PLATE-TYPE HEAT EXCHANGER AND REFRIGERATION CYCLE APPARATUS USING THE SAME

Abstract

A plate thicknesses t of each of heat transfer plates 2 and 3 is equal to or less than 0.2 mm; a pitch Λ of concaves/convexes 9 and 10 is 4 to 7 mm; a distance h between peaks of the concaves/convexes 9 and 10 is 1.0 to 1.2 mm; when a value obtained by dividing a wave length s corresponding to the length of each of the heat transfer plates 9 and 10 between peaks of the concaves/convexes 9 and 10 of the heat transfer plates 2 and 3 by the pitch Λ of each of the concaves/convexes 9 and 10 is defined as an area enlargement ratio Φ, the area enlargement ratio Φ is 1.05 to 1.15.

Description

TECHNICAL FIELD

The present invention relates to a plate-type heat exchanger and a refrigeration cycle apparatus using the same.

BACKGROUND ART

A plate-type heat exchanger has been proposed in which a plurality of heat transfer plates having a plurality of rows of wave-shaped concaves and convexes formed thereon are stacked and arranged such that a line connecting peaks of crest portions (or bottoms of trough portions) of the wave shape of each heat transfer plate intersects with an adjacent heat transfer plate. In such a plate-type heat exchanger, a wave height h corresponding to the interval between the peak of the crest portion of the wave shape of the heat transfer plate and the bottom of the trough portion of a heat transfer plate adjacent to this heat transfer plate is set at, for example, about 1.6 mm to 2.2 mm.

When such a wave height h is set, since the cross-section of a refrigerant flow path is large, the flow rate of the refrigerant is decreased. Thus, the heat transfer efficiency between refrigerants is reduced. Therefore, in order to restrain a reduction in heat transfer efficiency, a plate-type heat exchanger has been proposed in which the wave height h of each heat transfer plate is set at 0.5 to 1.5 mm (the hydraulic diameter is set at 1 to 3 m) (see, e.g., Patent Literature 1).

However, when the flow rate of the refrigerant increases, the refrigerant more easily flows beyond the wave shape formed in each heat transfer plate, and there is a possibility that the refrigerant flows in the longitudinal direction of each heat transfer plate. In other words, the refrigerant is less likely to spread in the short-side direction (widthwise direction) of each heat transfer plate, and then there is a possibility that the flow rate of the refrigerant in the short-side direction becomes nonuniform. This interrupts flow of refrigerant on the width side of the heat transfer plate, and then there is a possibility that the effective heat transfer area reduces or dust clogging occurs.

To overcome the above-described circumstances, a plate-type heat exchanger has been proposed in which a wave angle θ formed between the longitudinal direction of each heat transfer plate and a line connecting the peaks of the crest portions of the wave shape is set small (e.g., 45 degrees) so that the refrigerant easily spreads in the short-side direction of the heat transfer plate (see, e.g., Patent Literature 2).

It should be noted that in the technique described in Patent Literature 2, a wave pitch Λ which is the interval between the crests (troughs) of each heat transfer plate is set low (e.g., 4 mm or lower) so that the flow rate of the refrigerant increases.

CITATION LIST

Patent Literature

• Patent Literature 1: Japanese Unexamined Patent Application Publication No. 2001-56192 (e.g., paragraph [0017] of the specification and FIG. 2)
• Patent Literature 2: Japanese Unexamined Patent Application Publication No. 2011-516815 (e.g., paragraphs [0025] to [0028] of the specification)

SUMMARY OF INVENTION

Technical Problem

In the technique described in Patent Literature 1, the wave height h is reduced, whereby the cross section of the refrigerant flow path is decreased and the flow rate of the refrigerant is increased. In this case, when the flow rate of the refrigerant increases, the refrigerant is likely to be agitated at an intersection of adjacent heat transfer plates, and pressure loss is increased. This poses a problem that the power consumption of a compressor which supplies the refrigerant to the plate-type heat exchanger increases.

In addition, the refrigerant is less likely to spread in the short-side direction (widthwise direction) of each heat transfer plate, and then there is a possibility that flow of refrigerant in the short-side direction becomes nonuniform. This interrupts flow of refrigerant on the width side of the heat transfer plate, and then there is a possibility that the effective heat transfer area reduces or dust clogging occurs.

In the technique described in Patent Literature 2, since the wave angle θ is set at 45 degrees, it is possible to suppress the occurrence of a phenomenon that “flow of refrigerant in the short-side direction becomes nonuniform” as in the technique described in Patent Literature 1. However, since the wave pitch Λ is set equal to or lower than 4 mm, when the wave angle θ is thus set, the distance in the short-side direction between joining points decreases. Thus, when brazing the heat transfer plates, the joining points are filled with a brazing material, and then there is a possibility that the refrigerant flow path is blocked (pressure loss is increased).

Moreover, in the technique described in Patent Literature 2, since the wave pitch Λ is reduced, an area enlargement ratio Φ is increased. The increase in area enlargement ratio Φ amounts to an increase in elongation of a plate material when forming each heat transfer plate from the plate material. When the area enlargement ratio Φ increases, there is a possibility that the heat transfer plate suffers from, for example, cracks or nonuniformity in thickness t.

In other words, in the technique described in Patent Literature 2, there is a possibility that the plate-type heat exchanger suffers loss of strength. Thus, the thickness of the plate material cannot be decreased, and this increases the material cost and the weight. In addition, since the thickness of the plate material cannot be decreased, the set load in press working increases, and the cost incurred in this process, in turn, increases.

The present invention has been made in order to solve at least one of the above-described problems, and has as its object to provide a plate-type heat exchanger which increases the heat transfer efficiency, reduces pressure loss, restrains blocking of a refrigerant flow path and an increase in cost, and has a lightweight configuration.

Solution to Problem

A plate-type heat exchanger according to the present invention has an inlet through which a fluid flows therein, an outlet through which the fluid having flowed therein through the inlet flows out therefrom, and a flow path connecting the inlet to the outlet and formed in a space which is formed by waves of concaves/convexes of adjacent heat transfer plates among a plurality of heat transfer plates having a plurality of substantially V-shaped waves of concaves/convexes formed and aligned from the inlet toward the outlet, the plurality of heat transfer plates being stacked in alternate layers of inverted heat transfer plates and noninverted heat transfer plates. A plate thickness t of each heat transfer plate is equal to or less than 0.2 mm. A pitch Λ of the concaves/convexes is 4 to 7 mm. A distance h between peaks of the concaves/convexes is 1.0 to 1.2 mm. When a value obtained by dividing a wave length s corresponding to a length of the heat transfer plate between peaks of the wave of the concaves/convexes of the heat transfer plate by the pitch Λ of the concaves/convexes is defined as an area enlargement ratio Φ, the area enlargement ratio Φ is 1.05 to 1.15.

Advantageous Effects of Invention

In the plate-type heat exchanger according to the present invention, since the plate thickness t is equal to or less than 0.2 mm, the pitch Λ of the concaves/convexes is 4 mm to 7 mm, the distance h between the peaks of the concaves/convexes corresponding to the stacking direction is 1.0 mm to 1.2 mm, the area enlargement ratio Φ falls within the range of 1.05 to 1.15, it is possible to increase the heat transfer efficiency, to reduce pressure loss, to restrain blocking of a refrigerant flow path and an increase in cost, and to provide a lightweight configuration.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a diagram illustrating a plate-type heat exchanger according to Embodiment 1 of the present invention.

FIG. 2 is a schematic view of a heat transfer plate of the plate-type heat exchanger shown in FIG. 1.

FIG. 3 is a diagram illustrating various dimensions of the heat transfer plate of the plate-type heat exchanger shown in FIG. 1.

FIG. 4 is a diagram illustrating the relationship between a wave pitch Λ and an area enlargement ratio Φ of the plate-type heat exchanger shown in FIG. 1.

FIG. 5 is a graph showing the amount of weight reduction of the plate-type heat exchanger when a wave height h and a wave angle θ are changed as parameters with the amount of heat exchange of the plate-type heat exchanger being set at 15 kW.

FIG. 6 is a graph showing the amount of weight reduction of the plate-type heat exchanger when the area enlargement ratio Φ and the wave angle θ are changed as parameters.

FIG. 7 is a diagram illustrating a distance between joining points of adjacent heat transfer plates for each wave angle θ.

FIG. 8 is a diagram illustrating a refrigeration cycle apparatus according to Embodiment 2 of the present invention.

DESCRIPTION OF EMBODIMENTS

Embodiments of the present invention will be described hereinafter with reference to the drawings.

Embodiment 1

FIG. 1 is a diagram illustrating a plate-type heat exchanger 100 according to Embodiment 1. Here, FIG. 1(a) is a side view in a state where the plate-type heat exchanger 100 is assembled, FIG. 1(b) is a front view of a side plate 1. FIG. 1(c) is a front view of a heat transfer plate 2. FIG. 1(d) is a front view of a heat transfer plate 3. FIG. 1(e) is a front view of a side plate 4. FIG. 1(f) is a diagram illustrating a state where the heat transfer plate 2 and the heat transfer plate 3 are stacked on each other. FIG. 2 is a schematic view of a heat transfer plate 20 of the plate-type heat exchanger 100 shown in FIG. 1. It should be noted that in FIG. 2, each solid arrow represents flow of a first refrigerant, and each dotted arrow represents flow of a second refrigerant.

It should be noted that the relationship of size between individual constituent elements in the drawings described below, including FIG. 1, may be different from the actual relationship. The configuration of the plate-type heat exchanger 100 will be described with reference to FIGS. 1 and 2.

[Configuration of Plate-Type Heat Exchanger 100]

The configuration of the plate-type heat exchanger 100 will be described first.

The plate-type heat exchanger 100 exchanges heat between a heat source side refrigerant (the first refrigerant) conveyed from an outdoor unit on the heat source side and a heat medium (the second refrigerant) conveyed from an indoor unit on the use side. In other words, the plate-type heat exchanger 100 is a heat exchanger for refrigerant vs. refrigerant or for refrigerant vs. heat medium such as water or brine. It should be noted that in the plate-type heat exchanger 100, a first refrigerant flow path X through which the first refrigerant flows and a second refrigerant flow path Y through which the second refrigerant flows are formed such that the first refrigerant and the second refrigerant do not mix with each other.

As shown in FIG. 1, the plate-type heat exchanger 100 includes the heat transfer plate 20 and the side plates 1 and 4 which reinforce the plate-type heat exchanger 100.

(Heat Transfer Plate 20)

As shown in FIG. 2, the heat transfer plate 20 forms the first refrigerant flow paths X and Y for the first refrigerant and the second refrigerant, respectively. The heat transfer plate 20 includes two types of plates: heat transfer plates 2 and 3 each having a rectangular plane shape. It should be noted that the heat transfer plate 2 is obtained by inverting the heat transfer plate 3.

The heat transfer plate 20 includes heat transfer plates 2 each having substantially V-shaped waves of concaves/convexes formed on its surface and aligned on a plurality of rows in the “vertical direction”; and heat transfer plates 3 arranged in such a manner that the upper and lower sides thereof are opposite to those of the heat transfer plate 2. The heat transfer plates 2 and 3 are alternately arranged (stacked) so as to be opposed to each other. Therefore, the heat transfer plate 3 whose upper and lower sides are opposite to those of the heat transfer plate 2 is arranged behind the heat transfer plate 2, and the heat transfer plate 2 is arranged behind this heat transfer plate 3.

The “vertical direction” refers herein not only to a direction perpendicular to a surface on which the plate-type heat exchanger 100 is installed, but also to the general up-down direction.

In addition, a description will be given on the assumption that the heat transfer plate 20 has a rectangular shape when viewed in a plan view, but its shape is not limited to this and may be, for example, a square or the like. Moreover, the upper and lower sides of the heat transfer plate 20 correspond to the upper and lower sides, respectively, of the sheet surface of FIG. 1. In other words, the upper side of the heat transfer plate 20 corresponds to a side where a first opening 11 and a fourth opening 14 (to be descried later) are provided, and the lower side of the heat transfer plate 20 corresponds to a side where a second opening 13 and a third opening 12 are provided.

Each heat transfer plate 2 is a plate-shaped member that is provided parallel to the heat transfer plate 3 and the side plates 1 and 4, and opposed to an adjacent heat transfer plate 3.

As shown in FIG. 1(c), the heat transfer plate 2 has substantially inverted V-shaped waves of concaves/convexes 9 formed on a plurality of rows, when seen in a plan view. A line connecting peaks (or bottoms) of the concaves/convexes 9 is formed so as to be symmetrical about a center line parallel to the longitudinal direction of the heat transfer plate 2 (the up-down direction of the sheet surface of each of FIGS. 1 and 2). In addition, the concaves/convexes 9 are formed such that the line connecting the peaks of the concaves/convexes 9 forms a predetermined angle with the center line.

Each heat transfer plate 3 has the same configuration as that of each heat transfer plate 2, except for the angle formed between a line connecting peaks (or bottoms) of concaves/convexes 10 of the heat transfer plate 3 and the longitudinal direction of the heat transfer plate 3 (the up-down direction of the sheet surface of each of FIGS. 1 and 2). In other words, each heat transfer plate 3 is a plate-shaped member that is provided parallel to the heat transfer plate 2 and the side plates 1 and 4, and opposed to an adjacent heat transfer plate 2.

As shown in FIG. 1(d), the heat transfer plate 3 has substantially V-shaped waves of concaves/convexes 10 formed on a plurality of rows, when seen in a plan view. A line connecting the peaks (bottoms) of the concaves/convexes 10 is formed so as to be symmetrical about a center line parallel to the longitudinal direction of the heat transfer plate 3. In addition, the concaves/convexes 10 are formed such that the line connecting the peaks of the concaves/convexes 10 forms a predetermined angle with the center line.

The concaves/convexes 9 formed in the heat transfer plates 2 and the concaves/convexes 10 formed in the heat transfer plates 3 are formed in order to, for example, increase the area in which heating energy (or cooling energy) of each of the first refrigerant and the second refrigerant is released, to improve the heat exchange efficiency. Here, each heat transfer plate 2 and each heat transfer plate 3 are provided such that the line connecting the peaks of the concaves/convexes 9 of the heat transfer plate 2 and the line connecting the peaks of the concaves/convexes 10 of the heat transfer plate 3 intersect with each other, when seen in a state where they are opposed to each other as shown in FIG. 1(f). It should be noted that the cross-sections of the concaves/convexes 9 and 10 formed in the heat transfer plates 2 and 3 may have, for example, a saw shape or a wave shape (curved surface). The concaves/convexes 9 and 10 will be described in detail later with reference to FIG. 3.

In each heat transfer plate 2 and each heat transfer plate 3, as shown in FIG. 1(c) and FIG. 1(d), respectively, the first opening 11 through which the first refrigerant flowing into the plate-type heat exchanger 100 flows and the second opening 13 through which the first refrigerant flowing out from the plate-type heat exchanger 100 flows are formed. In addition, in each heat transfer plate 2 and each heat transfer plate 3, the third opening 12 through which the second refrigerant flowing into the plate-type heat exchanger 100 flows and the fourth opening 14 through which the second refrigerant flowing out from the plate-type heat exchanger 100 flows are formed.

In other words, the first opening 11 formed in each heat transfer plate 2 corresponds to an inlet for the first refrigerant flowing into the first refrigerant flow path X formed between the heat transfer plates 2 and 3, and the second opening 13 formed in each heat transfer plate 2 corresponds to an outlet for the first refrigerant having flowed into the first refrigerant flow path X. It should be noted that the second refrigerant passes from the third opening 12 and the fourth opening 14, formed in each heat transfer plate 2, without flowing into the first refrigerant flow path X.

In addition, the third opening 12 formed in each heat transfer plate 3 corresponds to an inlet for the second refrigerant flowing into the second refrigerant flow path Y formed between the heat transfer plates 3 and 2, and the fourth opening 14 formed in each heat transfer plate 3 corresponds to an outlet for the second refrigerant having flowed into the second refrigerant flow path Y. It should be noted that the first refrigerant passes from the first opening 11 and the second opening 13, formed in each heat transfer plate 2, without flowing into the second refrigerant flow path Y.

It should be noted that the first openings 11 in the heat transfer plates 2 and 3 communicate with each other. The same applies to the second opening 13, the third opening 12, and the fourth opening 14.

In addition, as shown in FIG. 2, each heat transfer plate 2 and each heat transfer plate 3 are configured such that the rear surface of the heat transfer plate 2 and the front surface of the heat transfer plate 3 form the first refrigerant flow path X through which the first refrigerant flows, and the rear surface of the heat transfer plate 3 and the front surface of the heat transfer plate 2 form the second refrigerant flow path Y through which the second refrigerant flows. It should be noted that in this example, the “front” corresponds to the “right of the sheet surface” of FIG. 2, and the “rear” corresponds to the “left of the sheet surface” of FIG. 2.

(Side Plates 1 and 4)

The side plates 1 and 4 reinforce the plate-type heat exchanger 100.

The side plate 1 is provided parallel to the heat transfer plate 20 and the side plate 4 and is opposed to the first heat transfer plate 2 from the front, as shown in FIG. 1(a). In addition, the side plate 4 is provided parallel to the heat transfer plate 20 and the side plate 1 and is opposed to the rearmost heat transfer plate 3, as shown in FIG. 1(a).

A first refrigerant inlet pipe 5 for causing the first refrigerant to flow into the plate-type heat exchanger 100 and a first refrigerant outlet pipe 7 for causing the first refrigerant to flow out from the plate-type heat exchanger 100 are provided in the side plate 1. In addition, a second refrigerant inlet pipe 6 for causing the second refrigerant to flow into the plate-type heat exchanger 100 and a second refrigerant outlet pipe 8 for causing the second refrigerant to flow out from the plate-type heat exchanger 100 are provided in the side plate 1.

[Dimensions of Heat Transfer Plate 20]

FIG. 3 is a diagram illustrating various dimensions of the heat transfer plate 20 of the plate-type heat exchanger 100 shown in FIG. 1. It should be noted that FIG. 3(a) is a diagram of the heat transfer plate 2 of the heat transfer plate 20 as seen in a plan view as an example. In addition, FIG. 3(b) is a cross-sectional view on a plane perpendicular to a line connecting the peaks (bottoms) of the concaves/convexes 9 of the heat transfer plate 2 in FIG. 3(a). FIG. 4 is a diagram illustrating the relationship between an area enlargement ratio Φ and a wave pitch Λ of the plate-type heat exchanger 100 shown in FIG. 1. Referring to FIG. 4, a plate thickness t is 0.2 mm, and a wave height h is 1.4 mm.

In Embodiment 1, variables including a wave angle θ, the wave pitch Λ, the wave height h, a wave length s, the area enlargement ratio Φ, and the plate thickness t are used to identify the shapes of the waves of the concaves/convexes 9 and 10 formed in the heat transfer plate 20.

The wave angle θ corresponds to the angle of spread of the waves with respect to the direction in which the substantially V-shaped waves of the concaves/convexes 9 and 10 are aligned. In other words, as shown in FIG. 3, the wave angle θ refers to an angle that a line connecting the peaks (or bottoms) of the concaves/convexes of the heat transfer plate 20 makes with the longitudinal direction of the heat transfer plate 20.

As shown in FIG. 3(b), the wave pitch Λ corresponds to the length between adjacent peaks.

As shown in FIG. 3(b), the wave height h corresponds to the length between the bottom and the peak of the concaves/convexes.

As shown in FIG. 3(b), the wave length s corresponds to the length of the heat transfer plate 20 between adjacent peaks. The wave length s is represented by (Equation 1) shown below. It should be noted that in the following equation R1 is the radius of curvature corresponding to the vertical distance from a center of curvature O shown in FIG. 3(b) to the wave of the concaves/convexes 9 or 10. Also, Θ is the range in which the distance from the center of curvature O to the wave of the concaves/convexes 9 or 10 corresponds to the same radius of curvature R1.

$\begin{array}{cc}s=\pi R1\frac{\Theta }{45}+2\sqrt{\frac{\Theta }{4}+{\left(h-2R1\right)}^2}& \mathrm{Equation}1\end{array}$

The plate thickness t corresponds to the thickness of the heat transfer plate 20.

As shown in FIG. 3(b), the area enlargement ratio Φ is obtained by dividing the wave length s at a predetermined wave height h by the wave pitch Λ. In addition, since the wave length s is represented by the above (Equation 1), the area enlargement ratio Φ can also be represented by the wave height h and the wave pitch Λ. When the area enlargement ratio Φ is low, the elongation of the plate material is small, and when the area enlargement ratio Φ is high, the elongation of the plate material is large. FIG. 4 illustrates the value of the area enlargement ratio Φ when the plate thickness t is 0.2 mm and the wave height h is 1.4 mm.

[Setting of Wave Height h and Wave Angle θ when Amount of Heat Exchange is 15 kW]

FIG. 5 is a graph showing the amount of weight reduction of the plate-type heat exchanger 100 when the wave height h and the wave angle θ are changed as parameters with the amount of heat exchange of the plate-type heat exchanger 100 being set at 15 kW. In FIG. 5, the plate thickness t is 0.2 mm.

As shown in FIG. 5, the wave height h is preferably 1.0 to 1.2 mm. This is because when the wave height h falls within this range, the area enlargement ratio Φ is low and the elongation of the plate material can therefore be kept small. Thus, it is possible to make the weight reduction ratio equal to or higher than 20% or close to 20% by adjusting the wave angle θ.

It should be noted that when the wave height h is 1.0 to 1.2 mm, the wave angle θ is desirably set to fall within the range of 40 degrees to 50 degrees. One reason is that when the wave angle θ falls within this range, while the weight reduction ratio is ensured, it is possible to restrain flow of refrigerant in the short-side direction of the heat transfer plate 20 from being nonuniform. Another reason is that when the wave angle θ falls within this range, it is possible to restrain flow of refrigerant on the width side of the heat transfer plate 20 from being interrupted and then the effective heat transfer area from decreasing or dust clogging from occurring, and thus it is possible to reduce pressure loss.

Furthermore, when the wave height h is 1.0 to 1.2 mm, the elongation of the plate material can be kept small, and thus it is also possible to restrain the heat transfer plate 20 from suffering from, for example, cracks or nonuniformity in thickness t.

Note that the area enlargement ratio Φ is higher and the elongation of the plate material is larger when the wave height h is higher than 1.2 mm than when the wave height h is 1.0 to 1.2 mm. Thus, in the former case, there is a possibility that the heat transfer plate 20 suffers from, for example, cracks or nonuniformity in thickness t.

When the wave height h is less than 1.0 mm, the area enlargement ratio Φ is decreased and the elongation of the plate material is decreased as compared to the case where the wave height h is 1.0 to 1.2 mm. However, when the wave height h falls within this range, the refrigerant flow path is small in size, and thus pressure loss becomes high. In other words, to make the amount of heat exchange to be 15 kw when the wave height h is less than 1.0 mm, it is necessary to increase the number of stacked plates in the heat transfer plate 20 in order to reduce the pressure loss. Thus, it is impossible to reduce the weight of the plate-type heat exchanger 100.

[Setting of Area Enlargement Ratio Φ and Wave Angle θ when Amount of Heat Exchange is 15 kW and 9 kW]

FIG. 6 is a graph showing the amount of weight reduction of the plate-type heat exchanger 100 when the area enlargement ratio Φ and the wave angle θ are changed as parameters. It should be noted that in FIG. 6, the plate thickness t is 0.2 mm. In FIG. 6, each solid line indicates a result in a plate-type heat exchange that exchanges heat in an amount of 15 kW, and each dashed line indicates a result in a plate-type heat exchanger that exchanges heat in an amount of 9 kW.

As shown in FIG. 6, when heat is exchanged in either amount, the area enlargement ratio Φ is desirably set to 1.05 to 1.15. This is because in the case of this area enlargement ratio Φ, it is possible to make the weight reduction ratio equal to or higher than 20% or close to 20% by adjusting the wave angle θ.

It should be noted that when the area enlargement ratio Φ is 1.05 to 1.15, the wave angle θ is desirably set to fall within the range of 40 degrees to 50 degrees. One reason is that when the wave angle θ falls within this range, while the weight reduction ratio is ensured, it is possible to restrain flow of refrigerant in the short-side direction of the heat transfer plate 20 from being nonuniform. Another reason is that when the wave angle θ falls within this range, it is possible to restrain flow of refrigerant on the width side of the heat transfer plate 20 from being interrupted and then the effective heat transfer area from decreasing or dust clogging from occurring, and thus it is possible to reduce pressure loss.

From the above description with reference to FIGS. 5 and 6, when the plate thickness t of the heat transfer plate 20 is reduced to be equal to or smaller than 0.2 mm, the wave height h, the area enlargement ratio Φ, and the wave angle θ are desirably set as follows in order to reduce the weight of the plate-type heat exchanger 100 regardless of the amount of heat exchange of the plate-type heat exchanger 100. Specifically, the wave height h is desirably set at 1.0 to 1.2 mm, the wave angle θ is desirably set to fall within the range of 40 degrees to 50 degrees, and the area enlargement ratio Φ is desirably set at 1.05 to 1.15. This makes it possible to obtain an optimum weight reduction effect when the plate thickness t is reduced to be equal to or smaller than 0.2 mm.

FIGS. 5 and 6 illustrate the case where the plate thickness t is 0.2 mm. However, even in the case where the plate thickness t is less than 0.2 mm, by setting the wave pitch Λ and the wave height h to fall within the above ranges (values), it is possible to increase the weight reduction ratio, restrain flow of refrigerant in the short-side direction of the heat transfer plate 20 from being nonuniform, and restrain the heat transfer plate 20 from suffering from, for example, cracks or nonuniformity in thickness t.

[Setting of Wave Pitch Λ]

In the above description with reference to FIGS. 5 and 6, when the wave angle θ is set at 40 degrees to 50 degrees and the wave height h is set at 1.0 to 1.2 mm, it is possible to obtain an optimum weight reduction effect by reducing the plate thickness t regardless of the amount of heat exchange of the plate-type heat exchanger 100. In addition, the wave pitch Λ is preferably set equal to or larger than 4 mm. Details of this setting will be described below with reference to FIG. 7.

FIG. 7 is a diagram illustrating the distance between joining points of adjacent heat transfer plates 2 and 3 for each wave angle θ. It should be noted that in FIG. 7(a), the wave angle θ is 65 degrees, and in FIG. 7(b), the wave angle θ is 45 degrees. A joining point corresponds to the position of a point at which a line connecting the peaks of the concaves/convexes of the heat transfer plate 2 intersects with a line connecting the peaks of the concaves/convexes of the heat transfer plate 3. Dashed lines shown in FIG. 7(a) and FIG. 7(b) represent lines connecting the peaks of the concaves/convexes formed in the adjacent heat transfer plates 2 and 3. Points “a” and b shown in FIG. 7 represent joining points of the heat transfer plates 2 and 3, which are closest to each other in the short-side direction among the joining points. L1 is the distance between the points “a” and b when the wave angle θ is 65 degrees, and L2 is the distance between the points a and b when the wave angle θ is 45 degrees.

As shown in FIG. 7(a) and FIG. 7(b), when the wave angle θ is decreased from 65 degrees to 45 degrees, it is possible to restrain flow of refrigerant in the short-side direction of the heat transfer plates 2 and 3 from being nonuniform, and it is possible to restrain flow of refrigerant on the width side of the heat transfer plates 2 and 3 from being interrupted and then the effective heat transfer area from decreasing or dust clogging from occurring.

On the other hand, when the wave angle θ is decreased from 65 degrees to 45 degrees, distance L1>distance L2. In other words, the distance between joining points closest to each other in the short-side direction decreases.

Therefore, when the wave angle θ is set at 45 degrees, if the wave pitch Λ is narrowed too much, the distance L2 further decreases. Thus, there is a possibility that the joining points are filled with a brazing material and the refrigerant flow path is blocked (pressure loss is increased). Therefore, the wave pitch Λ is desirably set at 4 to 7 mm. By this setting, the radius of a minimum fillet becomes about 1.5 mm and about 50% or more is ensured as a refrigerant flow path. Thus, the refrigerant flow path is restrained from being blocked.

It should be noted that if the wave pitch Λ is widened too much, the number of joining points of adjacent heat transfer plates 2 and 3 decreases, and thus the heat transfer efficiency lowers. However, when the wave pitch Λ is set at 4 to 7 mm, it is possible to restrain the heat transfer efficiency from decreasing.

[Advantageous Effects of Plate-type Heat Exchanger 100 According to Embodiment 1]

In the plate-type heat exchanger 100 according to Embodiment 1, when the plate thickness t of the heat transfer plate 20 is set equal to or smaller than 0.2 mm, the wave height h is set at 1.0 to 1.2, the wave angle θ is set at 40 degrees to 50 degrees, and the wave pitch Λ is set at 4 to 7 mm, it is possible to increase the heat transfer efficiency, to reduce pressure loss, to make flow of refrigerant in the short-side direction to be uniform, to restrain blocking of the refrigerant flow path and an increase in cost, and to provide a lightweight configuration.

Specifically, in the heat transfer plate 20, the wave height h is set at 1.0 to 1.2 mm, the wave angle θ is set at 40 degrees to 50 degrees, and the area enlargement ratio Φ is set at 1.05 to 1.15 (see FIGS. 5 and 6). When the wave height h falls within this range, the area enlargement ratio Φ is decreased and the elongation of the plate material can be decreased; flow of refrigerant in the short-side direction of the heat transfer plate 20 is made nonuniform; the heat transfer plate 20 suffers from, for example, cracks or nonuniformity in thickness t; and the refrigerant flow path is reduced in diameter. Thus, it is possible to increase the flow rate of the refrigerant to improve the heat transfer efficiency.

In addition, in the heat transfer plate 20 of the plate-type heat exchanger 100, the wave angle θ is set at 40 degrees to 50 degrees, and the wave pitch Λ is set at 4 to 7 mm (see FIG. 7).

By this setting, the joining points are restrained from being filled with the brazing material so as not to block the refrigerant flow path by narrowing the wave pitch Λ too much. In addition, the number of joining points of the heat transfer plates 2 and 3 is restrained from being decreased so as not to lower the heat transfer efficiency by widening the wave pitch Λ too much.

Furthermore, in the heat transfer plate 20 of the plate-type heat exchanger 100, since the wave height h is set at 1.0 mm to 1.2 mm and the wave pitch Λ is set at 4 to 7 mm, the area enlargement ratio Φ can fall within the range of 1.05 to 1.15. Thus, the refrigerant flow path is reduced in diameter, and hence it is possible to increase the flow rate of the refrigerant to improve the heat transfer efficiency.

Moreover, it is possible to keep the elongation of the plate material small in forming the heat transfer plate 20 from the plate material, and it is possible to restrain the heat transfer plate 20 from suffering from, for example, cracks or nonuniformity in thickness t. In other words, the plate-type heat exchanger 100 is less liable to loss of strength (high strength). Thus, it is possible to reduce the thickness of the plate material, and, in turn, to reduce the material cost and the weight. Since the thickness can be reduced, it is possible to set the set load in press working to be low, and thus it is possible to reduce the cost incurred in this process.

It should be noted that if the plate thickness t, the wave height h, the wave angle θ, and the wave pitch Λ in the heat transfer plate 20 fall outside the ranges described in Embodiment 1, the following adverse effects are produced.

First, if the plate thickness t falls outside the aforementioned range, the weight of the plate-type heat exchanger 100 increases.

Second, if the wave height h and the wave angle θ fall outside the aforementioned ranges, the weight of the plate-type heat exchanger 100 increases and the heat transfer plate 20 suffers from, for example, cracks or nonuniformity in thickness t, or the weight of the plate-type heat exchanger 100 is increased due to increase in the number of stacked plates in the heat transfer plate 20 with increase of pressure loss.

Lastly, if the wave pitch Λ falls outside the aforementioned range, the heat transfer efficiency decreases due to blocking of the refrigerant flow path or a decrease in the number of joining points of adjacent heat transfer plates 2 and 3.

[Others]

The plate-type heat exchanger 100 achieves a reduction in pressure loss and high strength as described above. Therefore, the plate-type heat exchanger 100 is able to reduce pressure loss and restrain deformation of the heat transfer plate 20 even when, for example, a CO₂ refrigerant which allows for operation at a high pressure, a hydrocarbon refrigerant, a low-GWP refrigerant which has a low density and is flammable, or the like is supplied thereto.

In addition, the plate-type heat exchanger 100 allows for a reduction in elongation ratio of the plate material as described above, and thus the heat transfer plate 20 may be formed from a metal such as stainless steel having an elongation ratio of 30% or higher (an elongation ratio of 40%), copper (an elongation ratio of 40%), industrial aluminum (an elongation ratio of 30%), titanium having an elongation ratio as low as 20% or lower (an elongation ratio of 14%), corrosion-resistant aluminum (an elongation ratio of 16%), or the like, or the heat transfer plate 20 may be formed from a synthetic resin or the like.

In the above description, each heat transfer plate 2 is obtained by inverting the heat transfer plate 3 and has the same configuration as the heat transfer plate 3, but is not limited to this form. In other words, each heat transfer plate 2 and each heat transfer plate 3 may have any form as long as the plate thickness t is set equal to or smaller than 0.2 mm, the wave height h is set to fall within the range of 1.0 to 1.2 mm, the wave angle θ is set to fall within the range of 40 degrees to 50 degrees, the wave pitch Λ is set to fall within the range of 4 to 7 mm, and the area enlargement ratio Φ is set to fall within the range of 1.05 to 1.15.

Embodiment 2

FIG. 8 is a diagram illustrating a refrigeration cycle apparatus (air-conditioning apparatus) according to Embodiment 2 of the present invention. In Embodiment 2, the same portions as those in Embodiment 1 are denoted by the same reference signs, and the difference from Embodiment 1 will be mainly described. It should be noted that the refrigeration cycle apparatus according to Embodiment 2 refers to, for example, an apparatus for air-conditioning, for generation of electricity, or for thermally sterilizing food, which includes a plate-type heat exchanger. In the following description, the case where the refrigeration cycle apparatus serves as an air-conditioning apparatus 200 will be taken as an example.

The air-conditioning apparatus 200 according to Embodiment 2 includes one outdoor unit 101 serving as a heat source device, one indoor unit 102, and a heat medium relay unit 103 for transferring cooling energy of a heat source side refrigerant flowing through the outdoor unit 101 to a heat medium flowing through the indoor unit 102.

The outdoor unit 101 and the heat medium relay unit 103 are connected to each other via a refrigerant pipe 120 through which the heat source side refrigerant (first refrigerant) flows, thereby forming a refrigerant circulation circuit A. Also, the heat medium relay unit 103 and the indoor unit 102 are connected to each other via a heat medium pipe 121 through which the heat medium (second refrigerant) flows, thereby forming a heat medium circulation circuit B.

The outdoor unit 101 is equipped with at least a heat source side heat exchanger 110, a compressor 118, and an expansion device 111.

The indoor unit 102 is equipped with at least a use side heat exchanger 112.

The heat medium relay unit 103 is equipped with at least the plate-type heat exchanger 100 according to Embodiment 1 and a pump 119.

It should be noted that an example will be described in which the plate-type heat exchanger 100 is provided in the heat medium relay unit 103, but the plate-type heat exchanger 100 may be used as at least one of heat exchangers of the outdoor unit 101 the indoor unit 102, and the heat medium relay unit 103.

In addition, in Embodiment 2, the air-conditioning apparatus 200 which performs cooling operation will be taken as an example of the refrigeration cycle apparatus, but it is needless to say that a four-way valve or the like may be provided in the refrigerant circulation circuit A to enable heating operation to be performed.

The heat source side heat exchanger 110 serves as a condenser and exchanges heat between the outdoor air and the heat source side refrigerant flowing through the refrigerant pipe 120. The heat source side heat exchanger 110 is connected on its one side to the plate-type heat exchanger 100 and is connected on its other side to the discharge side of the compressor 118.

The compressor 118 compresses the heat source side refrigerant and conveys the heat source side refrigerant to the refrigerant circulation circuit A. The compressor 118 is connected on its discharge side to the heat source side heat exchanger 110 and is connected on its suction side to the plate-type heat exchanger 100.

The expansion device 111 reduces the pressure of the heat source side refrigerant flowing through the refrigerant pipe 120, to expand the heat source side refrigerant. The expansion device 111 is connected on its one side to the heat source side heat exchanger 110 and is connected on its other side to the plate-type heat exchanger 100. The expansion device 111 may be composed of, for example, a capillary tube or a solenoid valve.

The use side heat exchanger 112 exchanges heat between the heat medium flowing through the heat medium pipe 121 and air in an air-conditioned space. The use side heat exchanger 112 is connected on its one side to the plate-type heat exchanger 100 and is connected on its other side to the suction side of the pump 119.

The plate-type heat exchanger 100 exchanges heat between the heat source side refrigerant and the heat medium. The plate-type heat exchanger 100 is connected to the suction side of the compressor 118 and the expansion device 111 via the refrigerant pipe 120. In addition, the plate-type heat exchanger 100 is connected to the use side heat exchanger 112 and the pump 119 via the heat medium pipe 121. In other words, the plate-type heat exchanger 100 is cascade-connected to the refrigerant circulation circuit A and the heat medium circulation circuit B.

The pump 119 conveys the heat medium to the heat medium circulation circuit B. The pump 119 is connected on its suction side to the use side heat exchanger 112 and is connected on its discharge side to the plate-type heat exchanger 100.

[Description of Operation]

Row of the heat source side refrigerant in the refrigerant circulation circuit A will be described.

A low-temperature and low-pressure heat source side refrigerant is compressed by the compressor 118 into a high-temperature and high-pressure gas refrigerant and discharged therefrom. The high-temperature and high-pressure gas refrigerant discharged from the compressor 118 flows into the heat source side heat exchanger 110. Then, the high-temperature and high-pressure gas refrigerant turns into a high-pressure liquid refrigerant while transferring heat to the outdoor air at the heat source side heat exchanger 110. The high-pressure liquid refrigerant having flowed out from the heat source side heat exchanger 110 is expanded by the expansion device 111 into a low-temperature and low-pressure two-phase refrigerant. The low-temperature and low-pressure two-phase refrigerant flows into the plate-type heat exchanger 100 which serves as an evaporator. Then, the low-temperature and low-pressure two-phase refrigerant turns into a low-temperature and low-pressure gas refrigerant while cooling the heat medium by removing heat from the heat medium circulating through the heat medium circulation circuit B. The gas refrigerant having flowed out from the plate-type heat exchanger 100 is sucked into the compressor 118 again.

Flow of the heat medium in the heat medium circulation circuit B will be described next.

The heat medium having compressed and flowed out from the pump 119 flows into the plate-type heat exchanger 100, and cooling energy of the heat source side refrigerant in the plate-type heat exchanger 100 is transferred to the heat medium. The heat medium flows out from the plate-type heat exchanger 100 and flows into the use side heat exchanger 112. Then, the heat medium removes heat from the indoor air at the use side heat exchanger 112, thereby cooling the air-conditioned space. The heat medium having flowed out from the use side heat exchanger 112 is sucked into the pump 119 again.

REFERENCE SIGNS LIST

1,4 side plate 2, 3, 20 heat transfer plate 5 first refrigerant inlet pipe 6 second refrigerant inlet pipe 7 first refrigerant outlet pipe 8 second refrigerant outlet pipe 9 concaves/convexes 10 concaves/convexes 11 first opening 12 third opening 13 second opening 14 fourth opening 100 plate-type heat exchanger 101 outdoor unit 102 indoor unit 103 heat medium relay unit 110 heat source side heat exchanger 111 expansion device 112 use side heat exchanger 118 compressor 119 pump 120 refrigerant pipe 121 heat medium pipe 200 air-conditioning apparatus A refrigerant circulation circuit B heat medium circulation circuit X first refrigerant flow path Y second refrigerant flow path

Claims

1. A plate-type heat exchanger having an inlet through which a fluid flows therein, an outlet through which the fluid having flowed therein through the inlet flows out therefrom, and a flow path connecting the inlet to the outlet and formed in a space which is formed by waves of concaves/convexes of adjacent heat transfer plates among a plurality of heat transfer plates having a plurality of substantially V-shaped waves of concaves/convexes formed and aligned from the inlet toward the outlet, the plurality of heat transfer plates being stacked in alternate layers of inverted heat transfer plates and noninverted heat transfer plates, wherein

a plate thickness t of each heat transfer plate is equal to or less than 0.2 mm,

a pitch Λ of the concaves/convexes is 4 to 7 mm,

a distance h between peaks of the concaves/convexes is 1.0 to 1.2 mm, and

when a value obtained by dividing a wave length s corresponding to a length of the heat transfer plate between peaks of the wave of the concaves/convexes of the heat transfer plate by the pitch Λ of the concaves/convexes is defined as an area enlargement ratio Φ, the area enlargement ratio Φ is 1.05 to 1.15.

2. The plate-type heat exchanger of claim 1, wherein an angle θ of spread of the waves with respect to a direction in which the substantially V-shaped waves of concaves/convexes are aligned is 40 degrees to 50 degrees.

3. The plate-type heat exchanger of claim 1, wherein the angle θ of spread of the waves, the plate thickness t, the pitch Λ, and the distance h are the same in all the heat transfer plates.

4. The plate-type heat exchanger of claim 1, wherein each of the heat transfer plates is formed from titanium, corrosion-resistant aluminum, or stainless steel.

5. A refrigeration cycle apparatus comprising two refrigerant circuits which are cascade-connected to each other via the plate-type heat exchanger of claim 1.

6. The plate-type heat exchanger of claim 2, wherein the angle θ of spread of the waves, the plate thickness t, the pitch Λ, and the distance h are the same in all the heat transfer plates.

7. The plate-type heat exchanger of claim 2, wherein each of the heat transfer plates is formed from titanium, corrosion-resistant aluminum, or stainless steel.

8. The plate-type heat exchanger of claim 3, wherein each of the heat transfer plates is formed from titanium, corrosion-resistant aluminum, or stainless steel.

9. A refrigeration cycle apparatus comprising two refrigerant circuits which are cascade-connected to each other via the plate-type heat exchanger of claim 2.

10. A refrigeration cycle apparatus comprising two refrigerant circuits which are cascade-connected to each other via the plate-type heat exchanger of claim 3.

11. A refrigeration cycle apparatus comprising two refrigerant circuits which are cascade-connected to each other via the plate-type heat exchanger of claim 4.