JP2005083653A - Refrigerant evaporator - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a lightweight refrigerant evaporator having high cooling performance, by relatively freely setting a dimension of a fin by improving drainability of a condensate with a fin structure. <P>SOLUTION: This refrigerant evaporator comprises a core 4 by alternately juxtaposing a flat tube 2 and the corrugate fin 3. A first contact plane part 11 and a second contact plane part of the fin 3 are formed along an outside surface of the tube 2, and the whole first and second contact plane parts are joined to the outside surface of the tube 2. Intermediate surface parts 13 and 14 continuing with the first and second contact plane parts 11 are formed so as to substantially vertically extend to the outside surface of the tube 2. A dimension in the juxtaposing direction of the tube 2 of the fin 3, is set in a range of 3.0 mm to 6.0 mm. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a refrigerant evaporator for performing a refrigerant evaporation process in a refrigeration cycle such as an air conditioner.

Conventionally, for example, as a refrigerant evaporator of an automotive air conditioner, a flat tube having a long cross-sectional shape of the refrigerant flow path in the air flow direction, and a heat transfer fin formed in a corrugated plate shape when viewed in the air flow direction, The thing provided with the core formed by arranging in parallel alternately is known (for example, refer to patent documents 1). The fin of the refrigerant evaporator has a U-shaped corrugated top, and the top is brazed to the outer surface of the tube.
JP 2001-324290 A (3rd page, 4th page, FIG. 1, FIG. 2)

  However, in the refrigerant evaporator of Patent Document 1, since the fin brazed portion of the tube is formed in a U shape, the gap between the fin and the outer surface of the tube becomes narrower as it goes to the brazed portion. Thus, when a narrow gap is formed between the fin and the tube, condensed water accumulates in the gap, making it difficult to drain, increasing the ventilation resistance of the fin, reducing the heat transfer efficiency, and improving the cooling performance. descend. For this reason, in the refrigerant evaporator of patent document 1, when setting each dimension of a fin, the drainage property of condensed water must fully be considered, and it cannot set to the dimension which considered cooling performance as important.

  Further, in general, in the refrigerant evaporator, there is a demand to obtain high cooling performance while reducing the amount of material used and making it lightweight, but in the refrigerant evaporator of Patent Document 1, setting of fin dimensions is required. Since the degree of freedom is low, it is difficult to achieve both weight reduction and high cooling performance.

  The present invention has been made in view of such a point, and the object of the present invention is to improve the drainage of condensed water by devising the structure of the fins, so that the dimensions of the fins can be set relatively freely. It is possible to obtain a refrigerant evaporator that is lightweight and has high cooling performance.

  In order to achieve the above object, in the present invention, a cooling performance value of a plurality of refrigerant evaporators having different dimensions of fins and tubes is obtained by computer simulation, and a value obtained by dividing the cooling performance value by the weight of the refrigerant evaporator. That is, the fin size range in which the cooling performance value per unit weight of the refrigerant evaporator is good was specified.

  Specifically, a plurality of tubes whose cross-sectional shape of the refrigerant flow path is long in the air flow direction are arranged side by side in the direction intersecting the air flow direction, and fins for heat transfer are arranged between adjacent tubes. The refrigerant evaporator thus formed is a target.

  The fins extend substantially flat along the opposing outer surfaces of adjacent tubes, and are joined to the first and second joint surface portions joined in surface contact with the outer surfaces of the tubes, respectively. The intermediate surface portion extending substantially perpendicular to both the joint surface portions is formed into a corrugated plate shape when viewed in the air flow direction, and the dimension of the fins in the tube juxtaposition direction is set to 3.0 mm or more and 6.0 mm or less. The configuration.

  According to this configuration, the entire first and second joining surface portions of the fin are joined to the tube outer surface, and no gap is formed between the tube outer surface and each joining surface portion joined to the tube outer surface. Further, since the intermediate surface portion of the fin is substantially perpendicular to the first and second joining surface portions, it extends substantially perpendicular to the tube outer surface, and no gap is formed between the intermediate surface portion and the tube outer surface. By these things, it is prevented that condensed water accumulates between a fin and a tube, and condensed water comes to be drained smoothly. As a result, the ventilation resistance of the fins is reduced, the heat transfer efficiency is improved, and the cooling performance is improved. Thus, since the cooling performance can be ensured by the fin structure, the fin dimensions can be set relatively freely.

  The cooling performance value per unit weight of the refrigerant evaporator using the fins having the above-described structure is set in the tube juxtaposition direction of the tube in the refrigerant flow path of the tubes when the dimension of the fin juxtaposition direction of the fins is set in the above range. It increases regardless of the dimensions and dimensions of the core in the air flow direction. On the other hand, when the dimension of the fins in the tube juxtaposition direction is set outside the above range, the cooling performance value per unit weight of the refrigerant evaporator is lowered.

  That is, when the dimension of the fins in the direction in which the tubes are juxtaposed is shorter than 3 mm, the interval between the tubes is narrowed and the tubes are densely arranged. As a result, the cooling performance value is improved. Weight increases. Further, when the dimension of the fins in the direction in which the tubes are juxtaposed is longer than 6 mm, the number of tubes is reduced and the weight of the refrigerant evaporator is reduced, but the cooling performance value is further reduced.

  According to a second aspect of the present invention, in the first aspect of the present invention, the dimension in the tube juxtaposition direction in the refrigerant flow path of the tube is set to 0.7 mm or more and 1.2 mm or less.

  According to this configuration, the cooling performance value per unit weight of the refrigerant evaporator is increased regardless of the dimension of the core in the air flow direction.

  That is, if the tube juxtaposition direction of the tube in the refrigerant flow path of the tube is shorter than 0.7 mm, the cross-sectional area of the refrigerant flow path is reduced, the refrigerant flow rate of the entire refrigerant evaporator is reduced due to an increase in pressure loss in the pipe, and the cooling performance The value becomes lower. Moreover, if the tube juxtaposition direction dimension of the tube in the refrigerant flow path is longer than 1.2 mm, the cross-sectional area of the refrigerant flow path is increased, the refrigerant flow rate is lowered, the heat transfer coefficient inside the pipe is lowered, and the cooling performance value Becomes lower.

  According to a third aspect of the present invention, in the first or second aspect of the present invention, the dimension of the core constituted by the tubes and the fins in the air flow direction is set to 15 mm or more and 40 mm or less.

  According to this configuration, since the core is relatively thin, the efficiency of heat transfer on the air side is improved, and the cooling efficiency is increased.

  In the invention of claim 4, in the invention of claim 3, the dimension of the core in the air flow direction is set to 15 mm or more and less than 20 mm, and a path through which the refrigerant flows is constituted by a tube in the core, and the number of the paths is one. It is set as the structure which is thru | or three.

  According to this configuration, the flow rate of the refrigerant in the refrigerant evaporator is balanced with the pressure loss in the pipe, so that the heat transfer efficiency is improved and the cooling performance value is increased.

  In the invention of claim 5, in the invention of claim 3, the dimension of the core in the air flow direction is set to 20 mm or more and less than 30 mm, and a path through which the refrigerant flows is constituted by a tube in the core, and the number of the paths is 2 It is set as the structure which is thru | or four.

  According to this configuration, the cooling performance value increases as in the fourth aspect of the invention.

  In the invention of claim 6, in the invention of claim 3, the dimension of the core in the air flow direction is set to 30 mm or more and 40 mm or less, and a path through which the refrigerant flows is constituted by a tube in the core, and the number of the paths is 2 It is set as the structure which is thru | or 5.

  According to this configuration, the cooling performance value increases as in the fourth aspect of the invention.

  According to the first aspect of the present invention, the condensed water is smoothly drained by the structure of the fins, and the dimensions of the fins can be set relatively freely while ensuring the cooling performance. And the refrigerant evaporator is constituted using this fin, and the cooling performance value per unit weight of the refrigerant evaporator can be increased by setting the dimension of the fins in the juxtaposed direction to the range of 3 mm or more and 6 mm or less. A refrigerant evaporator that is lightweight and has high cooling performance can be obtained.

  According to the invention of claim 2, the cooling performance value can be increased by setting the dimension in the tube juxtaposition direction in the refrigerant flow path of the tube to 0.7 mm or more and 1.2 mm or less, and the unit weight of the refrigerant evaporator The per unit cooling performance value can be further increased.

  According to the invention of claim 3, by setting the dimension of the core in the air flow direction to 15 mm or more and 40 mm or less, the cooling efficiency of the refrigerant evaporator is increased, and the cooling performance value per unit weight of the refrigerant evaporator is further increased. Can be increased.

  According to the invention of claim 4, the cooling performance value per unit weight of the refrigerant evaporator is set by setting the dimension of the core in the air flow direction to 15 mm or more and less than 20 mm and setting the number of passes to 1 to 3. Can be increased.

  According to the fifth and sixth aspects of the invention, similarly to the fourth aspect of the invention, the cooling performance value per unit weight of the refrigerant evaporator can be further increased.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

  FIG. 2 shows a refrigerant evaporator 1 according to an embodiment of the present invention, and this refrigerant evaporator 1 constitutes one element of a refrigeration cycle in an automotive air conditioner.

  The refrigerant evaporator 1 is arranged at a core 4 in which a plurality of tubes 2 extending in the vertical direction and fins 3 for heat transfer are alternately arranged in the vehicle width direction, and at both upper and lower ends of the tube 2. An upper header tank 6 and a lower header tank 7 into which the upper end portion and the lower end portion of the tube 2 are inserted in a communicating state are provided. And although this refrigerant evaporator 1 is not shown in figure, it is accommodated in the casing of an air-conditioner, and it is comprised so that the air from an air blower may pass a core in the direction of arrow A.

  As shown in FIG. 1, the tube 2 of the core 4 is a flat tube in which the cross-sectional shape of the refrigerant flow path R is long in the air flow direction. The tube 2 is formed by bending a plate made of an aluminum alloy, and the edges of the plate are brazed. Adjacent tubes 2 are arranged such that the outer surfaces of the opposing tubes 2 are substantially parallel.

  On the other hand, the fin 3 is a corrugated fin formed in a corrugated plate shape when viewed in the air flow direction, and extends from the upper end portion of the tube 2 to the lower end portion. The fin 3 is formed by molding an aluminum alloy plate having a brazing material provided on both sides in layers. Out of these fins 3, the outer end fins 3 positioned at the left end and the right end of the core 4 in the vehicle width direction are respectively held by end plates 9 formed by forming a relatively thick aluminum alloy plate.

  As shown in FIG. 3, the fin 3 has a first joint surface portion 11 joined to the outer surface of the tube 2 adjacent to the left side in the vehicle width direction, and a second joint surface 12 joined to the outer surface of the tube 2 adjacent to the right side. Are alternately formed in the vertical direction. The joint surface portions 11 and 12 are continuous with the first intermediate surface portion 13 and the second intermediate surface portion 14, and the joint surface portions 11 and 12 and the intermediate surface portions 13 and 14 are integrally formed.

  The first joint surface portion 11 and the second joint surface portion 12 have the same shape, and both the joint surface portions 11 and 12 are formed along the outer surface of the tube 2 so as to extend to both ends in the width direction which is the air flow direction of the tube 2. The tube 2 is in surface contact with the outer surface. When the fin 3 is viewed from the vehicle width direction, the lower edge of the first joint surface portion 11 overlaps the upper edge of the second joint surface portion 12, and the first joint surface portion formed below the first joint surface portion 11. 11 is overlapped with the lower edge of the second joint surface portion 12.

  The first intermediate surface portion 13 extends from the lower edge of the first joint surface portion 11 to the upper edge of the second joint surface portion 12 substantially perpendicularly to the first joint surface portion 11. On the other hand, the second intermediate surface portion 14 extends from the upper edge of the first bonding surface portion 11 to the lower edge of the second bonding surface portion 12 substantially perpendicularly to the first bonding surface portion 11. The second intermediate surface portion 14 has substantially the same shape.

  The upstream end portion of the first intermediate surface portion 13 in the air flow direction extends straight in the air flow direction as shown in FIG. A first louver 13a that is cut and raised to the upper side, which is one side in the plate thickness direction of the first intermediate surface portion 13, on the downstream side in the air flow direction from the upstream end portion of the first intermediate surface portion 13, and the plate thickness direction Second louvers 13b cut and raised on the lower side, which is the other side, are alternately formed in the air flow direction.

  As shown in FIG. 3A, a portion corresponding to the first louver 13a of the first joint surface portion 11 is cut out on one side of the first intermediate surface portion 13 in the plate thickness direction, and the first louver 13a is cut from the cut portion. Extends to the second joint surface portion 12 side substantially perpendicular to the first joint surface portion 11. The second joint surface portion 12 side of the first louver 13a is bent to the other side in the plate thickness direction of the first intermediate surface portion 13 along the outer surface of the tube 2, and is continuous with the second joint surface portion 12. Accordingly, a substantially rectangular through hole 15 is formed between the first louver 13a and the end of the first intermediate surface portion 13 on the upstream side in the air flow direction so that air flows through the through hole 16. It has become. The width of the through hole 15 is substantially the same from the first joint surface portion 11 side to the second joint surface portion 12 side.

  The second louver 13b extends from the first joint surface portion 11 to the other side in the plate thickness direction of the first intermediate surface portion 13 along the outer surface of the tube 2, and then is substantially perpendicular to the first joint surface portion 11. It is bent toward the joint surface portion 12 and extends toward the second joint surface portion 12. The portion corresponding to the second louver 13b of the second joint surface portion 12 is cut out to the other side in the plate thickness direction of the first intermediate plate portion 13, and the second joint surface portion 12 side of the second louver 13b is connected to this notch portion. Yes. Accordingly, between the second louver 13b and the first louver 13a, the through-hole 16 having substantially the same width from the first joint surface portion 11 side to the second joint surface portion 12 side, like the through-hole 15. Is formed. The first louver 13a and the second louver 13b are entirely inclined with respect to the air flow direction, and as shown in FIG. 3B, the inclination angle θ is, for example, not less than 10 ° and not more than 15 °. Is set to By setting the inclination angle θ in this manner, it is possible to suppress the airflow resistance while improving the heat transfer efficiency by disturbing the flow of air passing through the core 4.

  Further, the second intermediate surface portion 14 is formed in the same manner as the first intermediate plate portion 13, and an end portion on the upstream side in the air flow direction extends straight in the air flow direction, and on the downstream side of the air flow, First louvers 14a and second louvers 14b are alternately formed to face the first louver 13a and the second louver 13b of the first intermediate surface portion 13, respectively. The first louver 14a and the second louver 14b are inclined in the same manner as the louvers 13a and 13b. A through-hole 15 is formed between the first louver 14a and the end of the second intermediate surface portion 14 on the upstream side in the air flow direction, and between the first louver 14a and the second louver 14b. A through hole 16 is formed.

  On the other hand, as shown in FIG. 2, the upper header tank 6 includes a tank body 20 formed by forming a plate made of aluminum alloy into a cylindrical shape. Openings at both ends of the tank main body 20 are respectively closed by cap members 21 made of aluminum alloy plates, and the tank main body 20 and the cap member 21 are brazed to form the upper header tank 6. It is configured. A tube insertion hole (not shown) into which the upper end of the tube 2 is inserted is formed in the wall portion on the tube 2 side in the tank body 20, and the periphery of the tube insertion hole and the outer peripheral surface of the upper end of the tube 2 Is to be brazed.

  An inflow cooler pipe 22 through which refrigerant flows into the upper header tank 6 is connected to the right side of the tank body 20 in the vehicle width direction. A first partition plate 23 for partitioning the inside of the upper header tank 6 into a space T1 on the left side in the vehicle width direction and a space T2 on the right side is disposed in a portion of the tank body 20 close to the inflow cooler pipe 22. Yes. The first partition plate 23 has a disk shape, and its peripheral edge is brazed to the inner peripheral surface of the tank body 20.

  The lower header tank 7 is also configured in the same manner as the upper header tank 6 and includes a tank body 24 and a cap member 25. An outflow cooler pipe 26 that allows the refrigerant in the lower header tank 7 to flow out of the refrigerant evaporator 1 is connected to the left side of the tank body 24 in the vehicle width direction. A second partition plate 27 for partitioning the inside of the lower header tank 7 into a space S1 on the left side in the vehicle width direction and a space S2 on the right side is disposed near the outflow cooler pipe 26 of the tank body 24. Similar to the first partition plate 23, the second partition plate 27 is brazed to the inner peripheral surface of the tank body 24.

  The upper end portion of the tube 2 located on the right side in the vehicle width direction of the first partition plate 23 of the core 4 communicates with the right space T2 of the upper header tank 6, and the lower end portion of these tubes 2 is the lower header tank 7. The tube 2 communicates with the right space S2, and the tube 2 forms a first path P1. Further, the upper end portion of the tube 2 positioned between the first partition plate 23 and the second partition plate 27 of the core 4 communicates with the left space T1 of the upper header tank 6, and the lower end portions of these tubes 2 are arranged on the lower side. The tube 2 communicates with the right space S2 of the header tank 7, and the tube 2 forms a second path P2. Furthermore, the upper end portion of the tube 2 located on the left side in the vehicle width direction with respect to the second partition plate 27 of the core 4 communicates with the left space T1 of the upper header tank 6, and the lower end portion of these tubes 2 is the lower header tank. 7 and the left side space S1 and the tube 2 forms a third path P3.

  In other words, the refrigerant evaporator 1 includes three paths P1, P2, and P3, and the refrigerant flowing from the inflow cooler pipe 22 into the right space T2 of the upper header tank 6 through the refrigerant expansion valve (not shown) While flowing in the right space T2 to the left in the vehicle width direction, it flows into each tube 2 constituting the first path P1, flows through the tube 2 downward, and flows into the right space S2 of the lower header tank 7. The refrigerant flowing into the right space S2 flows into the tubes 2 constituting the second path P2 while flowing to the left in the vehicle width direction, flows upward through the tubes 2, and flows into the left space T1 of the upper header tank 6. To do. The refrigerant flowing into the left space T1 flows into each tube 2 constituting the third path P3 while flowing to the left in the vehicle width direction, flows downward through the tube 2, and enters the left space S1 of the lower header tank 7. After flowing in, it flows out through the cooler pipe 26 for outflow.

  In this embodiment, the refrigerant evaporator 1 has three paths P1, P2, and P3. The number of these paths increases or decreases the partition plates 23 and 27 of the upper header tank 6 and the lower header tank 7. Can be changed. Specifically, although not shown, by removing the first partition plate 23 and the second partition plate 27, the number of passes becomes one, and the outflow cooler pipe is connected to the left space T1 of the upper header tank 6. By leaving the first partition plate 23 and removing the second partition plate 27, there are two paths. Further, it is possible to configure four or more paths by increasing the number of partition plates and partitioning the inside of the upper header tank 6 and the lower header tank 7 into three or more spaces.

  When the refrigerant flowing into the refrigerant evaporator 1 flows through the tubes 2 of the paths P1, P2, and P3, the refrigerant and the air from the blower passing through the core 4 exchange heat through the fins 3, and the air Is cooled. When this air is cooled, condensed water is generated on the surface of the fin 2. At this time, since the first joint surface portion 11 and the second joint surface portion 12 of the fin 2 are formed substantially flat along the outer surface of the tube 2, the entire first joint surface portion 11 and the second joint surface portion 12 correspond to each other. Joined to the outer surface of the tube 2. As a result, no gap is formed between the outer surface of the tube 2 and the joint surface portions 11 and 12 joined to the outer surface of the tube 2. Further, the first intermediate surface portion 13 and the second intermediate surface portion 14 of the fin 3 extend substantially perpendicular to the outer surface of the tube 2, and a gap is formed between the intermediate surface portions 13 and 14 and the outer surface of the tube 2. Not. By these things, condensed water does not accumulate between the fin 3 and the tube 2 outer surface, but is drained smoothly.

  Furthermore, since the through holes 15 and 16 formed in the fin 3 have substantially the same width from the first joint surface portion 11 side to the second joint surface portion 12 side, there are no locally narrow portions. Condensed water does not accumulate inside 15 and 16 and is drained smoothly.

  As described above, the structure of the fin 3 allows the condensate to be smoothly drained, reduces the airflow resistance, improves the heat transfer efficiency, and improves the cooling performance. You can do it freely. In addition, you may make it give a super hydrophilic surface treatment to this fin, and, thereby, the drainage property of condensed water improves more.

  Next, the cooling performance value by the refrigerant evaporator 1 configured as described above is obtained by computer simulation, and this cooling performance value is divided by the weight of the refrigerant evaporator 1 per unit weight of the refrigerant evaporator 1. The cooling performance value will be described with reference to FIGS. Here, the weight of the refrigerant evaporator 1 is the sum of the weights of the tubes 2, fins 3, end plates 9 and header tanks 6 and 7.

  In this computer simulation, the fin 3 height dimension Fh (see FIG. 3A), which is the dimension of the fin 3 in the tube 2 juxtaposition direction, and the tube 2 juxtaposition direction dimension in the refrigerant flow path R of the tube 2. The cooling performance value is obtained by changing the channel height dimension Thi (see FIG. 1) in the tube 2 and the core thickness D (see FIG. 2), which is the dimension of the core 4 in the air flow direction. The core thickness D corresponds to the dimension of the tube 2 in the width direction.

  The precondition for the computer simulation will be described. The distance between the lower end of the upper header tank 6 and the upper end of the lower header tank 7, which is the vertical dimension on the air passage surface of the core 4, is 213 mm. Further, the distance between both end plates 9, 9 which is the dimension in the vehicle width direction on the air passage surface of the core 4 is 221 mm. Further, the fin pitch Fp (see FIG. 1), that is, the distance between the first intermediate surface portion 13 and the second intermediate surface portion 14 of the fin 3 is the air passage resistance of the core 4 when the dimensions of the respective portions are changed. Is changed to be constant. Specifically, in this refrigerant evaporator 1, the fin 4 has a structure that allows the drainage of condensed water to be improved, so that the fin pitch Fp is, for example, 0.8 mm to 1.5 mm, preferably 1. It can be set to a relatively small value of about 2 mm. Thereby, the wave shape of the fin 4 becomes dense, a wide heat transfer area is secured, and the cooling performance value is improved. Further, regarding the plate thickness of the tube 2 and the fin 3, when these plate thicknesses are changed within the practical range, only the weight of the refrigerant evaporator 1 is changed, and the cooling performance value is hardly affected. Therefore, the plate thickness of the tube 2 is 0.2 mm so that the pressure resistance and corrosion resistance of the tube 2 can be sufficiently obtained, and the plate thickness of the fin 3 is 0 so that the deformation of the fin 3 hardly occurs in a normal use state. .07 mm. The plate thickness of the tube 2 is preferably 0.1 mm or more and 0.4 mm or less, and the plate thickness of the fin 3 is preferably 0.05 mm or more and 0.09 mm or less.

  Moreover, temperature, humidity, and air volume are made constant for the air blown to the core 4, and temperature and pressure in the inflow cooler pipe 22 are made constant for the refrigerant flowing into the refrigerant evaporator 1. At the same time, the temperature and pressure in the outflow cooler pipe 26 are kept constant.

  FIG. 4 shows the relationship between the cooling performance value per unit weight of the refrigerant evaporator 1 and the fin height dimension Fh when the flow path height dimension Thi in the tube is fixed to 0.5 mm. Simulation is performed for each of the core thicknesses D of 20 mm, 30 mm, 40 mm, and 50 mm. As a result, the cooling performance value per unit weight of the refrigerant evaporator is high in the range where the fin height dimension Fh is 3 mm or more and 6 mm or less regardless of the core thickness D, while the fin height dimension Fh is Outside this range, the cooling performance value per unit weight of the refrigerant evaporator is low.

  As described above, when the fin height dimension Fh is shorter than 3 mm, the cooling performance value per unit weight of the refrigerant evaporator decreases because the fin height dimension Fh is shortened and the interval between the tubes 2 is reduced. This is because the tube 2 is narrowed and densely arranged. That is, when the tubes 2 are densely arranged, the number of the tubes 2 included in the refrigerant evaporator 1 is increased, so that the internal surface area is increased and the pressure loss in the tubes is reduced to improve the cooling performance value. This is because an increase in weight due to an increase in the number of slags greatly affects.

  Further, the cooling performance value per unit weight of the refrigerant evaporator 1 decreases when the fin height dimension Fh is longer than 6 mm because the interval between the tubes 2 increases. That is, if the interval between the tubes 2 is increased, the number of the tubes 2 included in the refrigerant evaporator 1 is reduced, so that the weight of the refrigerant evaporator 1 is reduced. However, the cooling performance is further increased by reducing the inner surface area and increasing the pressure loss in the pipe. This is because the value decreases.

  5 to 8 show cases where the channel height dimension Thi in the tube is fixed to 0.7 mm, 1.0 mm, 1.2 mm, and 1.4 mm, respectively, and the thickness D of the core for each is shown. Are changed to the same four as in FIG. Even in these cases, the cooling performance value per unit weight of the refrigerant evaporator 1 is high when the fin height dimension Fh is in the range of 3 mm to 6 mm. On the other hand, if the height dimension Fh of the fin is out of the range, the cooling performance value per unit weight of the refrigerant evaporator 1 becomes low for the same reason as in the case of FIG.

  Further, when FIG. 4 showing the case where the flow path height dimension Thi in the tube is 0.5 mm is compared with FIG. 5 showing the case where it is 0.7 mm, the refrigerant evaporates regardless of the core thickness D. The cooling performance value per unit weight of the vessel 1 is generally higher when 0.7 mm. This is because when the flow path height dimension Thi in the tube is lower than 0.7 mm, the cross-sectional area of the refrigerant flow path R in the tube 2 is reduced and the flow velocity in the pipe is increased, thereby improving the heat transfer coefficient inside the pipe. However, the effect of increasing the pressure loss inside the tube is larger than that, and the cooling performance value is lowered.

  Also, comparing FIG. 7 showing the case where the flow path height dimension Thi in the tube is 1.2 mm and FIG. 8 showing the case where the flow path height dimension Thi in the tube is 1.4 mm, the core Regardless of the thickness D, the cooling performance value per unit weight of the refrigerant evaporator 1 is generally higher when the channel height dimension Thi in the tube is 1.2 mm. This is because when the flow path height dimension Thi in the tube is higher than 1.2 mm, the cross-sectional area of the refrigerant flow path R in the tube 2 is increased and the pressure loss in the pipe is reduced. This is because the effect of lowering the heat transfer coefficient inside the tube due to lowering is large, and the cooling performance value is lowered. Further, when the flow path height dimension Thi in the tube is set to 1.0 mm, the flow path height dimension Thi within the tube is set to 0.7 mm or the same as when 1.2 mm. The cooling performance value per unit weight of the refrigerant evaporator 1 increases.

  FIG. 9 shows the relationship between the cooling efficiency and the thickness D of the core. In the cooling efficiency in this figure, the fin height dimension Fh is changed between 2.5 mm and 7.0 mm, and the flow path height dimension Thi in the tube is set to 0 within the range of the fin height dimension Fh. It is an average value of values obtained by changing between 5 mm and 1.4 mm.

  As apparent from FIG. 9, the cooling efficiency increases as the core 4 becomes thinner. This is because the heat transfer coefficient on the air side improves as the core thickness D decreases. Cooling efficiency sufficient for use as the refrigerant evaporator 1 of an air conditioner for automobiles is obtained when the core thickness D is in the range of 15 mm to 40 mm, and higher cooling efficiency is obtained particularly at 37 mm or less. That is, by setting the core thickness D within the above range, the cooling performance value per unit weight of the refrigerant evaporator is further increased.

  FIG. 10 shows the relationship between the cooling performance value and the number of passes, and the number of passes is changed as described above. Further, the cooling performance value in this figure is the same average value as in the case of FIG. In FIG. 10, when the core thickness D is 15 mm or more and less than 20 mm, a sufficient cooling performance value can be obtained when the number of passes is 1 to 3, particularly when the number of passes is two. A performance value is obtained. Further, when the core thickness D is 20 mm or more and less than 30 mm, a sufficient cooling performance value can be obtained when the number of passes is 2 to 4, and a high cooling performance value is obtained particularly when the number of passes is two. can get. Further, when the core thickness D is 30 mm or more and 40 mm or less, a sufficient cooling performance value can be obtained when the number of passes is 2 to 5, especially when the number of passes is two or three. A performance value is obtained.

  That is, when the core thickness D is reduced, the width dimension of the tube 2 is reduced and the cross-sectional area of the tube 2 is reduced, and the pressure loss in the pipe is increased. In-pipe pressure loss is reduced. Also, if the number of passes is set to be small without changing the number of tubes 2, the number of tubes 2 per pass increases, and the cross-sectional area of the refrigerant flow path in the pass increases, thereby reducing the pressure loss in the tube. As a result, the refrigerant flow rate increases. On the other hand, if the number of passes is set to be large without changing the number of tubes 2, the number of tubes 2 per pass is reduced, and the cross-sectional area of the refrigerant flow path in the pass is narrowed. The refrigerant flow rate increases and decreases. That is, since the refrigerant flow velocity and the pressure loss in the pipe change depending on the core thickness D and the number of passes, heat transfer is performed in the refrigerant evaporator 1 by setting the core thickness D and the number of passes to the above ranges, respectively. The cooling performance value is increased by balancing the pressure loss of the refrigerant that affects the rate and the flow velocity.

  Therefore, according to the refrigerant evaporator 1 according to this embodiment, the condensed water is smoothly drained by the structure of the fin 3 so as to ensure the cooling performance, so that the fin dimensions can be set freely. . And by setting the fin height dimension Fh to 3 mm or more and 6 mm or less, the cooling performance value per unit weight of the refrigerant evaporator 1 regardless of the flow path height dimension Thi or the core thickness D in the tube. Thus, it is possible to obtain the refrigerant evaporator 1 that is lightweight and has high cooling performance.

  In addition, by setting the channel height dimension Thi in the tube to 0.7 mm or more and 1.2 mm or less, the refrigerant flow rate can be secured, and the cooling flow rate and the heat transfer area are balanced to increase the cooling performance value. Thus, the cooling performance value per unit weight of the refrigerant evaporator 1 can be further increased.

  Furthermore, by setting the thickness of the core to 40 mm or less, the cooling efficiency of the refrigerant evaporator 1 is increased, and the cooling performance value per unit weight of the refrigerant evaporator can be further increased.

  In this embodiment, the case where the tubes 2 are arranged in a row in the air flow direction is described. However, the present invention is not limited to this, and the tubes 2 may be arranged in a plurality of rows in the air flow direction. In this case, a path is formed in each of the upstream tube row and the downstream tube row in the air flow direction so that the refrigerant flows from the upstream tube row to the downstream tube row and from the downstream tube row to the upstream side. It can be configured to flow to the tube row.

  In this embodiment, the first louvers 13a, 14a and the second louvers 13b, 14b are formed on the fin 3, but these louvers may not be formed.

  As described above, the refrigerant evaporator according to the present invention can be applied to, for example, an automobile air conditioner.

It is a perspective view which expands and shows the upper end side of a tube and a fin. It is the perspective view which looked at the refrigerant evaporator which concerns on embodiment of this invention from the air flow upstream. (A) is the elements on larger scale of a fin, (b) is the sectional view on the AA line in (a). It is a graph when the relationship between the cooling performance value per unit weight of the refrigerant evaporator and the height dimension of the fin is shown, and the flow path height dimension in the tube is 0.5 mm. FIG. 5 is a view corresponding to FIG. 4 when the flow channel height dimension in the tube is 0.7 mm. FIG. 5 is a view corresponding to FIG. 4 when the flow path height dimension in the tube is 1.0 mm. FIG. 5 is a view corresponding to FIG. 4 when the flow path height in the tube is 1.2 mm. FIG. 5 is a view corresponding to FIG. 4 when the flow path height in the tube is 1.4 mm. It is a graph which shows the relationship between the cooling efficiency and the thickness of a core. It is a graph which shows the relationship between the cooling performance value and the number of passes.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Refrigerant evaporator 2 Tube 3 Fin 4 Core 11 1st junction surface part 12 2nd junction surface part 13 1st intermediate surface part 14 2nd intermediate surface part R Refrigerant flow paths P1-P3 path

Claims (6)

Refrigerant evaporation in which a plurality of tubes whose cross-sectional shape of the refrigerant flow path is long in the air flow direction are arranged side by side in the direction intersecting the air flow direction, and fins for heat transfer are arranged between adjacent tubes A vessel,
The fins extend substantially flat along the opposing outer surfaces of adjacent tubes, and are joined to the first and second joint surface portions joined in surface contact with the outer surface of the tubes, respectively. With the intermediate surface portion extending substantially perpendicular to the joint surface portion, it is formed in a corrugated plate shape when viewed in the air flow direction,
The refrigerant evaporator according to claim 1, wherein a dimension of the fins in the tube juxtaposition direction is set to 3.0 mm or more and 6.0 mm or less. The refrigerant evaporator according to claim 1,
The refrigerant evaporator characterized by the dimension of the tube juxtaposition direction in the refrigerant flow path of a tube being set to 0.7 mm or more and 1.2 mm or less. The refrigerant evaporator according to claim 1 or 2,
A refrigerant evaporator, wherein a dimension of a core constituted by tubes and fins is set to 15 mm or more and 40 mm or less. The refrigerant evaporator according to claim 3, wherein
The dimension of the core in the air flow direction is set to 15 mm or more and less than 20 mm,
The refrigerant evaporator according to claim 1, wherein a path through which a refrigerant flows is formed by a tube in the core, and the number of the paths is 1 to 3. The refrigerant evaporator according to claim 3, wherein
The dimension of the core in the air flow direction is set to 20 mm or more and less than 30 mm,
The refrigerant evaporator according to claim 1, wherein a path through which a refrigerant flows is constituted by a tube in the core, and the number of the paths is two to four. The refrigerant evaporator according to claim 3, wherein
The dimension of the core in the air flow direction is set to 30 mm or more and 40 mm or less,
The refrigerant evaporator according to claim 1, wherein a path through which a refrigerant flows is formed by a tube in the core, and the number of the paths is 2 to 5.

Images