US20060237178A1

US20060237178A1 - Heat exchanger

Info

Publication number: US20060237178A1
Application number: US11/407,398
Authority: US
Inventors: Yoshiki Katoh; Yoshiyuki Okamoto; Masahiro Shimoya; Toshio Ohara; Noriyoshi Miyajima
Original assignee: Denso Corp
Current assignee: Denso Corp
Priority date: 2005-04-22
Filing date: 2006-04-19
Publication date: 2006-10-26
Also published as: JP2006322698A; DE102006018532A1

Abstract

Air passages 18 are formed between surfaces of a plurality of heat transfer plates 12 laid together. A plurality of rib sections 14 extending orthogonal to the air flowing direction A are formed on a surface of the heat transfer plate 12 and projected into the air passage 18. By shifting the positions of the rib sections 14 to each other in the air flow direction, coolant passages 15, 16 are formed inside the plurality of rib sections 14. Between the plurality of rib sections 14, fin sections 17 are formed integral with the heat transfer plate 12 and projected from the plate surface. The fin section 17 has a protruded shape formed by pressing and having a cut portion partially cut a plate thickness of the heat transfer plate 2.

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention
The present invention relates to a heat exchanger, wherein fins are formed integral with a heat transfer plate constituting internal passages through which flows a heat exchanging fluid and, for example, useful for a vehicle air conditioner.
2. Description of the Related Art
In the prior art and, for example, in Japanese Unexamined Patent Publication No. 11-287580 (first patent document), a heat exchanger has been proposed wherein a plurality of rib sections, constituting internal passages through which passes a heat exchanging fluid, are formed integral with a heat transfer plate and operate as turbulence generators for disturbing a straight flow of air stream flowing on the outside of the heat transfer plates.
According to this structure, as an air-side heat transfer rate is improved by forming turbulence in the air stream, it is possible to eliminate fin members such as the corrugated fins in the conventional fin and tube type heat exchanger. Thus, the heat transfer plate could be produced solely by press-forming and brazing heat transfer plates.
Also, in Japanese Unexamined Patent Publication No. 2002-147983 (second patent document), as shown in FIG. 27, a heat exchanger is proposed wherein a plurality of rib sections 14, for constituting internal passages 15 for a heat exchanging fluid, are formed integral with a heat transfer plate 12, and a base plate section 13 having a flat surface is formed between the rib sections 14 adjacent to each other and fin sections 17, projected toward an air passage 18, are provided on the base plate section 13. Also, in this second patent document, the fin members, such as corrugated fins, are not fixed to the heat transfer plate 12.
In this prior art, as the plurality of rib sections 14 constituting the internal passages are disposed at the same positions as seen in the air flowing direction A, the rib sections 14, 14 in the adjacent heat transfer plates 12, 12 are directly opposed to each other while interposing the air passage 18.
As a result, an area S1 of the air passage 18 at a position forming the base plate section 13, that is, at a position forming the fins 17 reduces to S2 at a position forming the rib section 14. Accordingly, in this prior art, the air passage 18 repeats the reduction and enlargement of the cross-sectional area in accordance with whether or not the rib section 14 exists.
In this regard, in Japanese Unexamined Patent Publication No. 11-287580, while a local heat transfer rate in the vicinity of the rib section is better than in the conventional fin and tube type heat exchanger, the air-side heat transfer area becomes insufficient, whereby there may be a case that a necessary heat transfer performance cannot be ensured.
Also, in the heat transfer plate, as the base plate section having no rib section forms a flat surface extending in the air-flow direction, a temperature boundary layer grows on this flat surface to lower the local heat transfer rate to a great extent. This is also a cause for the deterioration of heat transfer performance.
To ensure the necessary heat transfer performance, it is necessary to increase the number of heat transfer plates. However, as the heat transfer plate has a large wall thickness for maintaining the required pressure resistance in comparison with the fin member, the total weight of the heat exchanger increases. Also, as the material cost of the heat transfer plate increases, the production cost of the heat exchanger becomes high.
In the second patent document, as the air passage 18 repeats the reduction and enlargement of the cross-sectional area in accordance with whether or not the rib section 14 exists, a pressure loss in the air-stream increases.
According to the second patent document, as gaps between tops of the rib sections 14 directly opposed to each other via the air passage 18(the portions having the area S2) are arranged on a straight line in the air flowing direction A, a main air stream linearly passes a portion having a reduced area S2 as shown by an arrow E.
Therefore, in the base plate section 13 constituting the enlarged portion having the passage area S1 in the heat transfer plate 12, a region F in which air stream dwells is formed along the surface of the base plate section 13, which significantly deteriorates the heat transfer rate on the surface of the base plate section 13.
In the second patent document, as the main stream of air moves straight past the portion having the reduced area S2 as shown by an arrow E, unless the fin 17 is projected into the gap through which the main stream E passes, it does not serve to improve the heat transfer performance.
Accordingly, in the second patent document, it is necessary that a height of the fin 17 is higher than a height of a top of the rib section 14, which forces the metallic material forming the heat transfer plate 12 to be excessively stretched during the machining of the fin 17. Accordingly, it is difficult to machine the fins 17.
Also, if the fin 17 is made to be higher than the top of the rib section 14, in a process of assembling the heat exchanger, the fin 17 is liable to collide with a peripheral member and be damaged.

SUMMARY OF THE INVENTION

In view of the above problems in the prior art, an object of the present invention is to improve the heat transfer performance of the plate type heat exchanger in which separate fin members are not combined with heat transfer plates constituting internal passages, without increasing the number of heat transfer plates.
Also, another object of the present invention is both to improve the heat transfer performance of the plate type heat exchanger of this kind and to facilitate the production of the heat exchanger.
To achieve the above objects, according to the inventive heat exchanger, a plurality of heat transfer plates (12) forming plate surfaces extending in the flow direction (A) of an external fluid are stacked perpendicular to said plate surfaces,
a gap is provided between said plate surfaces of said adjacent heat transfer plates (12) to form an external passage (18) through which said external fluid flows,
a plurality of rib sections (14) extending orthogonal to the flowing direction A of said external fluid are projected from said plate surfaces into said external passage (18) to be integral with said heat transfer plates (12),
by shifting the positions of the plurality of rib sections (14) in one of said adjacent heat transfer plates (12) relative to the positions of the plurality of rib sections (14) in the other of said adjacent heat transfer plates (12) as seen in the flowing direction A of said external fluid, said external passage (18) is formed in a meandering manner,
the plurality of rib sections (14) form an internal passage (15, 16) inside thereof, through which flows an internal fluid,
fin sections (17) are projected from said plate surfaces at positions between the adjacent rib sections (14) to be integral with the heat transfer plate (12), and
said fin section is press-formed (lances) so as to protrude a cut surface partially cut a plate thickness of said heat transfer plate (12).
According to this structure, as the external fluid impinges on the rib section (14) to generate turbulence, the local heat transfer rate is improved in the vicinity of the rib section (14). Simultaneously therewith, the external passage (18) is formed in a meandering manner, whereby a main stream of the external fluid can positively impinge onto a plate surface located between the plurality of rib sections (14). Thus, the local heat transfer rate is also improved on the plate surface between the rib sections (14).
Further, due to the tip end effect (an effect of thinning the temperature boundary layer) of the fin section (17), the local heat transfer rate of the fin section (17) is largely improved and an external fluid-side heat transfer area of the heat transfer plate (12) is increased by forming the fin section (17).
For the above-mentioned reasons, it is possible to improve the heat transfer rate of the plate type heat exchanger without increasing the number of heat transfer plates, and the practical merit thereof is significant.
As a main stream of the external fluid impinges onto the plate surface located between the plurality of rib sections (14) by forming the external passage (18) in a meandering manner, it is unnecessary to make the fin section (17) higher than the rib section (14) as described in the second patent document. Thus, it is possible to make the height of the fin section (17) lower than that disclosed in the second patent document.
Thus, when the fin section (17) is press-formed so as to protrude a cut portion partially cut the plate thickness of the heat transfer plate (12), the stretching of the heat transfer plate material becomes small to allow easy formation of the fin section (17) and an inconvenience is avoided in that the fin section (17) impinges on the peripheral members and is damaged during the assembly of the heat exchanger.
In the plate type heat exchanger which is the subject matter of the present invention, there is a problem of abnormal air noise (wind sound) caused by vortices generated downstream of the rib section (14) as disclosed in Japanese Unexamined Patent Publication No. 2002-48491. According to the present invention, it is possible to shift (vary), in the longitudinal direction of the rib section (14), the timing at which the external fluid stream moves over the rib section (14) by providing the fin section (17) between the plurality of rib sections (14).
Thereby, as the overlap of sound waves based on the vortices generated behind the rib section (14) is suppressed to avoid the resonance, the abnormal air noise (wind sound) based on the provision of the rib sections (14) is reduced.
According to the present invention, said heat transfer plates (14) are combined to form pairs, and said rib sections (14) and said fin sections (17) are formed integral with said pair of heat transfer plates (12), and the pair of heat transfer plates (12) are fixed together to form said internal passage (15, 16) inside the plurality of rib sections (14).
Thus, as the rib sections (14) and the fin sections (17) are formed integral with the pair of heat transfer plates (12), the above-mentioned effects are effectively exhibited.
Brazing is a representative means for fixing the pair of heat transfer plates (12) to each other. When the fin sections (17) are provided, cut holes (17 d) are simultaneously formed, which operate as air-discharging holes during the brazing, whereby the brazing between the pair of heat transfer plates (12) is improved.
In this regard, according to the present invention, the pair of heat transfer plates (12) includes two completely separated plates as well as a single plate folded at a center thereof to be two parts, each being half the overall size.
According to the present invention, positions in the pair of heat transfer plates (12) at which said rib sections (14) are formed are shifted in the flowing direction (A) of the external fluid, and
said internal passage (15, 16) may be formed by said rib sections (14) formed in one of the pair of heat transfer plates (12) and a plate surface of the other.
According to the present invention, said rib sections (14) may be formed in said pair of heat transfer plates (12) at the same positions as seen in the flowing direction (A) of said external fluid, and
said internal passages (15, 16) are formed by the combination of said rib sections (14) formed in said pair of heat transfer plates (12), respectively.
Thus, as the internal passages (15, 16) are formed by the combination of the rib sections (14) in the pair of heat transfer plates (12), it is possible to increase the area of the internal passage in comparison with the above-mentioned invention. Accordingly, it is possible to enlarge the mutual distance between the rib sections (14) and easily increase the number of fin sections (1).
According to the present invention, said heat transfer plate (12) is constituted by a single extrusion-formed plate material,
said rib sections (14) are formed by extrusion-forming a tubular-shaped portion on said single extrusion-formed plate material, and
said fin sections (17) are formed integral with said single extrusion-formed plate material to be projected from a plate surface of said single extrusion-formed plate material.
As the rib sections (14), that is the internal passages (15, 16) are formed by extrusion-forming a tubular-shaped portion on the single extrusion-formed plate material, a coupling structure is unnecessary for the purpose of forming the internal passages (15, 16). As a result, the coupling portions in the heat exchanger as a whole are largely reduced to improve the productivity of the heat exchanger.
According to the present invention, said heat transfer plate (12) has a base plate section (13) having a flat surface between the adjacent rib sections (14), and
said fin section (17) is formed in said base plate section (13).
Thereby, the fin section (17) is easily formed on the flat surface of the base plate section (13).
According to the present invention, a width (Fw) in the flowing direction (A) of said external fluid of said fin section (17) is 5 mm or less. Thereby, the tip end effect (effect for thinning the temperature boundary layer) of the fin section (17) is effectively exhibited and the external fluid side heat transfer rate of the heat transfer plate (12) is advantageously improved.
According to the present invention, said fin section (17) is a slit fin having an offset wall surface (17 a) set apart from a plate surface of said heat transfer plate (12) with a predetermined gap, wherein said offset wall surface (17 a) are coupled to a plate surface of said heat transfer plate (12) at two positions.
By employing such a slit fin, it is possible to effectively improve the external fluid side heat transfer performance of the heat transfer plate (12).
When a gap, between positions on the pair of heat transfer plates (12) opposed to each other to define said external passage (18), at to which positions are formed said slit fins 17, is defined as (L1 to L3), and a projected height of said offset wall surface (17 a) from a plate surface of said heat transfer plate (12) is defined as Fha1 to Fha3, the following relationship is satisfied:
Fha1 to Fha3≦½(L1 to L3).
If the offset wall surface (17 a) is formed thus, it is possible to form a meandering stream of the external fluid closer to a flat plate surface of the heat transfer plate (12). Thus, the impingement of the external fluid to a surface of the heat transfer plate (12) is facilitated.
According to the present invention, a cross-sectional shape of said rib section (14) has a curved surface, projected from the surface of said heat transfer plate, which is generally semicircular,
said slit fin (17) is located at a position directly on downstream side from said external fluid relative to said rib section (14), and
said offset wall surface (17 a) is inclined in the same direction as the inclination of the downstream side curved surface in the generally semicircular curved surface of said rib section (14).
Thereby, it is possible to form a stream P approaching the downstream side curved surface of the rib section (14) due to the guiding operation of the inclined surface of the offset wall surface (17 a) as illustrated in FIG. 25 described later. Thus, as the vortices (M′) are reduced to minimize the dwelling region caused by the vortices (M′), it is possible to improve the heat transfer rate of the downstream side curved surface of the rib section (14) and of the flat surface of the heat transfer plate (12).
According to the present invention, the cross-sectional shape of said rib section (14) is such that it has a curved surface protruded semi-circularly from a surface of said heat transfer plate (12),
said slit fin is disposed adjacent to said rib section (14) at a position directly on the upstream side of said external fluid, and
said offset wall surface (17 a) is inclined in the same direction as the inclination of the upstream side curved surface in a generally semicircular curved surface of said rib section (14).
Thereby, as the inclined surface of the offset wall surface (17 a) is inclined in the same direction as the inclination of the upstream side curved surface of the rib section (14), it is possible to make the external fluid stream meander smoothly on the upstream side.
According to the present invention, said slit fin (17) is disposed opposite to a front of said rib section (14) while interposing said external passage (18), and said offset wall surface (17 a) is formed to be parallel to a flat surface of said heat transfer plate (12).
In this regard, as illustrated in FIG. 21 described later, a reversed position of the stream in the external passage (18) is formed in front of the rib section (14). If the offset wall surface (17 a) inclined in a predetermined direction is disposed at this reversed position of the stream, the inclination of the offset wall surface (17 a) prevents the reversing of the stream. However, in the present invention, as the offset wall surface (17 a) is parallel to a flat surface of the heat transfer plate (12), the offset wall surface (17 a) becomes neutral relative to the reversal of the stream and does not obstruct the reversal of the stream.
According to the present invention, said external fluid is air and said internal fluid is coolant for cooling said air, wherein said heat exchanger is constituted as a cooling heat exchanger generating condensation water on the surface of said heat transfer plate (12), and
a gap (Q1, Q2) between said offset wall surface (17 a) and the surface of said heat transfer plate (12) is 0.3 mm or more.
According to the study of the present inventors, it has been confirmed that, when the gap (Q1, Q2) is 0.3 mm or more, the blockade of this gap (Q1, Q2) is avoidable and the drainage of condensation water is performed.
According to the present invention, said fin section (17) is a protruded (lanced) fin to protrude at a predetermined angle relative to the surface of said heat transfer plate.
The protruded (lanced) fin is simple in shape and easily formed in comparison with the slit fin defined by the above invention.
According to the present invention, said protruded fin (17) is triangular.
Such a triangular protruded fin (17), that is a delta wing, is liable to generate a Karman vortex which improves the local heat transfer rate on the periphery of the fin section due to the release of the Karman vortex.
According to the present invention, said triangular protruded fin (17) is inclined to the flowing direction (A) of said external fluid at an angle from 15° to 45°.
According to the present invention, said protruded fin (17) is rectangular. Here, “rectangular” includes square and trapezoidal.
According to the present invention, the inclination angle of said protruded fin (17) relative to the flowing direction (A) of said external fluid is determined in a small angle range from −30° to +30° so that a surface of the protruded fin (17) follows the flowing direction (A) of said external fluid, whereby the ventilation resistance of the external fluid is reduced.
According to the present invention, said external fluid is air, and internal fluid for cooling said air flows through said internal passage (15, 16),
said heat transfer plate (12) is disposed so that the longitudinal direction of said rib section (14) coincides with the upward/downward direction, and
an inclination angle of said protruded fin (17) is in a range from 60° to 120° relative to the flowing direction (A) of said external fluid so that a surface of said protruded fin (17) follows the longitudinal direction of said rib section (14).
Thereby, when the condensation water generated on the surface of the heat transfer plate (12) drops in the longitudinal direction of the protruded fin (17), the drainage of the condensation water is facilitated since the protruded fin (17) does not disturb the dropping of the condensation water.
According to the present invention, said internal passage has an upstream side internal passage (16) disposed on the upstream side in the flowing direction (A) of said external fluid and a downstream side internal passage (15) disposed on the downstream side in the flowing direction (A) of said external fluid,
said upstream side internal passage (16) and said downstream side internal passage (15) are respectively sectioned vertically to the flowing direction (A) of said external fluid into a plurality of areas (X, Y), and
passages connected in parallel to each other are constituted between the plurality of areas (X, Y) of said upstream side internal passages (16) and the plurality of areas (X, Y) of said downstream side internal passages (15).
Thereby, it is possible to lower the pressure loss in the internal passages (15, 16) as a whole by the parallel passage structure. Also, it is possible to reduce the number of rib sections (14) as well as enlarge a gap between the heat transfer plates (12) laid to each other, resulting in the reduction of the external fluid side ventilation resistance.
According to the present invention, if said downstream side internal passage (15) is an inlet side passage for said internal fluid, and said upstream side internal passage (16) is an exit side passage for said internal fluid, a high efficiency orthogonally opposed type heat exchanger is obtained.
According to the present invention, if said parallel passages couple the plurality of areas (X, Y) in said upstream side internal passage (16) and the plurality of areas (X, Y) in said downstream side internal passage (15) to each other in an X pattern, both the reduction of pressure loss in the internal passage (15, 16) and the uniformity of the temperature distribution of the blown-out external fluid are attainable.
Note, reference numerals in brackets indicate the correspondence of the respective members with concrete means in embodiments described later.
The present invention may be more fully understood from the description of preferred embodiments of the present invention, as set forth below, together with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

In the drawings:
FIG. 1 is an exploded perspective view of an evaporator according to a first embodiment of the present invention;
FIG. 2 is an exploded perspective view illustrating a coolant flow passage of the evaporator according to the first embodiment;
FIG. 3 is cross-section taken along a line III-III in FIG. 1;
FIG. 4 is a perspective view of part of a heat transfer plate shown in FIG. 3;
FIG. 5 is a perspective view of part of a core portion of a prior art fin and tube type heat exchanger;
FIG. 6 is a cross-section of a prior art finless type heat exchanger (shown in the first patent document);
FIG. 7 is a table showing the comparison of various items in the heat exchanger of the prior art with those in the first embodiment;
FIG. 8 is a graph of the local heat transfer rate in the finless type heat exchanger shown in the prior art (the first patent document);
FIG. 9 is a cross-section of a core portion of an evaporator according to a second embodiment of the present invention;
FIG. 10A is a perspective view of part of a heat transfer plate according to a third embodiment of the present invention and
FIG. 10B is a perspective view of part of a heat transfer plate according to a comparative example for the third embodiment;
FIG. 11 is a cross-section of a core portion of an evaporator according to a fourth embodiment of the present invention;
FIG. 12 is a perspective view of part the heat transfer plate shown in FIG. 11;
FIG. 13 is an enlarged view of part of FIG. 12;
FIG. 14 is a perspective view of part of a heat transfer plate according to a fifth embodiment of the present invention;
FIG. 15 is an exploded perspective view illustrating a structure of a coolant flow passage according to a sixth embodiment of the present invention;
FIG. 16 is a schematic perspective view illustrating a structure of a coolant flow passage in an evaporator according to the sixth embodiment;
FIG. 17 is a perspective view of part of a heat transfer plate illustrating a fin shape according to a seventh embodiment of the present invention;
FIG. 18 is a cross-section of a core portion of an evaporator according to an eighth embodiment of the present invention;
FIG. 19 is a cross-section of part of a core portion of an evaporator according to a ninth embodiment of the present invention;
FIG. 20 is a cross-section of part of a core portion of an evaporator according to a tenth embodiment of the present invention;
FIG. 21 is a cross-section of part of a core portion of an evaporator according to the tenth embodiment of the present invention;
FIG. 22 is a cross-section of part of a core portion of an evaporator according to a comparative example of the tenth embodiment of the present invention;
FIG. 23 is a cross-section of part of a core portion of an evaporator according to an eleventh embodiment of the present invention;
FIG. 24 is a cross-section of part of a core portion of an evaporator according to a twelfth embodiment of the present invention;
FIG. 25A is an enlarged cross-section of part of a core portion of an evaporator according to a comparative example of the twelfth embodiment, and FIG. 25B is an enlarged cross-section of part of a core portion according to the twelfth embodiment;
FIG. 26 is a cross-section of part of a core portion of an evaporator according to a thirteenth embodiment of the present invention; and
FIG. 27 is a cross-section of a main part of a heat exchanger according to a second patent document.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

First Embodiment

A first embodiment is an evaporator for a vehicle air conditioner. Initially, the total structure of the evaporator 10 for a vehicle air conditioner will be described. FIG. 1 is an exploded perspective view illustrating a summary of the total structure of the evaporator, and FIG. 2 is an exploded perspective view wherein a coolant passage indicated by arrows is added to FIG. 1. FIG. 3 is a lateral cross-section illustrating a laminated structure of heat transfer plates 12, and is a cross-section taken along a line I-I in FIG. 1. FIG. 4 is an enlarged perspective view of part of a heat transfer plate 12.
The total structure of the evaporator shown in FIGS. 1 and 2 may be substantially the same as that disclosed in the above-mentioned first patent document (Japanese Unexamined Patent Publication No. 11-287580). The evaporator 10 is an orthogonal opposed-flow-type heat exchanger wherein a flowing direction A of conditioning air and a flowing direction B of coolant in a heat transfer plate (up-down direction in FIG. 1) are orthogonal to each other, and the upstream (inlet) side passage of the coolant stream is located downstream from the air flowing direction A, while the downstream (exit) side passage of the coolant is located upstream from the air flowing direction A. In this regard, in the evaporator 10, air is an external fluid (cooled fluid) and coolant is an internal cooling fluid.
This evaporator 10 constitutes a core section 11 for carrying out the heat exchange between the conditioned air and the coolant solely by stacking a number of heat transfer plates 12 in the direction vertical to the plate surface (in the direction orthogonal to the air flowing direction A). In this regard, at the uppermost and lowermost ends of these heat transfer plates 2, tanks 20 to 23, described later, are formed. As a portion in which the tank 20 to 23 is formed does not allow air to pass therethrough, the core section 11 is formed in an intermediate region of the heat transfer plate 12 except for the tanks 20 to 23 formed at the upper and lower ends.
The respective heat transfer plate 12 is press-formed from a metallic sheet and, more concretely, from A3000-type aluminum core material clad with A400-type aluminum material on both side surface thereof. A plate thickness t of the heat transfer plate 12 (FIG. 2) is as small as, for example, 0.15 mm. The heat transfer plates 12 have a generally rectangular planar shape having the same dimensions.
Next, a concrete shape of the heat transfer plate 12 will be described with reference to FIG. 3. The respective heat transfer plate 12 has rib sections 14 formed from a flat base plate 13 by the press-forming. The rib sections 14 are of a longitudinal shape continuously extending in parallel to each other in the longitudinal direction of the heat transfer plate 12. While a cross-sectional shape of the rib section is generally semicircular in FIG. 3, it may be other shapes, for example, a trapezoid having rounded corners.
An inside space of the rib section 14 forms an internal passage, more concretely, a coolant passage 15, 16, through which flows a low pressure side coolant after passing through a pressure-reduction means (an expansion valve or others) in a refrigerant cycle. As the longitudinal direction of the heat transfer plate 12 coincides with upward and downward, the longitudinal direction of the rib sections 14 is also coincides with upward and downward; that is, orthogonal to the air flowing direction A.
At a central position of a rib pitch Rp which is a mutual distance between the adjacent rib sections 14 in one heat transfer plate 12, there the rib section 14 in the other heat transfer plate 12 mated therewith. Accordingly, when the pair of heat transfer plates 12, 12 are located so that the rib sections 14 of the respective heat transfer plates are opposed to each other toward outside and the base plate sections 13 are in contact with each other, the inner side of the rib section in the one heat transfer plate 12 is tightly closed by a central wall surface of the base plate section 13 in the other heat transfer plate 12.
Accordingly, coolant passages 15 and 16 are formed between the respective inner side of the rib section 14 and the base plate section 13 in the mated heat transfer plate 12. The coolant passage 15 constitutes a leeward coolant passage disposed on the downstream side area of the air flowing direction A, while the coolant passage 16 constitutes a windward coolant passage disposed on the upstream side area of the air flowing direction A.
A fin section 17 is formed integrally at a position at which the base plate sections 13 in the respective heat transfer plates 12, 12 are in contact with each other. The fin section 17 is formed between the adjacent rib sections 14. In this embodiment, the fin sections 17 in the pair of heat transfer plates 12, 12 are formed at the same position as seen in the air flowing direction A.
The fin section 17 in this embodiment constitutes a slit fin. The slit fin is one having an offset wall surface 17 a which is a top wall surface apart from a surface of a mother material (concretely, a surface of the base plate section 13), and at a predetermined gap, to define a space for allowing air to pass therethrough between the offset wall surface 17 a and the surface of the mother material as shown in FIG. 4, wherein at least two positions of the offset wall surface 17 a are physically fixed to the surface of the mother material.
In the embodiment shown in FIG. 4, the fin section 17 is of a U-shape in which left and right ends of the offset wall surface 17 a are fixed to the base plate section 13 with two lateral walls 17 b and 17 c.
In this regard, a fin height Fh which is a height of the offset wall surface 17 a of the fin section 17 is the same as a height (rib height) Rh of the rib section 14 or slightly lower than height Rh as shown in FIG. 3. In the embodiment shown in FIG. 4, the slit fin 17 has a widthwise dimension Fw in the air flowing direction A smaller than a dimension orthogonal to the air flowing direction A (an upward/downward dimension in FIG. 4).
To form such a fin section 17, two cut lines are provided in a fin-forming area of the base plate section 13 at a distance corresponding to the fin width Fw, after which a region between the two cut lines is pressed to have a U-shaped cross-section.
The U-shape (a slit fin shape) of the fin section 17 constitutes a protruded shape having a cut surface passing through a plate thickness of the heat transfer plate 12. Thereby, a cut hole 17 d accompanied with the formation of the fin section 17 is formed in the fin-forming area of the base plate section 13.
In this regard, as the fin-forming area is provided at a position at which the base plate sections 13 in the pair of heat transfer plates 12, 12 are in contact with each other, there is no risk in that the coolant leaks from the coolant passages 15 and 16 even if the cut hole 17 d is formed in the base plate section 13.
In this regard, in FIGS. 1 and 2, the above-mentioned fin sections 17 are not shown for simplifying the illustration. In FIG. 3, the number of the rib sections 14 in the pair of heat transfer plates 12, 12 is five. On the other hand, in FIGS. 1 and 2, the number of the rib sections 14 in one of the pair of heat transfer plates 12, 12 is six, and that in the other of the pair is five. The number of the rib sections 14, that is, the number of coolant passages 15, 16 may, of course, be increased or decreased in accordance with the required performance or contour of the evaporator 10.
At each of opposite end areas of the respective heat transfer plate 12 as seen in the direction B (the longitudinal direction of the heat transfer plate) orthogonal to the air flowing direction A, two tank sections 20 to 23 divided in the widthwise direction of the heat transfer plate (in the air flowing direction A) are formed. That is, there are two tank sections 20 and 22 at the upper end area of the heat transfer plate 12, and two tank sections 21 and 23 at the lower end area thereof.
The tank sections 20 to 23 are formed to be projected in the same direction as the rib sections 14 in the respective heat transfer plate 12. A projected height of the tank section 20 to 23 is one half of a tube pitch Tp (see FIG. 3) so that tops of the adjacent tank sections 20 to 23 are brought into contact and fixed with each other.
In this regard, the projected height of the tank section 20 to 23 includes the plate thickness t of the heat transfer plate 12. The tube pitch Tp is a distance between the arranged heat transfer plates 12. Also, a space pitch Sp is a value obtained by subtracting the plate thicknesses t of two heat transfer plates 12 from the tube pitch Tp; i.e., Tp−2t.
In the embodiment shown in FIG. 3, while the rib height of the rib section Rh is determined to be one half of the tube pitch Tp, that is, generally equal to the projected height of the respective tank section 20 to 23, this is not limitative, but the rib height Rh of the rib section 14 may be slightly increased or decreased relative to the respective tank section 20 to 23.
As mentioned above, as the tank sections 20 to 23 are projected in the same direction as the rib sections 14, and recessed longitudinal opposite end areas formed by the projection of the rib sections 14 are continuous to the recessed shape of the tank sections 20 to 23, both end portions of the windward coolant passage 16 communicate with the leeward upper and lower tank sections 22, 23, and both ends portions of the leeward coolant passage 15 communicate with the windward upper and lower tank sections 20, 21.
In this regard, the leeward tank section 20 and the windward tank section 22 on the upper side of the heat transfer plate define the coolant passages independent from each other, and the leeward tank section 21 and the windward tank section 23 on the lower side of the heat transfer plate define the coolant passages independent from each other.
As communication openings 20 a to 23 a are provided at the centers of the tops of the respective tank sections 20 to 23, it is possible to communicate the communication openings 20 a to 23 a with each other by bringing the projected tops of the tank sections 20 to 23 adjacent to each other and fixing them together.
Thereby, it is possible to mutually communicate the coolant passages of the tank sections 20 to 23, between adjacent heat transfer plates, as seen in the left/right direction in FIGS. 1 and 2.
As the plurality of rib sections 13 in the respective heat transfer plate 12 are disposed while being shifted from those in the adjacent heat transfer plate 14 as seen in the widthwise direction of the heat transfer plate 12 (in the air flowing direction A) as shown in FIG. 3, it is possible to oppose the respective rib section 14 to the base plate section 13 in the adjacent heat transfer plate 12.
As the rib height Rh of the rib section 14 is determined to be approximately equal to half a tube pitch Tp as described before, a gap is formed between a top of the rib section 14 on the convex side and the base plate section 13 in the adjacent heat transfer plate 12, whereby a meandering curved air passage 18 is continuously formed along a total length (in the air flowing direction A) of the heat transfer plate 12 as shown in FIG. 3 by an arrow A1. The fin sections 17 constituting U-shaped slit fins are arranged in this wavy air passage 18 adjacent to the respective rib sections 14.
Then, a portion for feeding and discharging coolant relative to the core section 11 is described below. As shown in FIGS. 1 and 2, end plates 24, 25 having the same size as the heat transfer plate 12 are disposed at opposite ends in the laying direction of the heat transfer plates. The end plate 24, 25 is a flat plate capable of being in contact with a convex side of the tank sections 20 to 23 of the heat transfer plate 12 and fixed thereto.
Into holes provided in the vicinity of the upper end of the left side end plate 24 in FIGS. 1 and 2, a coolant inlet pipe 24 a and a coolant exit pipe 24 b are fixed, wherein the coolant inlet pipe 24 a communicates with a communication opening 20 a formed at a top of the leeward side tank section 20 formed at an upper end of the leftmost heat transfer plate 12. The coolant exit pipe 24 b communicates with a communication opening 22 a formed at a top of the windward tank section 22 formed at an upper end of the leftmost heat transfer plate 12.
The left side end plate 24 is formed of a both-side aluminum clad material in the same manner as in the heat transfer plate 12, and brazed to the coolant inlet and exit pipes 24 a, 24 b. The right side end plate 25 is formed of metallic material clad with brazing metal on one side to be brazed to the heat transfer plate 12.
A liquid-vapor-type two-phase coolant passing through a pressure-reduction means such as an expansion valve is fed into the coolant inlet pipe 24 a. On the other hand, the coolant exit pipe 24 b is connected to a suction side of a compressor, not shown, so that evaporated vapor-liquid type coolant evaporated in the evaporator 10 is guided to the suction side of the compressor.
In a group of the plurality of heat transfer plates 12 stacked to each other in the left/right direction in FIGS. 1 and 2, the leeward side coolant passage 15 formed in the interior of the rib sections 14 described before constitutes the inlet side coolant passage in all of the evaporator as the coolant is fed from the coolant inlet pipe 23.
On the other hand, the windward side coolant passage 16 formed in the interior of the rib sections 14, described before, constitutes the outlet side coolant passage directing the coolant passing through the leeward side (inlet side) coolant passage 15 as the coolant is fed to the coolant outlet pipe 24 b.
Next, all the coolant passages in the evaporator 10 will be described with reference to FIG. 2. The leeward tank sections 20 and 21 in the tank sections 20 to 23 disposed at the upper and lower ends of the evaporator 10 constitute the coolant inlet side tank sections, while the windward tank sections 22 and 23 constitute the coolant exit side tank sections.
The leeward and upper side coolant inlet side tank section 20 is divided by a partition (not shown) disposed at an intermediate position in the laying direction of the heat transfer plates 12 into the left side flow passage in FIG. 2 (a flow passage on the area X side) and the right side flow passage in FIG. 2 (a flow passage on the area Y side).
Similarly, the windward and upper side coolant exit side tank section 22 on is divide into the left side flow passage in FIG. 2 (a flow passage on the area X side) and the right side flow passage in FIG. 2 (a flow passage on the area Y side). These divided portions are simply constituted by using the heat transfer plate 12 among those described before, which are located at the intermediate position and the communication openings at the tops of the tank sections 20 and 22 thereof are blocked to be a barrier wall (a blind lid).
According to the coolant passage structure in FIG. 2, the vapor-liquid type two phase coolant depressurized by the expansion valve enters the leeward upper side inlet side tank section 20 from the coolant inlet pipe 24 a as shown by an arrow a. As the flow passage of the inlet side tank section 20 is divided into the left and right areas X and Y by the partition not shown, the coolant is fed solely into the left side area X of the inlet side tank section 20.
In the left side area X in FIG. 2, the coolant descends in the coolant passage 15 formed by the leeward side rib sections 14 as indicated by an arrow b and enters the lower side inlet side tank section 21. Next, the coolant moves through the lower side inlet side tank section 21 into the right side area Y in FIG. 2 as indicated by an arrow c, and rises in the coolant passage 15 formed in the right side area Y by the leeward side rib section 14 of the heat transfer plate 12, as indicated by an arrow d to enter the right side area Y of the upper side inlet side tank section 20.
Here, the communication opening 20 a of the inlet side tank section 20 in the rightmost side heat transfer plate 12 communicates with the communication opening 22 a of the exit side tank section 22 located on the upper side of the rightmost side heat transfer plate 12 via a communication passage (not shown, see an arrow f) formed in the vicinity of the upper end of the right side end plate 25.
Accordingly, the coolant entering the flow passage of the right side area Y in the upper side inlet side tank section 20 flows rightward as indicated by an arrow e, and thereafter, passes the communication passage (not shown) in the vicinity of the upper end of the right side end plate 25 as indicated by an arrow f and enters the flow passage in the right side area Y of the upper side exit side tank section 22.
Here, as the flow passage of the exit side tank section 22 is divided into the left and right side areas X and Y by the above-mentioned partition not shown, the coolant solely enters the flow passage in the right side area Y of the exit side tank section 22 as indicated by an arrow g. Next, the coolant entering the right side area Y in this tank section 22 descends the coolant passage 16 formed by the windward rib section 14 in the heat transfer plate 12 and enters the right side area Y of the lower side exit side tank section 23.
The coolant moves from the right side area Y to the left side area X in FIG. 2 through the lower side exit side tank section 23 as indicated by an arrow i, and thereafter, rises the coolant passage 16 formed by the windward rib section 14 of the heat transfer plate 12 as indicated by an arrow j to enter the flow passage in the left side area X of the upper side exit side tank section 22. The coolant passes through the exit side tank section 22 leftward as indicated by an arrow k, and is discharged from the coolant exit pipe 24 b to the outside of the evaporator.
In the evaporator 10 shown in FIGS. 1 and 2, the coolant passage is structured as described above, whereby it is possible to assemble the evaporator 10 by laying the respective components (12, 24, 25, 24 a and 24 b) to be in contact with each other, holding such a stacked state (assembled state) by a suitable jig, putting the assembly into the brazing furnace, and heating the same to a melting point of the brazing material. Thus, the assembly of the evaporator 10 is completed.
Next, the operation of the above-mentioned evaporator 10 will be described. The evaporator 10 shown in FIGS. 1 and 2 is accommodated in a case of an air conditioning unit not shown, upside down, so that air flows in the direction A due to the action of an air conditioning blower.
When the compressor for the refrigeration cycle operates, the low pressure vapor-liquid type two-phase coolant decompressed by the expansion valve not shown flows via the above-mentioned passages indicated by the arrows a to k shown in FIG. 2. On the other hand, due to the gap formed between the rib section 14 projected on the outer surface side of the heat transfer plate 12 and the base plate section 13, the air passage 18 meandering as shown by an arrow A1 in FIG. 3 is continuously formed in the widthwise direction of the heat transfer plate (the air flowing direction A).
As a result, the conditioned air sent in the direction A passes through the air passage 18 between the two heat transfer plates 12 and 12 while meandering as shown by the arrow Al. Since the coolant sucks the evaporation latent heat from this air stream and vaporizes, the conditioned air fed in the direction A becomes a cold wind.
At this time, as the inlet side coolant passage 15 is arranged on the leeward side and the exit side coolant passage 16 is arranged on the windward side relative to the flowing direction A of the conditioned air, the relationship of the coolant inlet and exit relative to the air flow becomes a counter flow relationship.
Further, as the air flowing direction A is orthogonal to the longitudinal direction of the rib section 14 of the heat transfer plate 12 (the coolant flowing direction B in the coolant passage 15, 16) on the air side, and the rib section 14 forms a convex heat transfer surface projected orthogonal to the air flow, the straight advancement of air is disturbed by this orthogonally extended rib section 14. Accordingly, the air stream is disturbed to be turbulent, whereby the air-side heat transfer rate is significantly improved.
In this regard, in the plate type heat exchanger wherein the core section 11 is structured solely by the heat transfer plates 12 as in this embodiment, there is a problem in that the air-side heat transfer area is largely reduced in comparison with the conventional fin-and-tube type heat exchanger, whereby it is difficult to ensure the required heat transfer performance.
In view of such a point, the present inventors have studied various countermeasures. For example, it is thought that the air-side heat transfer rate is improved by increasing a rib height Rh of the rib section 14 to further facilitate the generation of turbulence. However, as the ventilation resistance becomes naturally larger if the rib height Rh increases, it is impossible to improve the performance in view of the ventilation resistance ratio. Similarly, as the increase in the number of the rib sections 14 results in the large ventilation resistance, it is impossible to improve the performance in view of the ventilation resistance ratio.
Although the reduction of the tube pitch Tp is advantageous for improving the heat transfer performance, this is defective in that the number of heat transfer plates 12 increases, resulting in a larger weight of heat exchanger as well as a larger ventilation resistance.
Under such circumstances, according to this embodiment, a fin section 17 constituting a U-shaped slit fin is provided at a position between the every adjacent rib sections 14, that is, a position corresponding to the flat base plate section 13.
According to this structure, as air flows along inner and outer surfaces of the U-shaped fin section 17 whereby the inner and outer surfaces of the U-shaped fin section 17 become the air-side heat transfer area, the air-side heat transfer area is largely increased in comparison with one having no fin section 17.
In addition thereto, it is possible to effectively improve the air-side heat transfer rate of the heat transfer plate 12. That is, while the air-side heat transfer rate is liable to be reduced in the base plate section 13 as a temperature boundary layer grows to be thick on a flat surface of the base plate section 13 in the heat transfer plate 2 in the air flowing direction A, it is possible to divide the temperature boundary layer on the flat surface of the base plate section 13 to restrict the growth of the temperature boundary layer by providing the fin section 17. Also, the air-side heat transfer rate of the fin section 17 itself is sufficiently improved by the tip-end effect of the fin section.
Further, due to the meandering of the air stream in the air passage 18 as shown in FIG. 3 by an arrow Al, it is possible to alternately impinge a main air stream on the surface of the rib section 14 and the flat surface of the base plate section 13. Thereby, it is possible to improve the air-side heat transfer rate in the base plate section 13 by thinning the temperature boundary layer on the flat surface of the base plate section 13.
As described above, according to this embodiment, it is possible to effectively improve the heat transfer performance of the plate type heat exchanger while restricting the increase of the ventilation resitance.
The improvement effect of the heat transfer performance according to this embodiment will be concretely described below in comparison with the conventional fin-and-tube type heat exchanger and the finless type heat exchanger disclosed in the first patent document.
FIG. 5 is a perspective view of part of a core portion of the conventional fin and tube type heat exchanger, wherein a corrugated fin 51 is fixed between flat tubes 50A and 50B FIG. 5.
FIG. 6 is a cross-section of the finless type heat exchanger shown in the first patent document, corresponding to a cross-section taken along a line I-I in FIG. 1, having no fin section 17 according to this embodiment.
FIG. 7 is a table illustrating the comparison of various items of the conventional fin-and-tube type heat exchanger (1) shown in FIG. 5, the finless type heat exchanger (2) according to the first patent document shown in FIG. 6 and the inventive heat exchanger according to this embodiment shown in FIGS. 3 and 4. In this Table, each of the item values of the fin-and-tube type heat exchanger (1) is selected as a reference value (100), and the item values of the heat exchangers (2) and (3) are represented as ratios to the reference values of the heat exchanger (1).
The items in FIG. 7 are calculated in accordance with the following conditions.
Contour size of heat exchanger: Width W 260 mm×height H 215 mm×depth D 38 mm
Note the width W is a dimension in the plate-stacking direction and the depth D is a thickness dimension in the air flowing direction.
Air flow: 500 m³/h, the ventilation resistance in the core section is the same in the heat exchangers (1) to (3).
Fin pitch fp: 2.6 mm and fin height fh: 6 mm in the heat exchanger (1).
Thickness t: 0.15 mm, space pitch Sp: 2.6 mm, pitch Rp of the rib sections: 7.1 mm and height Rh of the rib section: 1.45 mm in the heat transfer plate 12 in the heat exchanger (2).
Thickness t: 0.15 mm, space pitch Sp: 3.0 mm, pitch Rp of the rib sections: 7.1 mm, height Rh of the rib section: 1.45 mm: fin height Fh: 1.0 mm, fin width Fw: 0.8 mm in the heat transfer plate 12 in the heat exchanger (3). Note the fin pitch Fp is one half of the pitch Rp.
As is apparent from a comparison of the items in the respective heat exchangers (1), (2) and (3) shown in FIG. 7, in the finless type heat exchanger (2) disclosed in the first patent document, while the air-side heat transfer rate is largely improved relative to the fin-and-tube type heat exchanger (1), there is a defect in that the air-side heat transfer area Fa is largely reduced.
FIG. 8 illustrates the variation of the air-side local heat transfer rate in the finless type heat exchanger (2) of the first patent document. As the air stream impinges on a convex front surface of the rib section 14 in the heat transfer plate 12 to become turbulent, the local heat transfer rate is largely improved. On the contrary, on the flat surface of the base plate section 13 having no rib section 14, it is apparent that the temperature boundary layer grows to largely deteriorate the local heat transfer rate.
Contrarily, according to this embodiment, the fin section 17 constituting the U-shaped slit fin is disposed at a position between the adjacent rib sections 14 in the heat transfer plate 12; i.e., in the flat base plate section 13. As the heat transfer area of the heat transfer plate 12 is significantly increased by the formation of the fin section 17 and the temperature boundary layer on the flat surface of the base plate section 13 is divided by the fin section 17 and thinned due to the tip end effect, the heat transfer rate in the base plate section 13 is also improved.
As mentioned above, according to the heat exchanger (3) of this embodiment, it is possible to largely increase the air-side heat transfer area Fa and, simultaneously therewith, to improve the air-side heat transfer rate αa, in comparison with the finless type heat exchanger (2) disclosed in the first patent document, as illustrated in FIG. 7.
In this regard, in the heat exchanger (3) of this embodiment, the fin section 17 is added to the heat exchanger (2). Therefore, if the structure is as it is, the ventilation resistance increases. Actually, the space pitch Sp is increased from 2.6 mm in the heat exchanger (2) to 3.0 mm. Accordingly, it is possible to equalize the ventilation resistance of the inventive heat exchanger (3) to that in the heat exchanger (2) as described in the preconditions for the calculation described before.
According to the heat exchanger (3) in this embodiment, the number of heat transfer plates 12 to be used is reduced due to the enlargement of the space pitch Sp. Thereby, an area of the coolant passage becomes smaller than that in the heat exchanger (2), and the in-tube side heat transfer rate αr is more improved than in the heat exchanger (2).
When the heat exchanger is structured as an evaporator 10 for cooling air as in this embodiment, the moisture in air is condensed due to the cooling action of the evaporator 10 to generate condensation water. The drainage of this water is an important problem.
In the finless type heat exchanger (2) and the inventive heat exchanger (3) in this embodiment, the air stream impinges onto a front side of the convex of the rib section 14 extending upward/downward and generates the condensation water which moves to a rear side of the convex of the rib section 14 due to a wind pressure of the air stream and drops down along the rear side of the convex surface of the rib section 14 due to the gravity.
At that time, as fin section 17 is disposed more behind the convex surface of the rib section 14 in the heat exchanger (3) of this embodiment, condensation water is favorably drained along the rear side of the convex surface of the rib section 14 even if the fin section 17 is provided. Thus, an inconvenience, such as the increase in ventilation resistance caused by the dwell of the condensation water within the core section, is avoidable.
Further, according to this embodiment, an effect is obtainable in that abnormal air noise (wind sound) generated behind the rib section 14 (downstream of the air flow) is minimized due to the existence of the fin section 17.
That is, as described in Japanese Unexamined Patent Publication No. 2002-48491, in the finless type heat exchanger (2) disclosed in the first patent document, a layer peeled off from the main air stream is generated at the rear end of the rib section 14 as seen in the air flowing direction, and generates vortices therein. Further, as the rib sections 14 linearly extend orthogonal to the air flowing direction A while maintaining the same height, they generate vortices at the same time at the rear ends of the rib sections 14. The simultaneous generation of vortices coupled in the longitudinal direction of the rib sections causes the overlap of sound waves to amplify the abnormal air noise (wind sound).
On the contrary, according to this embodiment, as a position at which the fin section 17 is formed and a position at which the fin section is not formed are alternately present in the widthwise direction of the air passage 18 (in the longitudinal direction of the rib section 14), the variation occurs in the air stream in the longitudinal direction of the rib section 14, whereby it is possible to shift the timing at which the air stream moves over the rib section 14 in the longitudinal direction of the rib section 14. Thus, it is possible to divide vortices generated at the rear end of the rib section 14 as seen in the air flowing direction.
Thereby, as the overlap of sound waves based on the vortices generated at the rear ends of the rib sections 14 is restricted to avoid the resonance, the abnormal air noise (wind sound) caused by the rib sections 14 is suppressed.
As a result, it is possible to restrict the generation of the overlap of sound waves in the longitudinal direction of the rib section 14 and suppress the resonance phenomenon, resulting in the reduction of abnormal air noise (wind sound).
Also, according to this embodiment, as the fin sections 17 are pressed into a shape from the base plate section 13 of the heat transfer plate 12, the a hole 17 d is formed in the base plate section 13 at a position at which the fin section 17 is formed. Due to this cut hole 17 d, an additional effect is obtained to improve the brazing function of the heat exchanger.
That is, as shown in FIG. 3, according to this embodiment, a relatively wide fixing surface is formed between the adjacent rib sections 14 wherein flat surfaces of the base plate sections 13 in two heat transfer plates 12 are brought into contact with each other. In such a relatively wide fixing surface, the brazing defect is liable to occur due to the existence of air layers in micro-gaps of the fixing surface. In this embodiment, however, as the cut holes 17 d operate as air-discharging holes for discharging air on the fixing surface, the base plate sections 13 are favorably brazed to each other via the relatively wide fixing surface.
According to the above-mentioned first embodiment, the basic configuration of the heat transfer plate 12 is a plain plate disposed to form a flat surface in the air flowing direction A, and the rib section 14, the fin section 17 and the tank section 20 to 23 are formed in this plain plate. An intermediate portion of the heat transfer plate 12, except for the upper and lower end tank sections 20 to 23; that is, the core section 11, may not be a flat surface but may be a wavy surface (a curved surface gradually meandering in a wavy form). Even in such a structure, the same operation and effect are obtainable as in the first embodiment.

Second Embodiment

The rib sections 14 in the two heat transfer plates 12 fixing the base plate sections 13 to each other are disposed at positions shifted from each other in the air flowing direction A in the first embodiment. Contrarily, in a second embodiment, as shown in FIG. 9, the rib sections 14 in the two heat transfer plates 12 fixing the base plate sections 13 to each other are disposed at the same position in the air flowing direction A.
In the second embodiment, as the rib sections 14 having a semicircular cross-section in the two heat transfer plates 12 are combined at the same position to form circular coolant passages 15 and 16, the passage area of the respective coolant passage 15, 16 becomes larger.
Thereby, it is possible to decrease the number of the rib portions 14 to lengthen a mutual distance between the adjacent rib sections 14; i.e. a length of the base plate section 13 in the air flowing direction. Therefore, as shown in FIG. 11, it is possible to arrange three fin sections 17 between the adjacent rib sections 14.
According to the second embodiment, in accordance with the increase in the passage areas of the coolant passages 15 and 16, the coolant flow speed becomes lower and, as a result, the in-tube side heat transfer rate αr is smaller than that in the first embodiment. However, as the air side heat transfer performance is improved due to the increase of the number of the fin sections 17, and compensates for the reduction of the coolant side heat transfer performance, the heat transfer performance becomes better as a whole than in the first embodiment.
In this regard, it is, of course, possible to variously increase or decrease the number of the fin sections 17 in accordance with the specifications of the evaporator 10.

Third Embodiment

In the second embodiment, the rib sections 14 having a semicircular cross-section in the two heat transfer plates 12 are combined together at the same position to obtain the coolant passages 15 and 16 having a circular cross-section. According to a third embodiment, as shown in FIG. 10A, tubular coolant passages 15 and 16 having a circular cross-section are formed in a single heat transfer plate 12 by extrusion. Due to this tubular shape, rib sections 14 having a semicircular cross-section are projected from front and rear surfaces of the single heat transfer plate 12.
After this extrusion process, the fin sections 17 are pressed from a flat surface of the base plate section 13 between the adjacent rib sections 14. In the embodiment shown in FIG. 10A, the fin section 17 is formed as a U-shaped slit fin.
According to the third embodiment, as the tubular coolant passages 15 and 16 are formed in a single heat transfer plate 12 by the extrusion, the number of heat transfer plates 12 to be stacked is halved. Thereby, the positions necessary for brazing are largely decreased to improve the productivity of the heat exchanger to a large extent.
FIG. 10B illustrates a comparative example of the third embodiment, wherein the fin sections 17 are not formed. As the front and rear surfaces of a single heat transfer plate 12 are the air side heat transfer surface in this comparative example together with the third embodiment, even if the fin sections 17 are pressed as in the third embodiment, a large increase in the air side heat transfer area is not expected.
However, as the heat transfer rate in the base plate section 13 is largely improved by the tip end effect derived from the provision of the fin sections 17, it is possible to realize an improvement in the heat transfer performance as a whole.

Fourth Embodiment

In the first to third embodiments, while the description has been made on the structure wherein the fin section 17 is a U-shaped slit fin having an offset wall surface 17 a, the fin section 17 should not be limited to the slit fin but may be a simple protruded fin. Here, the protruded fin is one which is coupled to a mother material surface (concretely the surface of the base plate section 13) at at least one point and protruded by pressing to have a predetermined angle to the mother material surface.
In the fourth embodiment, as shown in FIGS. 11 and 12, the fin section 17 is a triangular fin protruded of a triangular piece cut at a right angle from the flat surface of the base plate section 13. Due to this protruding of the triangular fin, a cut hole 17 d is formed on the flat surface of the base plate section 13. This cut hole 17 d serves for discharging air when the brazing is carried out.
The fin sections 17 are provided at the same position in the two heat transfer plates 15 and 16 constituting the coolant passages 15 and 16 (the same position in the air flowing direction A). Also, the triangular piece constituting the fin section 17 is inclined at a predetermined angle 0 relative to the air flowing direction A. FIG. 13 is an enlarged view illustrating such a slanted arrangement of fin section 17.
The triangular fin section 17 constitutes a delta wing which is liable to generate a Karman vortex. In this regard, if the inclination angle θ of the fin section 17 constituting the delta wing is determined in a range from 15 to 45°, it is possible to facilitate the effect for improving the heat transfer rate in the base plate section 13 by the generation of the Karman vortex.
While the projected height of the fin section 17 is larger than half a tube pitch Tp in FIG. 11, the projected height may, of course, be increased or decreased if necessary and to, for example, smaller than half a tube pitch Tp.
The protruded fin (the fin section 17) according to the fourth embodiment is not limited to a triangle but may be other shapes, such as a rectangle.
If the protruded fin (the fin section 17) of the fourth embodiment is arranged generally parallel to the air flowing direction A, it is advantageous for reducing the ventilation resistance. Here, “generally parallel to the air flowing direction A” means that the inclination angle θ is within a range from −30 to +30°.
As a surface of the protruded fin (the fin section 17) becomes generally parallel to the longitudinal direction of the rib section 14 (that is, upward/downward direction of the evaporator) if the protruded fin (the fin section 17) of the fourth embodiment is arranged generally orthogonal to the air flowing direction A, the discharge of the condensation water is hardly disturbed by the protruded fin (the fin section 17) when the condensation water drops down in the longitudinal direction of the rib of the rib section 14. Here, “generally parallel to the longitudinal direction of the rib section 14” means that the inclination angle θ is in a range from 60 to 120°.

Fifth Embodiment

While a plurality of fin sections 17 constituted by slit fins are linearly arranged parallel to the air flowing direction A in the first embodiment, a plurality of fin sections 17 constituted by slit fins are arranged in a zigzag manner relative to the air flowing direction A as shown in FIG. 14 according to a fifth embodiment. Here, the zigzag arrangement means that the plurality of fin sections 17 are arranged while being shifted with respect to each other in the direction orthogonal to the air flowing direction A.
In this regard, when the fin section 17 is constituted by the protruded fin as in the fifth embodiment, the fin sections 17 may be arranged in a zigzag manner.

Sixth Embodiment

According to the first embodiment, coolant passages indicated by arrows a to k are arranged in series between the coolant inlet pipe 24 a and the coolant exit pipe 24 b as shown in FIG. 2. Contrarily, in a sixth embodiment, two coolant passages are arranged in parallel between the coolant inlet pipe 24 a and the coolant exit pipe 24 b.
The sixth embodiment will be explained with reference to FIGS. 15 and 16, wherein FIG. 15 is an exploded perspective view corresponding to FIG. 2, and FIG. 16 is a schematic perspective view illustrating the coolant passages in the FIG. 15.
According to the sixth embodiment, a tank section 20 located on the downstream side of the air flow and a tank section 22 located on the upstream side of the air flow are formed at an upper end of the heat transfer plate 12 as in the first embodiment. Contrarily, at a lower end of the heat transfer plate 12, a tank section divided into three tank sections is provided; that is, two tank sections 21 a and 21 b located on the downstream side of the air flow and one tank section 23 located on the upstream side of the air flow is provided.
Note that, at the lower end of the leftmost heat transfer plate 12 adjacent to the left side end plate 24 having the coolant exit pipe 24 b and the coolant inlet pipe 24 a, the tank section 21 a is solely provided and the tank section 21 b is not provided on the downstream side of the air flow. A barrier wall for interrupting the coolant passage (a blind lid structure having no communicating opening) is provided at the position at which the tank section 21 b is not formed.
The coolant inlet pipe 24 a in the left end plate 24 communicates with a flow passage of the tank section 20 at the upper end of the heat transfer plate 12 on the downstream side of the air flow. In the flow passage of this tank section 20, as no partition is arranged at an intermediate position in the stacking direction of the heat transfer plates 12 (a boundary between the left side area X and the right side area Y), the flow passage of the tank section 20 passes throughout the length thereof in the stacking direction of the heat transfer plates 12 (in the leftward/rightward direction).
Accordingly, the coolant entering from the coolant inlet pipe 24 a flows through the passage of the tank sections 20 along a total length in the stacking direction of the heat transfer plates 12. The coolant descends the air flow downstream side coolant passage 15 of the heat transfer plate 12 as indicated by arrows n1 and n2. Here, the arrow n1 indicates the coolant descending the coolant passage 15 located in the left side area X, and the arrow n2 indicates the coolant descending the coolant passage 15 located in the right side area Y.
The heat transfer plate 12 is constituted so that coolant passage 15 in the left side area X communicates solely with the air flow downstream side tank section 21 b at the lower end of the heat transfer plate 12, and the coolant passage 15 in the right side area Y communicates solely with the air flow downstream side tank section 21 a.
The flow passage in the tank section 21 a communicates with the left end flow passage of the air flow upstream side lower tank section 23 via the communication passage 24 c formed in the vicinity of the lower end of the left side end plate 24.
In this flow passage of the lower tank section 23, a partition (not shown) is disposed at an intermediate position in the stacking direction of the heat transfer plates 12 (the boundary between the left side area X and the right side area Y) to divide the flow passages in the left side area X and the right side area Y. Accordingly, the communication passage 24 c communicates solely with the flow passage in the left side area X of the lower tank section 23.
On the other hand, the flow passage in the tank section 21 b communicates with the right end low passage of the air flow upstream side lower tank section 23 via the communication passage 25 a formed in the vicinity of the lower end of the right end plate 25. That is, the communication passage 25 a communicates solely with the flow passage in the right side area Y among the flow passages in the lower tank section 23. The coolant, descending as indicated by the arrow n1, flows rightward through the lower tank section 21 b as indicated by an arrow p1, and then flows into the right side flow passage of the air flow upstream side lower tank section 23 via the communication passage 25 a of the right end plate 25 as indicated by an arrow q1.
The coolant in the right side flow passage of the lower tank section 23 rises in the air flow upstream side coolant passage 16 in the right side area Y as indicated by an arrow r1, and flows into the right side flow passage of the air flow upstream side upper tank 21.
On the other hand, the coolant descending the coolant passage 15 in the right side area Y located on the air flow downstream side, as indicated by an arrow n2, flows leftward in the lower tank section 21 a as indicated by an arrow p2, and then flows into the left side flow passage of the air flow upstream side lower tank section 23 via the communication passage 24 c of the left end plate 24 as indicated by an arrow q2.
The coolant in the left side flow passage of the lower tank section 23 rises the air flow upstream side coolant passage 16 in the left side area Y as indicated by an arrow r2, and flows into the left side flow passage of the air flow upstream side upper tank 21.
The coolant coming from the coolant passage 16 in the right side area Y and the coolant coming from the coolant passage 16 in the left side area X join together in the upper tank 21 and flow toward the coolant exit pipe 24 b as indicated by an arrow s.
Thereby, between the air flow downstream side upper tank section 20 communicating with the coolant inlet pipe 24 a and the air flow upstream side upper tank section 21 communicating with the coolant exit pipe 24 b, a first coolant passage indicated by the arrows n1, p1, q1 and r1 and a second coolant passage indicated by the arrows n2, p2, q2 and r2 are arranged in parallel to each other.
In this regard, according to the inventive plate type evaporator 10, the fin sections 17 are arranged between the adjacent rib sections 14. Thus, when the space pitch Sp is enlarged for the purpose of restricting an increase in ventilation resistance caused by the arrangement of the fin sections 17, the number of heat transfer plates 12 to be stacked together decreases.
The reduction of the number of heat transfer plates 12 causes the reduction of the coolant passage area, which increases the pressure loss of the coolant passage in the evaporator 10. The increase in pressure loss of the coolant flow passage causes the rise of the coolant evaporation temperature, whereby the cooling performance of the evaporator 10 becomes worse.
In the first embodiment, as the coolant inlet pipe 24 a and the coolant exit pipe 24 b are coupled to each other by a single coolant passage arranged in series indicated by the arrows a to k, the above-mentioned increase in pressure loss is liable to occur in the coolant passage.
Contrarily, in the coolant passage structure of the sixth embodiment, as the first coolant passage and the second coolant passage are coupled in parallel to each other in the evaporator 10, it is possible to effectively suppress the increase in the pressure loss in the evaporator 10.
By coupling the first and second coolant passages in an X pattern between the air flow downstream side upper tank section 20 and the air flow upstream side lower tank section 21, it is possible to make the distribution of the air temperature blown out from the evaporator uniform.

Seventh Embodiment

As shown in FIG. 4, according to the first embodiment, the fin section 17 is a slit fin having a U-shape, but the slit fin is not limited to have such a U-shape. The seventh embodiment relates to another shape of the slit fin constituting the fin section 17. As shown in FIG. 17, the slit fin constituting the fin section 17 is protruded to have a smoothly curved surface (a dome-like contour).
According to the curved surface (the dome-like contour) of the slit fin shown in FIG. 17, the offset wall surface 17 a and the left and right side walls 17 b and 17 c are continuously coupled by a smooth curve.

Eighth Embodiment

The width dimension Fw of the fin section 17 constituted by the slit fin is sufficiently smaller than the rib pitch Rp; in other words, a width dimension of a flat surface of the base plate section 13; in the first and second embodiments as shown in FIGS. 3 and 9. Contrarily, in the eighth embodiment, as shown in FIG. 18, a width dimension Fw of the fin section 17 constituted by the slit fin is sufficiently larger than in the first embodiment.
In the eighth embodiment, similarly to the second embodiment, the rib sections 14 in the two heat transfer plates 12 are arranged at the same position in the air flowing direction A. In this structure, the fin section (slit fin) 17 is formed to have a fin width dimension Fw nearly equal to a width dimension (a dimension of the flat surface in the air flowing direction) of the flat surface of the base plate section 13 located between the rib sections 14.
Concrete dimensions in the eighth embodiment are as follows; the space pitch Sp (=Tp−2t): 3.0 mm; the thickness t of the heat transfer plate 12: 0.15 mm; the rib section pitch Rp: 7.1 mm; the height Rh of the rib section: 1.45 mm; the fin pitch Fp=the rib section pitch Rp; the fin width Fw: 4.0 mm; and the fin height Fh: 1.0 mm.
According to the eighth embodiment, it is possible to increase the heat transfer area because the fin width Fw is enlarged, from (0.8 mm×2) in the first embodiment, to 4.0 mm.

Ninth Embodiment

In a ninth embodiment, the space pitch Sp which is a mutual distance between the base plate sections 13 (that is, flat surface portions) of the heat transfer plates 12 adjacent to each other while interposing the air passage 18, is studied.
As shown in FIG. 19, in the ninth embodiment, the rib height Rha of the rib section 14 is equal to a height from a surface of the base plate section 13 of the heat transfer plate 12 (that is, a height projected into the air passage 18). Accordingly, the projected height Rha is a value obtained by subtracting the thickness t of the heat transfer plate 12 from the rib height Rh in FIGS. 6 and 18 (Rha=Rh−t).
If the space pitch Sp becomes larger, a gap G between the rib sections 14 projected into the air passage 18 increases, whereby the action of the rib section 14 for guiding the air stream becomes insufficient and, finally, the air stream linearly flows through the air passage 18.
The present inventors have specifically studied the relationship between the space pitch Sp and the projected height Rha, and found that the meandering stream A1 is certainly formed by determining the space pitch Sp to be three times the projected height Rha or less, that is, the space pitch Sp≦3×Rha. Thereby, it has been confirmed that the dwelling region F of the air stream along the surface of the base plate section 13 (see FIG. 27) can be eliminated.
In this regard, as the pressure loss in the air stream increases if the space pitch Sp is extremely small, the space pitch Sp must be larger by a predetermined amount than the rib height Rha of the rib section 14. Preferably, the space pitch Sp is selected within a range from Sp=(2.0 to 2.3)×Rha for the purpose of forming the meandering air stream as well as reducing the pressure loss of the air stream.

Tenth Embodiment

A tenth embodiment relates to the projected height Fha of the fin section 17 when the fin section 17 is constituted by the slit fin.
As shown in FIG. 20, according to the tenth embodiment, a projected height Fha of the fin section 17 is equal to a height from the base plate section 13 of the heat transfer plate 12 (that is, a projected height into the air passage 18). More concretely, the projected height Fha is equal to a distance between the surface of the base plate section 13 of the heat transfer plate 12 and a center of a thickness of the offset wall surface 17 a. Therefore, the projected height Fha of the fin section 17 is a value obtained by subtracting a thickness t of the heat transfer plate 12 and half a thickness t′ of the offset wall surface 17 a from the fin height Fh, that is Fha=Fh−t−0.5t′.
On the other hand, an axis H of a plate is parallel to the base plate section 13 of the heat transfer plate 12 (see FIG. 20). A perpendicular line I orthogonal to the axis H of the plate is drawn. A length of a line is defined as L, which connects intersecting points J1 and J2 to each other, of the perpendicular line I on the surfaces of the two heat transfer plates 12 opposed to each other while interposing the air passage 18. The projected height Fha of the fin section 17 is determined to be half the length L or less at a position at which the fin section 17 is formed. That is, Fha≦0.5×L.
In FIG. 20, the fin section 17 at a position a is disposed to oppose the base plate section 13 in the adjacent heat transfer plate 12, the fin section 17 at a position b is disposed to oppose the front (top) of the rib section 14 in the adjacent heat transfer plate 12, and the fin section 17 at a position c is disposed to oppose an intermediate height portion between a top and a root of the curved surface in the adjacent heat transfer plate 12.
Accordingly, the lengths defined as described above have the relationship of L1>L3>L2. In either of the fin section 17 in a, b or c, the projected height Fha1, Fha2 or Fha3 is half a line length L1, L2 or L3 or less.
That is, the following relationship is established; Fha1≦0.5×L1, Fha2≦0.5×L2, and Fha3≦0.5×L3.
The line lengths L1, L2 and L3 are plate gaps determining the cross-sectional area of the passage variously changing in accordance with the directions A of air stream in the air passage 18 formed between the adjacent two heat transfer plates 12.
Thus, by setting the projected heights Fha1, Fha2 and Fha3 of the fin section 17 as described above, even if the position of the fin section 17 varies to the position a, b or c, a position of a center of a plate thickness of the offset wall surface 17 a in the fin section 17 is always located closer to the base plate section 13 (the base plate section 13 n which the fin section 17 is formed) than to the center of the above-mentioned “plate gap determining the cross-sectional area of the air passage”.
As the offset wall surface 17 a in the fin section 17 is located in the air stream in the air passage 18 and extends parallel to the flat surface of the base plate section 13 (parallel to the air flowing direction A), the air is liable to flow along the offset wall surface 17 a.
Therefore, as the offset wall surface 17 a is located closer to the base plate section 13 than to the center of the “plate gap determining the cross-sectional area of the air passage”, it is possible to cause the air stream flowing along the offset wall surface 17 a to approach the base plate section 13. As a result, as shown in FIG. 21, it is possible to assuredly form the air stream A1 largely meandering closer to the base plate section 17 than to the top of the curved surface of the rib section 14. Thus, the dwelling region F (see FIGS. 22 and 27) of the air stream flowing along the surface of the base plate section 13 is eliminated.
Contrarily, when the projected height Fha of the fin section 17 is too high, that is, when the projected height Fha of the fin section 17 is larger than the above-mentioned line length L, the offset wall surface 17 a of the fin section 17 is closer to the top of the opposed rib section 14 as shown in FIG. 22, whereby the air stream flowing along the offset wall surface 17 a is away from the base plate section 13 and, instead, approaches the top of the rib section 14.
In other words, according to the comparative example shown in FIG. 22, the offset wall surface 17 a disturbs the formation of the meandering stream to be established by the rib section 14. As a result, the air stream becomes almost linear as indicated by an arrow A2, causing the dwelling region F of the air stream along the surface of the base plate section 13, and the extreme reduction of the heat transfer rate on the surface of the base plate section 13.
In this regard, if the projected height Fha of the fin section 17 is extremely small, it is difficult to pass air through the interior of the offset wall surface 17 a, whereby it is necessary that the projected height Fha of the fin section 17 is a predetermined height or more capable of ensuring the air stream within the interior of the offset wall surface 17 a.
According to the tenth embodiment, while the projected height Fha of the fin section 17 is set to be half a line length L or less at a position at which the fin section 17 is formed; that is, Fha≦0.5×L, it is inevitable that the projected height Fha of the fin section 17 has the production variance (machining tolerance) when the heat exchanger is manufactured. Concretely, the machining tolerance is usually approximately ±17%, and if the projected height Fha is, for example, 3 mm or less, the projected height Fha of the fin section 17 has approximately ±0.5 mm.
Accordingly, “to suppress the projected height Fha of the fin section 17 to a value half a line length L or less at the position at which the fin section 17 is formed” does not strictly mean that the height Fha must be half a line length L or less, but means that it is generally half a line length or less including the excess amount due to the above-mentioned machining tolerance.

Eleventh Embodiment

In the above-mentioned embodiments, the offset wall surface 17 a of the fin section 17 is formed to be parallel to the flat surface of the base plate section 13. Contrarily, in an eleventh embodiment, the offset wall surface 17 a of the fin section 17 is inclined to the flat surface of the base plate section 13.
As shown in FIG. 23, according to the eleventh embodiment, when the fin sections 17 are located adjacent to each other both on windward and leeward sides of the rib section 14, the offset wall surface 17 a of the fin section 17 is inclined in the same direction as the curved surface of the closest rib section 14 on the same heat transfer plate 12.
That is, the offset wall surface 17 a of the fin section 17 located on the windward side of the rib section 14 is inclined to be away from the flat surface of the base plate section 13 as going from the upstream to the downstream. Contrarily, the offset wall surface 17 a of the fin section 17 located on the leeward side of the rib section 14 is inclined to be closer to the flat surface of the base plate section 13 as going from the upstream to the downstream.
Thereby, the offset wall surface 17 a of the fin section 17 performs the operation for facilitating the guiding of the air stream due to the curved surface of the rib section 14 (the guiding operation for the meandering stream). As a result, as shown in FIG. 23, the meandering stream A3 is assuredly formed, whereby the region F in which the air stream dwells along the surface of the base plate section 13 (see FIGS. 22 and 27 is eliminated.

Twelfth Embodiment

The fin sections 17 are arranged adjacent to the rib section 14 both on the windward and leeward sides in the eleventh embodiment. Contrarily, as shown in FIG. 24, according to a twelfth embodiment, the fin section 17 is disposed adjacent to the rib section 14 solely on the leeward side, so that the offset wall surface 17 a of the fin section 17 is inclined in the same direction as the leeward side curved surface of the rib section 14. That is, the offset wall surface 17 a is inclined closer to the flat surface of the base plate section 13 from upstream to downstream.
FIG. 25A illustrates a comparative example wherein the fin sections 17 are not disposed both on the windward and leeward sides of the rib section 14. As the air stream indicated by an arrow K impinges to the windward curved surface of the rib section 14, the heat transfer rate becomes high. However, vortices are generated on the leeward side of the rib section 14 as indicated by an arrow M due to the impingement on the windward side indicated by the arrow K to result in the dwelling of the air stream.
As a result, the heat transfer rate on the leeward side curved surface of the rib section 14, disposed in the region in which the vortices M are generated, becomes extremely low. Similarly, also in the base plate section 13, the heat transfer rate extremely deteriorates in the region in which the vortices M are generated. In this regard, O in FIG. 25A denotes a position at which the air stream impinges again to the base plate section 13. In a portion of the base plate section 13 upstream from the position O, the heat transfer rate is low.
Contrarily, in the twelfth embodiment, as shown in FIG. 25B, it is possible to cause the air stream P passing through the inside of the offset wall surface 17 a of the fin section 17 to flow closer to the leeward side curved surface of the rib section 4.
Thus, as the region in which the vortices M′ generate (the air stream dwelling region) can be minimized to a great extent in comparison with the region in which the vortices M are generated, it is possible to largely improve the heat transfer rate of the leeward side curved surface of the rib section 14 and the base plate section 13.
In this regard, in the above-mentioned eleventh and twelfth embodiments, when the offset wall surface 17 a of the fin section 17 is disposed directly adjacent to the rib section 14, the offset wall surface 17 a of the fin section 17 is inclined in the same direction as in the curved surface of the closest rib section 14. However, as shown in FIGS. 19 and 21, when the fin section 17 is disposed at a center of the base plate section 17 as seen in the air flowing direction and the fin section 17 is disposed opposite to a front of the rib section 14 in the opposite side heat transfer plate 12, it is better to form the offset wall surface 17 a of the fin section 17 parallel to the base plate section 13, instead of inclining the same.
That is, if the fin section 17 is disposed opposite to a front of the rib section 14 of the opposite side heat transfer plate 12, the offset wall surface 17 a is just located at a position at which the air stream is reversed. Accordingly, if the offset wall surface 17 a is inclined, air flows along the inclination thereof to disturb the reverse of the air stream. Contrarily, if the offset wall surface 17 a is formed parallel to the base plate section 13, the offset wall surface is neutral against the air stream, whereby the offset wall surface 17 a does not guide the air stream to a particularly inclined direction. Thus, as shown in FIGS. 19 and 21, the air stream is smoothly reversed and the meandering stream A1 is suitably formed.

Thirteenth Embodiment

A thirteenth embodiment relates to the determination of dimensions between the offset wall surface 17 a of the fin section 17 and the surface of the heat transfer plate 12.
When the heat exchanger is a cooler type generating condensation water as air is being cool, such as the air conditioner evaporator 10 shown in FIG. 1, the drainage of condensation water is an important problem in the design of the heat exchanger.
According to the thirteen embodiment, gaps Q1, Q2 are determined to be 0.3 mm or more between the offset wall surface 17 a of the fin section 17 and the surface of the heat transfer plate 12; more concretely, a gap Q1 between the inside surface of the offset wall surface 17 a and the surface of the base plate section 13, and a gap Q2 between the outside surface of the offset wall surface 17 a and the surface of the rib section on the heat transfer plate 12 disposed on the opposite side shown in FIG. 26.
According to a study made by the present inventors, it has been confirmed that by setting the above-mentioned gaps Q1 and Q2 to 0.3 mm or more (Q1, Q2≧0.3 mm), the condensation water does not block these gaps Q1 and Q2 but is smoothly drained.
In the cooler type heat exchanger generating condensation water, the installation posture of the heat exchanger during the use is determined so that the longitudinal direction of the rib section 14 (vertical to a paper surface in FIG. 26) coincides with the direction of gravity (upward/downward direction). Thereby, the condensation water generated on the surface of the heat transfer plate 12 smoothly flows down in the longitudinal direction of the rib section 12.

Other Embodiments

In the above-mentioned embodiments, a description has been made for the cases wherein the coolant passages (internal passages) 15 and 16 are formed inside the rib section 14 by laying and fixing two heat transfer plates 12 completely separated from each other. As disclosed in FIG. 36 of Japanese Unexamined Patent Publication No. 2001-41678, however, the two heat transfer plates 12 and 12 constituting the coolant passages (internal passages) 15, 16 may be formed of a press-formed single plate member which is bent at a widthwise center to be two sections 12, 12, and thereafter base plate sections 13, 13 thereof are fixed together to form the coolant passages 15 and 16.
Further, lateral surfaces of the respective plate members constituting the above-mentioned two heat transfer plates 12, 12 may be coupled together with a cleat-like coupler. This coupler is designed to have the same length as the space pitch Sp. Such a coupling structure is also disclosed in FIG. 36 of Japanese Unexamined Patent Publication No. 2001-41678.
As understood from such modifications, “two heat transfer plates 12 are used as one pair” in the present invention includes both of a case wherein the completely separated two heat transfer plates 12 are stacked together and another case wherein a single plate member 120 is bent at a center 121 and two portions of half a size are laid together.
In the above embodiments, while the description has been made on a case wherein the present invention is applied to an evaporator 10 which is a heat-suction side heat exchanger for a refrigeration cycle, the present invention may be applicable to heat exchangers for various uses.
For example, the present invention may be applicable to a condenser which is a heat-radiation side heat exchanger for a refrigeration cycle. Also, the present invention may be applicable to a heat exchanger wherein hot water flows through the internal passage of the heat transfer plate 12 (the coolant passages 15 and 16 in the above-mentioned embodiments) such as a hot water type radiator for a heater or a radiator for cooling an engine.
Similarly, the present invention may be applicable to a heat exchanger such as an engine oil cooler wherein oil flows through internal passages or an heat exchanger wherein cold water flows through internal passages.
While the invention has been described by reference to specific embodiments chosen for purposes of illustration, it should be apparent that numerous modifications could be made thereto by those skilled in the art without departing from the basic concept and scope of the invention.

Claims

1. A heat exchanger wherein

a plurality of heat transfer plates forming plate surfaces extending in the flowing direction (A) of external fluid are stacked together,

a gap is provided between said plate surfaces of said adjacent heat transfer plates to form an external passage through which said external fluid flows,

a plurality of rib sections extending orthogonal to the flowing direction A of said external fluid are projected from said plate surfaces into said external passage to be integral with said heat transfer plates,

by shifting positions of the plurality of rib sections in one of said adjacent heat transfer plates relative to positions of the plurality of rib sections in the other of said adjacent heat transfer plates as seen in the flowing direction A of said external fluid, said external passage is formed in a meandering manner,

the plurality of rib sections form an internal passage inside thereof, through which flows internal fluid,

fin sections are projected from said plate surfaces at positions between the adjacent rib sections to be integral with the heat transfer plate, and

said fin section is press-formed so as to protrude a cut portion partially cut a plate thickness of said heat transfer plate.

2. A heat exchanger as defined by claim 1, wherein said heat transfer plates are combined to form pairs, and said rib sections and said fin sections are formed integral with said pair of heat transfer plates, and

the pair of heat transfer plates are fixed together to form said internal passage inside the plurality of rib sections.

3. A heat exchanger as defined by claim 2, wherein positions in the pair of heat transfer plates at which said rib sections are formed are shifted in the flow direction A of the external fluid, and

said internal passage is formed by said rib sections formed in one of the pair of heat transfer plates and a plate surface of the other.

4. A heat exchanger as defined by claim 2, wherein said rib sections are formed in said pair of heat transfer plates at the same positions as seen in the flowing direction A of said external fluid, and

said internal passages are formed by the combination of said rib sections formed in said pair of heat transfer plates, respectively.

5. A heat exchanger as defined by claim 1, wherein said heat transfer plate is constituted by a single extrusion-formed plate material,

said rib sections are formed by extrusion-forming a tubular-shaped portion on said single extrusion-formed plate material, and

said fin sections are formed integral with said single extrusion-formed plate material to be projected from a plate surface of said single extrusion-formed plate material.

6. A heat exchanger as defined by claim 1, wherein said heat transfer plate has a base plate section having a flat surface between the adjacent rib sections, and

said fin section is formed in said base plate section.

7. A heat exchanger as defined by claim 1, wherein a width (Fw) in the flowing direction (A) of said external fluid of said fin section is 5 mm or less.

8. A heat exchanger as defined by claim 1, wherein said fin section is a slit fin having an offset wall surface apart from a plate surface of said heat transfer plate at a predetermined gap, wherein said offset wall surface are coupled to a plate surface of said heat transfer plate at two positions.

9. A heat exchanger as defined by claim 8 wherein, when a gap between positions on the pair of heat transfer plates opposed to each other to define said external passage, at which positions are formed said slit fins, is defined as L, and a projected height of said offset wall surface from a plate surface of said heat transfer plate is defined as Fha, the following relationship is satisfied:

Fha≦½L.

10. A heat exchanger as defined by claim 8, wherein a cross-sectional shape of said rib section has a curved surface projected from the surface of said heat transfer plate, which is generally semicircular,

said slit fin is located at a position directly on downstream side from said external fluid relative to said rib section, and

said offset wall surface is inclined in the same direction as the inclination of the downstream side curved surface in the generally semicircular curved surface of said rib section.

11. A heat exchanger as defined by claim 8, wherein the cross-sectional shape of said rib section is such that it has a curved surface protruded semi-circularly from a surface of said heat transfer plate,

said slit fin is disposed adjacent to said rib section at a position directly on the upstream side of said external fluid, and

said offset wall surface is inclined in the same direction as the inclination of the upstream side curved surface in a generally semicircular curved surface of said rib section.

12. A heat exchanger as defined by claim 8, wherein said slit fin is disposed opposite to a front of said rib section while interposing said external passage, and said offset wall surface is formed to be parallel to a flat surface of said heat transfer plate.

13. A heat exchanger as defined by claim 8, wherein said external fluid is air and said internal fluid is a coolant for cooling said air, wherein said heat exchanger is constituted as a cooling heat exchanger generating condensation water on the surface of said heat transfer plate, and

a gap (Q1, Q2) between said offset wall surface and the surface of said heat transfer plate is 0.3 mm or more.

14. A heat exchanger as defined by claim 1, wherein said fin section is a protruded fin having a predetermined angle relative to the surface of said heat transfer plate.

15. A heat exchanger as defined by claim 14, wherein said protruded fin is triangular.

16. A heat exchanger as defined by claim 15, wherein said triangular protruded fin is inclined to the flowing direction (A) of said external fluid at an angle from 15° to 45°.

17. A heat exchanger as defined by claim 14, wherein said protruded fin is rectangular.

18. A heat exchanger as defined by claim 14, wherein the inclination angle of said protruded fin relative to the flowing direction (A) of said external fluid is determined in a small angle range from −30° to +30° so that a surface of the protruded fin follows the flow direction (A) of said external fluid.

19. A heat exchanger as defined by claim 14, wherein said external fluid is air and internal fluid for cooling said air flows through said internal passage,

said heat transfer plate is disposed so that the longitudinal direction of said rib section coincides with the upward/downward direction, and

an inclination angle of said protruded fin is in a range from 60° to 120° relative to the flowing direction (A) of said external fluid so that a surface of said protruded fin follows the longitudinal direction of said rib section.

20. A heat exchanger as defined by claim 1, wherein said internal passage has an upstream side internal passage disposed on the upstream side in the flow direction (A) of said external fluid and a downstream side internal passage disposed on the downstream side in the flowing direction (A) of said external fluid,

said upstream side internal passage and said downstream side internal passage are respectively sectioned vertically to the flowing direction (A) of said external fluid into a plurality of areas (X, Y), and

passages connected in parallel to each other are constituted between the plurality of areas (X, Y) of said upstream side internal passages and the plurality of areas (X, Y) of said downstream side internal passages.

21. A heat exchanger as defined by claim 20, wherein said downstream side internal passage is an inlet side passage for said internal fluid, and said upstream side internal passage is an exit side passage for said internal fluid.

22. A heat exchanger as defined by claim 20, wherein said parallel passages couple the plurality of areas (X, Y) in said upstream side internal passage and the plurality of areas (X, Y) in said downstream side internal passage to each other in an X pattern.