US20070227715A1

US20070227715A1 - Heat exchanger

Info

Publication number: US20070227715A1
Application number: US11/732,097
Authority: US
Inventors: Masahiro Shimoya; Akira Itoh
Original assignee: Denso Corp
Current assignee: Denso Corp
Priority date: 2006-04-04
Filing date: 2007-04-02
Publication date: 2007-10-04
Also published as: JP2007278558A; DE102007015530A1; CN101050925A

Abstract

A heat exchanger has tubes defining refrigerant passages therein and fins disposed between the tubes. The tubes have tube main walls opposed to each other. The fins are joined to the tube main walls. The tube main walls have projections that project inside of the tubes and define recesses on outer sides of the tubes. Each of the tubes has an outer dimension, in a direction perpendicular to the tube main walls, in a range between equal to or greater than 0.8 mm and equal to or less than 1.9 mm.

Description

CROSS REFERENCE TO RELATED APPLICATION

This application is based on Japanese Patent Application No. 2006-103093 filed on Apr. 4, 2006, the disclosure of which is incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates to a heat exchanger, which is for example used as a refrigerant radiator for cooling a refrigerant flowing in tubes.

BACKGROUND OF THE INVENTION

U.S. Pat. No. 6,595,273 B2 (JP-A-2004-3787) discloses a heat exchanger having flat tubes, as a refrigerant radiator. The flat tubes have recesses on flat walls (tube main walls) thereof for allowing air to flow, thereby to improve the efficiency of heat exchange. The recesses are defined by projections formed on the tube main walls.
The projections have serpentine side walls such that air flows in the recesses in a serpentine or meandering manner. Because the flow of air adjacent to outer surfaces of the tube main walls is disturbed, development of a temperature boundary layer adjacent to the outer surfaces of the tube main walls is reduced. Thus, a coefficient of heat transfer of the air improves.
The recesses are formed by pressing the outer surfaces of the tube main walls in an inward direction of the tubes. Therefore, inside of the flat tubes, the refrigerant flows over the inward projections in a serpentine manner, and hence the flow of refrigerant is disturbed. Because development of a temperature boundary layer adjacent to inner surfaces of the tube main walls is reduced, a coefficient of heat transfer of the refrigerant improves. Accordingly, in this heat exchanger, the coefficient of heat transfer of both of the air and the refrigerant is improved by disturbing the flows of air and refrigerant, thereby improving the efficiency of heat exchange.
However, since the refrigerant flows in the serpentine manner, resistance to flow of the refrigerant increases, resulting in pressure loss of the refrigerant. If the temperature of the refrigerant reduces due to the pressure loss of the refrigerant, the temperature difference between the refrigerant and the air reduces. Further, this will affect the efficiency of heat exchange.

SUMMARY OF THE INVENTION

The present invention is made in view of the foregoing matter, and it is an object of the present invention to provide a heat exchanger having tubes with projections for disturbing flows of an internal fluid and an external fluid, which is capable of sufficiently maintaining or improving efficiency of heat exchange.
According to an aspect of the present invention, a heat exchanger has tubes and fins. The tubes define passages therein through which a refrigerant as an internal fluid flows. The tubes have tube main walls opposed to each other. The fins are disposed between the tubes and joined to the tube main walls. The tube main walls have projections projecting inside of the tubes and defining recesses outside of the tubes for allowing an external fluid to flow. Each of the tubes has a tube outer dimension (tube height), in a direction perpendicular to the tube main walls, in a range between equal to or greater than 0.8 mm and equal to or less than 1.9 mm.
As the tube outer dimension increases, a passage area of a refrigerant passage increases and resistance to flow of the refrigerant reduces. Therefore, pressure loss of the refrigerant is reduced and hence the decrease of efficiency of heat exchange may be suppressed. However, if the tube outer dimension is increased more than necessarily, the resistance to flow of the refrigerant is reduced excessively. In this case, although the refrigerant flows smoothly, disturbing effect of the refrigerant is reduced.
Accordingly, the tube outer dimension is set in a range between equal to or greater than 0.8 mm and equal to or less than 1.9 mm. When the tube outer dimension is in this range, the pressure loss of the refrigerant is reduced while the disturbing effect of the refrigerant is maintained. Therefore, the efficiency of heat exchange is sufficiently provided.

BRIEF DESCRIPTION OF THE DRAWINGS

Other objects, features and advantages of the present invention will become more apparent from the following detailed description made with reference to the accompanying drawings, in which like parts are designated by like reference numbers and in which:

FIG. 1 is a perspective view of a heat exchanger according to a first embodiment of the present invention;

FIG. 2 is a schematic perspective view of a part of the heat exchanger according to the first embodiment;

FIG. 3A is a perspective view for showing a step of forming projections and recesses on a plate member for the heat exchanger according to the first embodiment;

FIG. 3B is a perspective view for showing a step of folding the plate member according to the first embodiment;

FIG. 4 is a perspective view for showing a step of joining formed tube members as another example of forming a tube for the heat exchanger according to the first embodiment;

FIG. 5 is a graph showing a relationship between the height of the tube and efficiency of heat transfer of air according to the first embodiment;

FIG. 6 is a graph showing a relationship between pitch of projections of the tubes and efficiency of heat transfer of air according to the first embodiment;

FIG. 7 is a graph showing a relationship between the height of fins of the heat exchanger and efficiency of heat transfer of air according to the first embodiment;

FIG. 8 is a schematic perspective view of a part of a heat exchanger according to a second embodiment of the present invention;

FIG. 9 is a schematic perspective view of a part of a heat exchanger according to a third embodiment of the present invention;

FIG. 10 is a schematic perspective view of a part of a heat exchanger according to a fourth embodiment of the present invention;

FIG. 11 is a schematic perspective view of a part of a heat exchanger according to a fifth embodiment of the present invention;

FIG. 12 is a schematic perspective view of a part of a heat exchanger according to a sixth embodiment of the present invention;

FIG. 13 is a schematic perspective view of a part of a heat exchanger according to a seventh embodiment of the present invention;

FIG. 14 is a schematic perspective view of a part of a heat exchanger according to an eighth embodiment of the present invention;

FIG. 15 is a perspective view for showing a step of forming a tube of the heat exchanger according to the eighth embodiment; and

FIG. 16 is a schematic perspective view of a part of a heat exchanger according to a ninth embodiment of the present invention.

DETAILED DESCRIPTION OF EXAMPLE EMBODIMENT

First Embodiment

A first embodiment will be described with reference to FIGS. 1 to 7. As shown in FIG. 1, a heat exchanger 10 is for example used as a refrigerant condenser of a refrigerating cycle for a vehicle air conditioner, and is mounted in an engine compartment of a vehicle at a position where outside air is sufficiently supplied when the vehicle is running.
The heat exchanger 10 has a generally rectangular outline and includes a heat exchanging part 13 and tanks 14, 15. The heat exchanging part 13 includes flat tubes and fins 12. The flat tubes 11 define refrigerant passages therein through which a refrigerant flows. The fins 12 are for example corrugated fins. The heat exchanging part 13 performs heat exchange between the refrigerant and air flowing outside of the flat tubes 11.
The tanks 14, 15 are coupled to opposite longitudinal ends of the tubes 11. The refrigerant is distributed into the flat tubes 11 from one of the tanks 14, 15 (e.g., a left tank in FIG. 1) and is collected in the other one of the tanks 14, 15 (e.g., a right tank in FIG. 1). At the ends of the heat exchanging part 13, i.e., at longitudinal ends of the tanks 14, 15, side plates 16, 17 are provided as members for maintaining the outline of the heat exchanger 10. The side plates 16, 17 are arranged parallel to the tubes 11 and ends of the side plates 16, 17 are connected to the tanks 14, 15. The tubes 11, the fins 12, and the tanks 14, 15 are integrally joined by brazing, for example.
The tanks 14, 15 are made of metal such as aluminum alloy, and in the form of cylindrical container. The tanks 14, 15 are formed with slits (not shown), and the slits are arranged at predetermined intervals in a longitudinal direction of the tanks 14, 15. The longitudinal ends of the tubes 11 are inserted in the slits to make communication with the tanks 14, 15.
The tanks 14, 15 are provided with connecting blocks 14 a, 15 a. For example, a first connecting block 14 a is brazed to the left tank 14 at a position adjacent to one end (e.g., lower end in FIG. 1). An inlet pipe (not shown) is coupled to the connecting block 14 a for introducing a high temperature, high pressure refrigerant, which has been discharged from a compressor (not shown) of the refrigerating cycle, into the left tank 14.
Also, a second connecting block 15 a is brazed with the right tank 15 at a position adjacent to an opposite end (e.g., upper end in FIG. 1). An outlet pipe (not shown) is coupled to the connecting block 15 a for discharging a liquid-phase refrigerant, which has passed through the heat exchanger 10, toward an expansion valve (not shown) of the refrigerating cycle.
Further, the tanks 14, 15 are provided with engaging projections 14 b, 15 b at ends thereof (lower ends in FIG. 1) for fixing the heat exchanger 10 to a body of the vehicle.
Next, a structure of the heat exchanging part 13 will be described with reference to FIG. 2.
Each of the tubes 11 has generally flat walls 20, 21 (hereafter, referred to as tube main walls) opposed to each other. The tube main walls 20, 21 extend substantially parallel to a general flow direction Ar1 of the air. The tubes 11 and the fins 12 are stacked in a direction perpendicular to the tube main walls 20, 21, thereby to construct the heat exchanging part 13. The tubes 11 are joined to the fins 12 through the tube main walls 20, 21.
Here, each tube 11 has an outer dimension (hereafter, referred to as a tube height) H in a direction perpendicular to the tube main walls 20, 21. The tube height H is in a range between equal to or greater than 0.8 mm and equal to or less than 1.9 mm. Also, each fin 12 has a height (hereafter, referred to as a fin height) F in the direction perpendicular to the tube main walls 20, 21. The fin height F is in a range between equal to or greater than 2.0 mm and equal to or less than 9.0 mm.
Each of the tube main walls 20 has projections 22 projecting inside of the tube 11. For example, the projections 22 are formed by pressing or embossing corresponding portions of the tube main wall 20 from the outer side to the inner side. Thus, the projections 22 define recesses 20 a on a tube outer side for allowing air to pass through as shown by an arrow Ar2. Similarly, the tube main walls 21 have projections 23 that define recesses 21 a on the tube outer side.
Specifically, each of the projections 22, 23 ( recesses 20 a, 21 a) extends in a serpentine or meandering manner with a constant width along the tube main wall 20, 21. Also, the projection 22, 23 extends entirely from an air upstream end to an air downstream end of the tube 11 with respect to the air general flow direction Ar1. In this embodiment, the projections 22 of the tube main wall 20 and the projections 23 of the tube main wall have the same shape.
Each projection 22, 23 has an end wall, which corresponds to a bottom of each recess 20 a, 21 a, and side walls. The end wall defines a substantially flat wall. The side walls connect to the outer surfaces of the tube main wall, but form rounded corners with the outer surface of the tube main wall. Namely, corners 22 a, 23 a between the side walls of the projection 22, 23 and the outer surfaces of the tube main wall are chamfered in an arc shape.
The projections 22, 23 are arranged at predetermined intervals (pitch) P with respect to a general flow direction Rf1 of the refrigerant, i.e., a longitudinal direction of the tube 11. Here, the pitch P is within a range between equal to or greater than 1.0 mm and equal to or less than 6.5 mm.
The end walls of the projections 22, 23 have first depressed portions 22 b, 23 b at positions corresponding to apexes or most curved portions of the meandering recesses 20 a, 21 a. The first depressed portions 22 b, 23 b are further recessed inside of the tube 11 from the end walls of the projections 22, 23 in stepwise.
The projections 22 of the tube main wall 20 and the projections 23 of the tube main wall 21 are staggered with respect to the refrigerant general flow direction Rf1. Further, the projection 22 of the tube main wall 20 and the projection 23 of the tube main wall 21 overlap with each other at the first depressed portions 22 b, 23 b. Also, the tube main walls 20, 21 are in contact with and joined to each other at the first depressed portions 22 b, 23 b.
Also, the end wall of the projections 22, 23 have second depressed portions 22 c, 23 c at positions corresponding to upstream and downstream ends of the projections 22, 23 with respect to the air general flow direction Ar1. The second depressed portions 22 c, 23 c are further recessed from the end walls of the projections 22, 23 inside of the tube 11, similar to the first depressed portions 22 b, 23 b. Thus, the second depressed portions 22 c, 23 c are in contact with and joined to each other. In this embodiment, the dimension of the first and second depressed portions 22 b, 23 b, 22 c, 23 c from the outer surfaces of the tube main walls 20, 21 in the direction perpendicular to the tube main walls 20, 21 is 0.65 mm, for example.
FIGS. 3A, 3B show an example of a method of forming the tube 11. As shown in FIG. 3A, first, the projections and recesses as described above are formed on a metal plate, which is for example made of aluminum alloy, by roll forming using rollers 24, 25. Then, the formed metal plate is folded relative to its centerline, as shown by an arrow B in FIG. 3B, and joined. In this case, therefore, the tube main walls 20, 21 of the tube 11 are formed from a single metal plate member. Instead of the roll forming shown in FIG. 3A, the metal plate can be shaped by pressing.
FIG. 4 shows an another example of the method of forming the tube 11. The tube 11 can be formed by joining two metal plates. Specifically, the projections and recesses are formed on a first tube member 11 a and a second tube member 11 b, separately. Thereafter, the first member 11 a and the second member 11 b are arranged opposite to each other and joined to each other. Thus, the tube main walls 20, 21 are provided by the first and second tube members 11 a, 11 b.
In this embodiment, the projections 22, 23 of the tube main walls 20, 21 have the same shape. Thus, the first tube member 11 a and the second tube member 11 b are shaped into the same shape. That is, the first tube member and the second tube member 11 a, 11 b are provided by the same members. Therefore, productivity of the tubes 11 improves, and hence manufacturing costs of the tubes 11 reduces.
As shown in FIG. 2, complex refrigerant passages are provided inside of each tube 11. Since the meandering projections 22, 23 are formed on the tube main walls to project inside of the tube 11, refrigerant passages are formed in a serpentine or meandering manner with respect to the direction perpendicular to the tube main walls 20, 21, as shown by arrows Rf2.
Specifically, since inner surfaces of the first depressed portions 22 b, 23 b of the projections 22, 23 are in contact with each other, the refrigerant passages are divided over the first depressed portions 22 b, 23 b and then merged together. The flows of the refrigerant repeat the divergence and mergence while flowing along the tube main walls 20, 21 in the meandering manner.
The fins 12 are for example corrugated fins. Each of the fins 12 is formed by bending a thin plate member, which is for example made of aluminum alloy, into a corrugated shape. The fin 12 has joining walls (first walls) 12 a, 12 b to be joined to the outer surfaces of the tube main walls 20, 21.
Also, the fin 12 has connecting walls (second walls) 12 c, 12 d extending perpendicular to the joining walls 12 a, 12 b. The connecting walls 12 c, 12 d are formed with louvers 12 e, 12 f. The louvers 12 e, 12 f are formed by cutting out from the flat walls 12 c, 12 d and bending relative to the connecting walls 12 c, 12 d so as to oppose a flow of air passing through the connecting walls 12 c, 12 d (arrow Ar3).
Next, an operation of the heat exchanger 10 will be briefly described. The high temperature, high pressure refrigerant, which has been discharged from the compressor (not shown), flows into the left tank 14 of heat exchanger 10 through the first connector block 14 a. The refrigerant is distributed into the tubes 11 from the left tank 14.
While the refrigerant flowing in the tubes 11, heat of the refrigerant is transferred to the air flowing outside of the tubes 11 through entire surfaces of the tubes 11 and the fins 12. Thus, the refrigerant is condensed into the liquid-phase. The liquid-phase refrigerant is collected in the right tank 15 and discharged from the heat exchanger 10 through the second connecting block 15 a. Then, the refrigerant is introduced into the expansion valve (not shown), for example.
Next, an effect of heat exchange between the refrigerant and the air in the heat exchanging part 13 will be described. As shown by the arrows Rf2 in FIG. 2, since the refrigerant flows in the tubes 11 while meandering complexly, the flow of the refrigerant is disturbed. As such, the coefficient of heat transfer from the refrigerant improves. Accordingly, efficiency of heat transfer improves.
On the other hand, the air that flows at positions separated from the tube main walls 20, 21 flows along the fins 12, as shown by the arrow Ar3 in FIG. 2. This air receives heat from the fin 12 and then flows out of the fins 12. Thus, the fins 12 are cooled by the air passing through the fins 12.
Also, the air flowing adjacent to the tube main walls receives heat from the tubes 11 and is discharged from the heat exchanging part 13 after cooling the tube 11. In this case, as the air flows through the recesses 20 a, 21 a in the meandering manner, as shown by the arrows Ar2, the flow of this air is disturbed. As such, the coefficient of heat transfer of the air improves.
In addition, as the air is contracted when flowing into the recesses 20 a, 21 a, the coefficient of heat transfer of the air improves. Further, because the surface of heat transfer is increased by the recesses 20 a, 21 a, the amount of heat radiation from the tube 11 to the air is increased.
FIG. 5 shows a graph showing a relationship between the tube height H and efficiency Q of heat transfer. The efficiency Q is represented by the following equation (1):
Q=φ·Cp·ρ·Wa(Tr−Ta) (1)
Here, φ represents temperature efficiency of the heat exchanger 10; Cp represents specific heat of air; ρ represents density of air; Wa represents the volume of air; Tr represents temperature of refrigerant; and Ta represents inlet temperature of air.
In FIG. 5, a solid line L1 shows a measured result of the heat exchanger 10 of this embodiment in which the projection pitch P is 3.6 mm and the fin height F is 5.0 mm. In FIG. 5, a vertical axis represents the efficiency Q. When the tube height H is 0.8 mm and 1.9 mm, the efficiency Q is set to 100%.
As shown by the solid line L1 in FIG. 5, when the tube height H is 1.3 mm, the efficiency Q is at the maximum level. Namely, as the tube height H reduces smaller than 1.3 mm, a passage area inside of the tube 11 reduces. Thus, the flow speed of the refrigerant increases. Because pressure loss of the refrigerant increases, the pressure of the refrigerant reduces. As a result, the refrigerant temperature Tr reduces, and hence the difference between the refrigerant temperature Tr and the air inlet temperature Ta reduces. Accordingly, the efficiency Q expressed by the equation (1) reduces.
On the other hand, as the tube height H increases larger than 1.3 mm, the passage area inside of the tube 11 increases. Although the refrigerant flows smoothly, the disturbing effect of the refrigerant reduces. Thus, the coefficient of heat transfer of the refrigerant reduces. With the decrease of the coefficient of heat transfer of the refrigerant, the temperature efficiency φ of the heat exchanger 10 reduces. Accordingly, the efficiency Q reduces.
In the case that the tube height H is in a range between equal to or greater than 0.8 mm and equal to or less than 1.9 mm, the pressure loss of the refrigerant is reduced while maintaining the disturbing effect of the refrigerant. As such, the decrease of the efficiency Q due to the pressure loss of the refrigerant is suppressed. That is, when the tube height H is in the above range, the efficiency Q is provided sufficiently.
Also, in the case that the tube height H is in a range between equal to or greater than 1.0 mm and equal to or less than 1.6 mm, the decrease of the efficiency Q due to the pressure loss of the refrigerant is further suppressed.
Further, in the case that the tube height H is in a range between equal to or greater than 1.2 mm and equal to or less than 1.4 mm, the decrease of the efficiency Q due to the pressure loss of the refrigerant is further effectively suppressed.
In FIG. 5, a dashed line L2 shows a measured result of the heat exchanger 10 without having the projections 22, 23, as a comparative example. In the comparative example without having the projections 22, 23, the efficiency Q is at the maximum level when the tube height H is 1.0 mm. With the decrease or increase of the tube height H relative to 1.0 mm, the efficiency Q reduces by the same reason described in the above.
Accordingly, the efficiency Q of this embodiment having the projections 22, 23 improves, as compared with the comparative example without having the projections 22, 23. Also, in this embodiment, the tube height H where the efficiency Q is at the maximum level is greater than that of the comparative example. Namely, in the comparative example, the efficiency Q is at the maximum level when the tube height H is 1.0 mm. On the other hand, in this embodiment, the efficiency Q is at the maximum level when the tube height H is 1.3 mm.
In this embodiment, the flow of the refrigerant is disturbed by the projections 22, 23. Therefore, the pressure loss of the refrigerant of this embodiment is larger than the pressure loss of the refrigerant of the comparative example even when the tube height H is the same between this embodiment and the comparative example.
In FIG. 5, the projection pitch P is 3.6 mm and the fin height F is 5.0 mm, for example. When the projection pitch P and the fin height F are varied from these values, the relationship between the tube height H and the efficiency Q have the similar trend as FIG. 5, though the efficiency Q entirely, slightly reduces. That is, even when the projection pitch P and the fin height F are varied, the efficiency Q is at the maximum level when the tube height H is approximately 1.3 mm. Also, the efficiency Q reduces when the tube height H is reduced or increased relative to approximately 1.3 mm.
A graph of FIG. 6 shows a relationship between the projection pitch P and the efficiency Q of this embodiment in which the tube height H is 1.3 mm and the fin height F is 5.0 mm. In FIG. 6, a vertical axis represents the efficiency Q. When the projection pitch P is 1.0 mm and 6.5 mm, the efficiency Q is set to 100%.
As shown in FIG. 6, when the projection pitch P is 3.6 mm, the efficiency Q of the heat exchanger 10 is at the maximum level. As the projection pitch P reduces from 3.6 mm, the number of the projections 22, 23 of the tube 11 increases. Therefore, the disturbing effect of the flow of the refrigerant increases, and hence the pressure loss of the refrigerant increases. As a result, the pressure of the refrigerant reduces, and the temperature difference between the refrigerant temperature Tr and the air inlet temperature Ta reduces. Accordingly, the efficiency Q reduces.
On the other hand, as the projection pitch P increases from 3.6 mm, the number of the projections 22, 23 of the tubes 10 reduces. Therefore, the disturbing effect of the flow of the refrigerant reduces, and hence the flow of the refrigerant becomes close to a natural convection current. As a result, the coefficient of heat transfer of the refrigerant reduces. Thus, the temperature efficiency φ reduces, and hence the efficiency Q reduces.
Accordingly, when the projection pitch P is in a range between equal to or greater than 1.0 mm and equal to or less than 6.5 mm, the efficiency Q is improved by effectively providing the disturbing effect of the refrigerant.
Also, when the projection pitch P is in a range between equal to or greater than 1.6 mm and equal to or less than 5.7 mm, the efficiency Q is improved by providing the disturbing effect of the refrigerant further effectively.
Further, when the projection pitch P is in a range between equal to or greater than 2.3 mm and equal to or less than 5.0 mm, the disturbing effect of the refrigerant is further effectively improved. Thus, the efficiency Q is improved.
In FIG. 6, the tube height H is 1.3 mm and the fin height F is 5.0 mm, for example. When the tube height H and the fin height F are varied relative to these values, the relationship between the tube height H and the efficiency Q has the similar trend as FIG. 6, though the efficiency Q entirely, slightly reduces. That is, even when the tube height H and the fin height F are varied, the efficiency Q is at the maximum level when the projection pitch P is approximately 3.6 mm. Also, as the projection pitch P is reduced or increased from approximately 3.6 mm, the efficiency Q reduces.
In this embodiment, the end walls of the projections 22, 23 have the first and second depressed portions 22 b, 23 b, 22 c, 23 c that are further projected inside of the tube 11 in a stepwise manner. Therefore, the flows of air in the recesses 20 a, 21 a are further disturbed. As such, the coefficient of heat transfer of the air is further improved.
A graph of FIG. 7 shows a relationship between the fin height F and the efficiency Q of the heat exchanger 10 of this embodiment. Here, the tube height H and the projection pitch P are set to optimum values. Specifically, the tube height H is 1.3 mm and the projection pitch P is 3.6 mm. In FIG. 7, a vertical axis represents the efficiency Q. When the fin height F is 2 mm and 9 mm, the efficiency Q is set to 100%.
As shown in FIG. 7, in this embodiment, the efficiency Q is at the maximum level when the fin height F is 5.0 mm. As the fin height F reduces smaller than 5.0 mm, the heat transfer area reduces. As a result, the temperature efficiency φ reduces. Therefore, the efficiency Q reduces.
On the other hand, as the fin height F increases greater than 5.0 mm, the heat transfer area excessively increases. As a result, because the fin efficiency reduces, the temperature efficiency φ reduces. Therefore, the efficiency Q reduces.
Accordingly, when the fin height F is in a range between equal to or greater than 2.0 mm and equal to or less than 9.0 mm, heat radiation is effectively performed by the fin 12. As such, the efficiency Q improves.
Also, when the fin height F is in a range between equal to or greater than 3.0 mm and equal to or less than 7.3 mm, the heat radiation is further effectively performed by the fin 12. As such, the efficiency Q improves.
Further, when the fin height F is in a range between equal to or greater than 4.0 mm and equal to or less than 6.0 mm, the heat radiation is much further effectively performed by the fin 12. As such, the efficiency Q further improves.
In FIG. 7, the tube height is 1.3 mm and the projection pitch P is 3.6 mm. Even when the tube height H and the projection pitch P are varied from these values, the relationship between the tube height H and the efficiency Q have the similar trend as FIG. 7, though the efficiency Q entirely, slightly reduces. That is, even when the tube height H and the projection pitch P are varied, the efficiency Q is at the maximum level when the fin height F is approximately 5.0 mm. Also, as the fin height F is reduced or increased relative to approximately 5.0 mm, the efficiency Q reduces.
Since the projections 20, 21 extend continuously from the air upstream end to the air downstream end of the tube 11, the flow of air is introduced by the projections 20, 21 ( recesses 20 a, 21 a). Thus, the disturbing effect of the air improves, and the efficiency of heat transfer improves. Further, the projections 20, 21 extend in the serpentine or curved manner, the disturbing effect of the air improves.

Second Embodiment

A second embodiment will be described with reference to FIG. 8. As shown in FIG. 8, the fins 12 do not have the louvers 12 e, 12 f on the flat walls 12 c, 12 d. Structures of the heat exchanger 10 other than the louvers 12 e, 12 f are similar to those of the heat exchanger 10 of the first embodiment.
In the heat exchanger 10 of the second embodiment, because the efficiency of heat radiation of the fins 12 slightly reduces due to elimination of the louvers 12 e, 12 f, the efficiency Q slightly reduces as compared with the efficiency Q of the first embodiment. However, the tubes. 11 have the similar structure as the first embodiment. Therefore, effects substantially similar to the effects of the first embodiment are provided in the second embodiment.
Accordingly, the relationships between the efficiency Q and the tube height H, projection pitch P and fin height F have the similar trends as those of the first embodiment, though the efficiency Q is slightly reduced in this embodiment.

Third Embodiment

A third embodiment will be described with reference to FIG. 9. In the third embodiment, the heat exchanger 10 has the similar structure as that of the second embodiment other than corners 22 a, 23 a of the projections 22, 23.
In the second embodiment, the corners 22 a, 23 a are chamfered into the arc shape. In the third embodiment, on the other hand, the corners 22 a, 23 a are not chamfered. Instead, the side walls of the projections 22, 23 and the outer surfaces of the tube main walls 20, 21 form edged corners, as shown in FIG. 9.
Also in this case, the effects similar to the second embodiment are provided. Therefore, the relationships between the efficiency Q and the tube height H, projection pitch P and fin height F have the similar trends as those of the second embodiment.

Fourth Embodiment

A fourth embodiment will be described with reference to FIG. 10. In the third embodiment, the projections 22, 23 have the serpentine shape for allowing the air to flow in the serpentine manner. On the other hand, in the fourth embodiment, the projections 22, 23 have straight shape.
As shown in FIG. 10, the projections 22, 23 extend straight in a direction oblique to the general air flow direction along the tube main walls 20, 21 with a constant width. Also in this embodiment, the projections 22 of the tube main wall 20 and the projections 23 of the tube main wall 21 have the same shape. However, the projections 22, 23 are not parallel to each other. The projections 22, 23 are arranged to intersect with each other at a predetermined angle. With this arrangement, the refrigerant passages are formed inside of the tubes 11 in complexly serpentine manner.
On the other hand, the air flowing adjacent to the tubes 11 are biased and disturbed by the recesses 20 a, 21 a. Therefore, the coefficient of heat transfer of the air improves.
Accordingly, even when the projections 22, 23 are formed to extend straight in the oblique direction relative to the general air flow direction Ar1, the effects similar to the third embodiment are provided. Therefore, the relationships between the efficiency Q and the tube height H, projection pitch P and fin height F have the similar trends as those of the third embodiment.

Fifth Embodiment

A fifth embodiment will be described with reference to FIG. 11. In the fifth embodiment, the projections 22, 23 are formed into V-shape that extend in the refrigerant general flow direction Rf1 along the tube main walls 20, 21 with a constant width, as shown in FIG. 11.
Also in this case, the projections 22 of the tube main wall 20 and the projections 23 of the tube main wall 21 have the same V-shape, but are arranged in opposite directions with respect to the refrigerant general flow direction Rf1. The diverged ends of the V-shaped projection 22 are opposed to the diverged ends of the V-shaped projection 23. Therefore, the projections 22, 23 intersect with each other at the predetermined angle. Accordingly, the refrigerant passages are formed inside of the tubes 11 in complexly serpentine manner.
On the other hand, the air flowing adjacent to the tube main walls 20, 21 are biased and disturbed by the recesses 20 a, 21 a. Therefore, the coefficient of heat transfer of the air improves.
Accordingly, even when the projections 22, 23 are formed in the V-shape diverging in the refrigerant general flow direction Rf1, the effects similar to the third embodiment will be provided. Further, the relationships between the efficiency Q and the tube height H, projection pitch P and fin height F have the similar trends as those of the third embodiment.

Sixth Embodiment

A sixth embodiment will be described with reference to FIG. 12. In the third embodiment, the projections 22, 23 entirely extend from the upstream end to the downstream end of the tube 11 with respect to the air general flow direction Ar1. In the sixth embodiment, on the other hand, the projections 22, 23 are ended between the upstream end and the downstream end of the tube 11 with respect to the air general flow direction Ar1.
In the example of FIG. 12, the projections 22, 23 are ended at positions upstream of the downstream end of the tube 11 with respect to the air general flow direction Ar1. That is, the projections 22, 23 do not extend to the downstream end of the tube 11. Alternatively, the projections 22, 23 can be ended adjacent to the upstream end of the tube 11 or at middle positions of the tube 11 with respect to the air general flow direction Ar1. Also, it is not always necessary that all of the projections 22, 23 are ended at the same position with respect to the air general flow direction Ar1. The projections 22, 23 may be ended at different positions with respect to the air general flow direction Ar1, appropriately.
Also in this embodiment, the projections 22, 23 provide the disturbing effect of the flows of the refrigerant and air, similar to the third embodiment. Accordingly, the similar effects as the third embodiment are provided. Also, the relationships between the efficiency Q and the tube height H, projection pitch P and fin height F have the similar trends as those of the third embodiment.

Seventh Embodiment

A seventh embodiment will be described with reference to FIG. 13. In the seventh embodiment, the projections 22, 23 are formed as mesh, as shown in FIG. 13.
Also in this embodiment, the refrigerant passages are formed inside of the tube 11 in the complexly serpentine manner. On the outside of the tube 11, the recesses 20 a, 21 a are formed as mesh. Thus, the air flows in the recesses 20 a, 21 a while diverging and merging repetitively, and is disturbed sufficiently. Therefore, the coefficient of heat transfer of the air improves.
Accordingly, even when the projections 22, 23 have the meshed shape, the similar effects as the third embodiment are provided. Also, the relationships between the tube height H, projection pitch P and fin height F and the efficiency Q have the similar trends as those of the third embodiment.

Eighth Embodiment

An eighth embodiment will be described with reference to FIGS. 14 and 15. In the above embodiments, the tube 11 is formed by folding the single plate member into two or joining two plate members after the projections and recessed are formed. In the eighth embodiment, on the other hand, the tube 11 is integrally formed without folding or joining.
As shown in FIG. 14, the tube 11 has separation walls 13 therein for separating the inner space of the tube 11 into plural spaces with respect to the air general flow direction Ar1. The separation walls 13 extend between inner surfaces of the tube main walls. Also, the separation walls 13 for example have flat plate shape and extend in the refrigerant general flow direction Rf1 (i.e., in the longitudinal direction of the tube 11). Thus, the refrigerant passages are aligned in the air general flow direction Ar1 inside of the tube 11.
In this embodiment, the projections 22, 23 have the serpentine shape, similar to the projections 22, 23 of the third embodiment. However, the end walls of the projections 22, 23 do not have the first and second depressed portions 22 b, 23 b, 22 c, 23 c.
FIG. 15 shows an example of a method of forming the tube 11 of this embodiment. First, a flat multi-passage tube 30 having the separation walls 31 therein is formed by extrusion using an extrusion die (not shown) that includes a female die and a male die. Then, the projections and recesses having the predetermined shape are formed on the flat multi-passage tube 30 by roll forming using a pair of rollers 32, 33. The separation walls 31 are maintained even after the projections and recesses are formed. Therefore, the inner space of the tube 11 is separated into the plural spaces with respect to the air general flow direction Ar1 by the separation wall 31. Here, the projections and recesses can be formed by pressing, instead of the roll forming.
In the tube 11 formed in the above-described manner, the effects similar to the third embodiment will be provided. Therefore, the relationships between the tube height H, projection pitch P and fin height F and the efficiency Q have the similar trends as those of the third embodiments. Also, since the separation walls 31 are integrally formed into the tube 11, strength of the tube 11 against pressure improves.

Ninth Embodiment

A ninth embodiment will be described with reference to FIG. 16. In the ninth embodiment, the tube 11 is integrally formed, similar to the eighth embodiment. However, the projections 22, 23 have the straight shape, similar to the fourth embodiment, as shown in FIG. 16.
Also in this case, the similar effects as the eighth embodiment will be provided. Thus, the relationships between the tube height H, projection pitch P and fin height F and the efficiency Q have the similar trends as those of the eighth embodiments.
Further, since the tube 11 has the separation walls 31 therein, the strength of the tube 11 against pressure improves, similar to the eighth embodiment. The shape of the projections 22, 23 can be changed into any other shapes as the above first to seventh embodiments.

Other Embodiments

In the above embodiments, the projections 22, 23 have the constant width. However, it is not always necessary that the projections 22, 23 have the constant width. The width of the projections 22, 23 can be changed or varied appropriately.
In the third to ninth embodiments, the fins 12 do not have the louvers 12 e, 12 f. However, the fins 12 may have the louvers 12 e, 12 f in the third to ninth embodiments. Also, the heat exchanger 10 may be implemented by any combinations of the above first to ninth embodiments.
In the above embodiments, the heat exchanger 10 is exemplary employed as the refrigerant condenser. However, the heat exchanger 10 can be employed as a refrigerant radiator of a supercritical refrigerating cycle in which the pressure of a refrigerant exceeds a critical pressure at a high pressure side. Further, use of the heat exchanger 10 may not be limited to the above.
The example embodiments of the present invention are described above. However, the present invention is not limited to the above example embodiment, but may be implemented in other ways without departing from the spirit of the invention.

Claims

1. A heat exchanger for performing heat exchange between a refrigerant and an external fluid, comprising:

tubes defining passages therein for allowing the refrigerant to flow, the tubes having tube main walls opposed to each other; and

fins disposed between the tubes and joined with the tube main walls, wherein

the tube main walls have projections that project inside of the tubes and define recesses outside of the tubes for allowing the external fluid to flow, and

each of the tubes has a tube outer dimension in a range between at least 0.8 mm and at most 1.9 mm, in a direction perpendicular to the tube main walls.

2. The heat exchanger according to claim 1, wherein

the tube outer dimension is in a range between at least 1.0 mm and at most 1.6 mm.

3. The heat exchanger according to claim 2, wherein

the tube outer dimension is in a range between at least 1.2 mm and at most 1.4 mm.

4. The heat exchanger according to claim 1, wherein

the projections are arranged at a predetermined pitch with respect to a longitudinal direction of the tube, and

the predetermined pitch is in a range between at least 1.0 mm and at most 6.5 mm.

5. The heat exchanger according to claim 4, wherein

the predetermined pitch is in a range between at least 1.6 mm and at most 5.7 mm.

6. The heat exchanger according to claim 5, wherein

the predetermined pitch is in a range between at least 2.3 mm and at most 5.0 mm.

7. The heat exchanger according to claim 1, wherein

each of the fins has a fin outer dimension in a range between at least 2.0 mm and at most 9.0 mm, in the direction perpendicular to the tube main walls.

8. The heat exchanger according to claim 7, wherein

the fin outer dimension is in a range between at least 3.0 mm and at most 7.3 mm.

9. The heat exchanger according to claim 8, wherein

the fin outer dimension is in a range between at least 4.0 mm and at most 6.0 mm.

10. The heat exchanger according to claim 1, wherein

the projections extend continuously from upstream ends to downstream ends of the tubes with respect to a flow direction of the external fluid flowing outside of the tubes.

11. The heat exchanger according to claim 1, wherein

each of the tubes has a first tube member and a second tube member joined with the first tube member, and the tube main walls of each tube are included in the first and second tube members.

12. The heat exchanger according to claim 11, wherein

the first tube member and the second tube member have an identical shape.

13. The heat exchanger according to claim 1 wherein

each of the tubes has a separation wall for separating an inner space of the tube into a plurality of spaces in a direction perpendicular to a longitudinal direction of the tube, and

the separation wall is integrated with the tube main walls.

14. The heat exchanger according to claim 1, wherein

the projections have curved shapes extending along the tube main walls in a meandering manner in a direction perpendicular to a longitudinal direction of the tubes.

15. The heat exchanger according to claim 1, wherein

the projections have straight shapes extending along the tube main walls and obliquely with respect to a longitudinal direction of the tubes.

16. The heat exchanger according to claim 1, wherein

the projections have V-shapes that extend along the tube main walls and diverging in a longitudinal direction of the tubes.

17. The heat exchanger according to claim 1, wherein

the projections are meshed along the tube main walls.

18. The heat exchanger according to claim 1, wherein

the projections have side walls extending perpendicular to outer surfaces of the tube main walls, and

the side walls define edged corners with the outer surfaces of the tube main walls.

19. The heat exchanger according to claim 1, wherein

the side walls define curved corners with the outer surfaces of the tube main walls.

20. The heat exchanger according to claim 1, wherein

the fins are corrugated fins including first fin walls and second fin walls connecting the first fin walls, and

the first fin walls are flat and joined with outer surfaces of the tube main walls.

21. The heat exchanger according to claim 20, wherein

the second fin walls have louvers that are angled with respect to a flow direction of the external fluid flowing through the fins.

22. The heat exchanger according to claim 1, wherein the external fluid is air for cooling the refrigerant.