Double pipe for heat exchanger

Abstract

A double pipe is configured with an inner pipe (3) inside an outer pipe. The double pipe is used to perform heat exchange between a fluid flowing inside the inner pipe (3) and a fluid flowing between the inner pipe (3) and the outer pipe. In a straight pipe portion of the double pipe, the inner pipe (3) has a plurality of convex portions (31) bent in a manner projecting toward an outer peripheral side and extending in a length direction. The convex portions (31) are displaced in a spiral shape in the length direction of the double pipe. In the straight pipe portion, an inner peripheral surface of the outer pipe is in contact with the convex portions (31) of the inner pipe (3), and an outer side flow path divided into a plurality of portions in a peripheral direction is formed between the outer pipe and the inner pipe (3). In a cross section of the straight pipe portion orthogonal to the length direction, when a maximum depth of each of the outer side flow paths is set as D [mm] and a circular arc length in the peripheral direction is set as L [mm], an average value of values of D / L of all the outer side flow paths is greater than 0.09 and less than 0.20.

Description

Technical Field

[0001] This invention relates to double tubes for heat exchangers. Background Technology

[0002] The heat exchange cycle (also known as the refrigeration cycle) of automotive air conditioning systems includes a condenser, an evaporator, a compressor, and an expansion valve. Freon, CO2, ammonia, and other refrigerants circulate in the cycle path connecting these components. In this heat exchange cycle, a proposed scheme involves arranging a double-pipe system in the cycle path. In the two-layer space formed by these double pipes, the high-temperature refrigerant exiting the condenser and the low-temperature refrigerant exiting the evaporator flow relative to each other, thus exchanging heat and improving heat exchange performance (see Patent Document 1).

[0003] On the other hand, as refrigerants used in heat exchange cycles, refrigerants with lower global warming coefficients have been studied to address environmental concerns. However, there are concerns that these environmentally problematic refrigerants may exhibit reduced heat exchange performance compared to existing refrigerants. Therefore, to mitigate the overall performance degradation of the heat exchange cycle, it is effective to actively adopt a structure that further improves heat exchange performance by assembling the aforementioned double-tube configuration.

[0004] In systems where a compressor compresses gaseous refrigerant discharged from the evaporator, if the refrigerant flows into the compressor in an incompletely vaporized state—that is, while the gaseous state is mixed with liquid—it may result in insufficient heat exchange. However, this problem can be eliminated by assembling a double-pipe system. This is because, in a double-pipe system, the refrigerant can be heated before flowing into the compressor, ensuring complete vaporization.

[0005] Here, the double tube described in Patent Document 1 uses an inner tube having multiple spiral grooves as the inner tube, and combines this inner tube with a smooth cylindrical outer tube. Furthermore, in the straight tube section of the double tube, the inner diameter of the outer tube is larger than the outer diameter of the inner tube, and the peaks between the grooves in the inner tube do not abut against the outer tube. Thus, in the straight tube section of the double tube, the grooves are interconnected circumferentially, and the outer flow path formed between the outer tube and the inner tube is formed throughout the entire circumference.

[0006] Prior art literature

[0007] Patent documents

[0008] Patent Document 1: Japanese Patent Application Publication No. 2006-162241 Summary of the Invention

[0009] The problem that the invention aims to solve

[0010] In the dual-pipe system described in Patent Document 1, as mentioned above, the inner diameter of the outer pipe is larger than the outer diameter of the inner pipe in the straight pipe section. The outer flow path becomes a single flow path connected circumferentially, resulting in a larger cross-sectional area compared to the curved section. Therefore, the refrigerant flow velocity in the outer flow path tends to decrease in the straight pipe section, potentially preventing the full utilization of heat exchange performance at the straight pipe section.

[0011] While double-tube heat exchangers have some bends, they are mostly composed of straight sections. Therefore, it can be assumed that improving the heat exchange performance of the straight sections can improve the overall heat exchange performance of the double-tube. On the other hand, although the heat exchange performance of a double-tube is considered to be affected not only by the flow velocity, flow area, and pressure loss in the inner and outer flow paths, but also by the flow path cross-sectional shape and other optimizations, the optimal conditions cannot yet be definitively determined. In particular, the effective flow path cross-sectional shape of the outer flow path has not yet been fully elucidated.

[0012] The present invention was made in view of the following background, and its object is to provide a double tube for a heat exchanger that optimizes the flow path cross-sectional shape of the outer flow path in the straight tube section, thereby improving the heat exchange performance compared to the past.

[0013] Methods for solving problems

[0014] One aspect of the invention lies in a double-tube heat exchanger having a double-tube structure in which an inner tube is disposed inside an outer tube, for heat exchange between a fluid flowing inside the inner tube and a fluid flowing before the inner tube and the outer tube, wherein...

[0015] In the straight section of the double tube for the heat exchanger, the inner tube has a plurality of protrusions that bend outwards and extend along the length direction.

[0016] The protrusion is helically displaced along its length.

[0017] In the cross-section of the straight tube section orthogonal to its length, the inner circumferential surface of the outer tube is circular.

[0018] In the straight tube section, the inner circumferential surface of the outer tube is in contact with the protrusion of the inner tube, and an outer flow path divided into multiple circumferential points is formed between the outer tube and the inner tube.

[0019] In the cross-section of the straight pipe section orthogonal to the length direction, when the maximum depth of each of the outer flow paths is set as D [mm] and the circumferential arc length is set as L [mm], the average value of D / L of all the outer flow paths is greater than 0.09 and less than 0.20.

[0020] Invention Effects

[0021] In the straight section of the dual-tube heat exchanger, the inner circumferential surface of the outer tube is in contact with the protrusion of the inner tube, and an outer flow path divided into multiple circumferential points is formed between the outer and inner tubes. Thus, by forming the outer flow path in the straight section in a circumferentially segmented manner, the flow velocity of the fluid flowing in the outer flow path of the straight section can be ensured, thereby improving the heat exchange performance of the straight section.

[0022] Furthermore, in the cross-section of the straight pipe section orthogonal to the length direction, when the maximum depth of each outer flow path is set as D [mm] and the circumferential arc length is set as L [mm], the average value of D / L for all outer flow paths is greater than 0.09 and less than 0.20. That is, in the straight pipe section, each outer flow path is formed with a relatively small depth and a relatively long circumferential arc length. Therefore, in the straight pipe section, the contact area between the fluid flowing in the outer flow paths and the outer and inner pipes can be increased. As a result, the heat exchange performance in the straight pipe section can also be improved.

[0023] Furthermore, by setting the average D / L value of all outer flow paths in the cross-section of the straight pipe section to a value greater than 0.09, it is possible to suppress the increase in pressure loss of the fluid flowing in each outer flow path in the straight pipe section. Moreover, by setting the average value to a value less than 0.20 in the cross-section of the straight pipe section, it is possible to suppress the increase in pressure loss of the fluid flowing in the flow path within the inner pipe of the straight pipe section.

[0024] As described above, according to this method, a double tube for a heat exchanger can be provided, which optimizes the cross-sectional shape of the outer flow path in the straight tube section, thereby improving the heat exchange performance compared to the past. Attached Figure Description

[0025] Figure 1 This is a partial three-dimensional cross-sectional view of the double tube used in the heat exchanger in Example 1.

[0026] Figure 2 This is a partial cross-sectional side view of the double tube used in the heat exchanger in Example 1.

[0027] Figure 3 yes Figure 2 Sectional view along line III-III.

[0028] Figure 4 It is Figure 3 The enlarged image of the portion enclosed by the single-dot dashed line.

[0029] Figure 5 This is a cross-sectional view of the inner tube used to illustrate the maximum depth D of the outer flow path and the circumferential arc length L in Embodiment 1.

[0030] Figure 6This is an explanatory diagram showing the configuration of the pressing plates as viewed from the length direction of the inner tube blank in Embodiment 1.

[0031] Figure 7 This is a cross-sectional view of the inner tube used to illustrate the maximum depth D of the outer flow path and the circumferential arc length L in Embodiment 2.

[0032] Figure 8 This is a cross-sectional view of the double tube of sample 1, which has an average D / L of 0.08 in Experimental Examples 1 to 3.

[0033] Figure 9 This is a cross-sectional view of the double tube of sample 5, which has an average D / L of 0.15 in Experiments 1-3.

[0034] Figure 10 This is a cross-sectional view of the double tube of sample 7, which has an average D / L of 0.20 in Experimental Examples 1-3.

[0035] Figure 11 This is a cross-sectional view of the double tube of sample 8, which served as a comparative sample in Experimental Examples 1-3.

[0036] Figure 12 This is a three-dimensional view of the double tube of sample 5 in Experimental Examples 1-3.

[0037] Figure 13 This is a graph showing the relationship between the average D / L value and the heat absorbed by the refrigerant in the inner flow path in Experimental Example 1.

[0038] Figure 14 This is a graph showing the relationship between the average D / L value and the pressure loss of the refrigerant in the outer flow path in Experiment Example 2.

[0039] Figure 15 This is a graph showing the relationship between the average D / L value and the pressure loss of the refrigerant in the inner flow path in Experiment Example 3. Detailed Implementation

[0040] The heat exchanger double tube can be configured, for example, for use in an air conditioning system for automobiles. The air conditioning system can have a structure including a condenser, an evaporator, a compressor, an expansion valve, and a circulation path connecting them, with the heat exchanger double tube arranged within this circulation path. In this case, for example, the low-temperature, low-pressure gaseous refrigerant from the evaporator is guided to a flow path within the inner tube, and the high-temperature, high-pressure liquid refrigerant from the condenser is guided to an outer flow path. The heat exchanger double tube is assembled into the circulation path such that the liquid refrigerant flowing in the outer flow path and the gaseous refrigerant flowing in the inner tube flow in opposite directions. By configuring the circulation path in this way, heat exchange between the high-temperature, high-pressure liquid refrigerant and the low-temperature, low-pressure gaseous refrigerant is efficiently performed within the heat exchanger double tube, improving the overall heat exchange performance of the heat exchange cycle.

[0041] The heat exchanger is constructed with dual tubes, wherein the average D / L value of all outer flow paths in the cross-section of the straight tube section orthogonal to the length direction is greater than 0.09 and less than 0.20. By setting the average D / L value of the outer flow paths to this specific range, as described above, it is possible to suppress the increase in pressure loss of the fluid flowing in each outer flow path in the straight tube section, and it is also possible to suppress the increase in pressure loss of the fluid flowing in the inner tube. From the viewpoint of more reliably achieving this effect, it is preferable to set the average D / L value of all outer flow paths in the cross-section of the straight tube section orthogonal to the length direction to be greater than 0.10 and less than 0.20, more preferably greater than 0.11 and less than 0.20, even more preferably greater than 0.12 and less than 0.20, even more preferably greater than 0.13 and less than 0.19, and particularly preferably greater than 0.13 and less than 0.18.

[0042] A double tube for a heat exchanger can have a straight tube section extending in a straight line, or it can be partially (in one or more places) bent through bending processes. In this case, at least in the cross-section of the straight tube section orthogonal to the length direction of the double tube, the average value of the D / L ratio for all outer flow paths should be set to a value greater than 0.09 and less than 0.20. This allows for improved heat exchange performance and reduced pressure loss in the straight tube section, which occupies a large portion of the double tube for the heat exchanger.

[0043] The outer tube and the protrusion in the heat exchanger double tube can also be brought into contact with each other by pressing the outer tube toward the inner tube. In this case, the outer tube and the protrusion can be easily brought into direct contact, which can improve the productivity of the heat exchanger double tube having multiple circumferentially divided outer flow paths.

[0044] In the cross-section of the straight pipe section orthogonal to the length direction, the total arc length L of the outer flow path is preferably 60% or more of the circumference C of the virtual circumcircle of the inner pipe. This allows for improved heat exchange performance and reduced pressure loss in the straight pipe section by forming the outer flow path over a wider circumferential region. Furthermore, in the cross-section of the straight pipe section orthogonal to the length direction, the total arc length L of the outer flow path is preferably 90% or less of the circumference C of the inner pipe's circumference. In this case, it is easy to ensure the contact width of the abutment between the inner and outer pipes, reliably separating the outer flow path.

[0045] In the straight pipe section, the number of protrusions provided in the inner pipe is preferably four or more and six or less. By setting the number of protrusions to four or more, it is easy to increase the flow velocity in the outer flow path of the straight pipe section, which can further improve the heat exchange performance. In addition, by setting the number of protrusions to six or less, it is possible to suppress the excessive increase in pressure loss caused by the outer flow path being divided into multiple circumferential points in the straight pipe section.

[0046] In the cross-section of the straight pipe section orthogonal to the length direction of the double tube for heat exchangers, multiple protrusions are preferably formed at equal intervals. This ensures that the refrigerant flow rate in each outer flow path is approximately equal in the circumferential direction. Therefore, uneven heat exchange between the refrigerant flowing in the outer flow path and the refrigerant flowing in the inner pipe flow path is easily suppressed.

[0047] In the straight tube section, the torsion angle of each protrusion is preferably 20° or more and 60° or less. The torsion angle is the angle between the forming direction of the protrusion and a straight line parallel to the length direction of the double tube for heat exchangers. By setting the torsion angle to 20° or more, it is easier to increase the contact area between each outer flow path and the inner tube, thereby improving heat exchange performance. By setting the torsion angle to 60° or less, it is possible to prevent excessive pressure loss of the refrigerant flowing in the outer flow path.

[0048] The boundary between the convex portion and the spiral recess on the outer circumferential surface of the inner tube can also be formed into an angle. Here, "angled" refers to a state where the radius of curvature of the convex portion and the radius of curvature of the spiral recess change in a stepped manner with the boundary as the boundary. When the boundary is angled, the convex portion tends to protrude sufficiently outward relative to the spiral recess, and the radius of curvature tends to be smaller. Therefore, when an angled boundary is formed, it is easy to reliably fit the convex portion against the outer tube, and the outer flow path can be reliably interrupted in the circumferential direction.

[0049] Example

[0050] (Example 1)

[0051] use Figures 1-5 An example of a heat exchanger using a double tube is described.

[0052] like Figure 1 , Figure 2 As shown, the heat exchanger in this example uses a double tube 1 (hereafter sometimes simply referred to as double tube 1) with a double tube structure consisting of an inner tube 3 arranged inside an outer tube 2. The double tube 1 is used for heat exchange between the fluid flowing inside the inner tube 3 and the fluid flowing between the inner tube 3 and the outer tube 2.

[0053] like Figure 1 , Figure 3 As shown, in the straight section 11 of the double tube 1, the inner tube 3 has a plurality of protrusions 31 that bend outwards and extend along the length direction of the double tube 1 (hereinafter referred to as the "X direction"). Figure 2 As shown, the protrusion 31 is displaced in a spiral shape in the X direction. Figure 1 , Figure 3 As shown, in the cross-section of the straight tube 11 orthogonal to the X direction, the inner circumferential surface of the outer tube 2 (i.e., the inner circumferential end edge of the outer tube 2) is circular.

[0054] like Figures 1-3 As shown, in the straight section 11 of the double tube 1, the inner circumferential surface of the outer tube 2 is connected to the protrusion 31 of the inner tube 3, and an outer flow path 4, divided into multiple circumferential sections, is formed between the outer tube 2 and the inner tube 3. Figure 5 As shown, in the cross section of the straight pipe section 11 orthogonal to the X direction, when the maximum depth of each outer flow path 4 is set to D [mm] and the circumferential arc length is set to L [mm], the average value of D / L of all outer flow paths 4 is greater than 0.09 and less than 0.20.

[0055] The double tube 1 in this example will be explained in detail later.

[0056] From now on, when referred to only as circumferential, unless otherwise specified, it refers to the circumferential direction of the double tube 1; when referred only as radial, unless otherwise specified, it refers to the radial direction of the double tube 1.

[0057] like Figure 1 , Figure 2 As shown, the outer tube 2 and the inner tube 3 are formed into a tubular shape, for example, from 1000 series pure aluminum, 3000 series, or 5000 series aluminum alloy. However, they are not limited to this; other metals with excellent thermal conductivity can also be used to form the outer tube 2 and the inner tube 3. In this example, the double tube 1 is formed such that after the inner tube 3 is inserted into the outer tube 2, the outer tube 2 is compressed in such a way that its diameter decreases circumferentially towards the inner circumference, thereby abutting against all the protrusions 31. Furthermore, the area sandwiched between the outer tube 2 and the inner tube 3 becomes the outer flow path 4, and the inner side of the inner tube 3 becomes the inner flow path 6.

[0058] like Figure 2As shown, when using the double tube 1, the refrigerant circulation path is formed such that the direction of the refrigerant flowing in the inner flow path 6 is opposite to the direction of the refrigerant flowing in the outer flow path 4. Figure 2 In the diagram, the direction of the refrigerant flowing in the inner flow path 6 is indicated by a dashed arrow, and the direction of the refrigerant flowing in the outer flow path 4 is indicated by a solid arrow.

[0059] like Figure 1 As shown, the outer tube 2 is a smooth tube with a circular annular cross-section orthogonal to the X direction. On the other hand, the inner tube 3 is formed with a cross-sectional shape orthogonal to the X direction. Figure 3 The shape is uneven as shown. When the cross-sectional shape of the inner tube 3 in the straight tube section 11 is observed at various positions in the X direction, the cross-sectional shape at each position in the X direction is approximately the same as the shape obtained by rotating the cross-sectional shape at other positions circumferentially.

[0060] The inner tube 3 has four spiral recesses 32 for forming an outer flow path 4 between itself and the outer tube 2. For example... Figure 2 As shown, the spiral recess 32 is formed in a spiral shape that rotates circumferentially on the outer surface of the inner tube 3 while advancing in the X direction. The angle formed by the forming direction of the spiral recess 32 and a straight line parallel to the X direction, i.e., the twist angle of the spiral recess 32, is 20° or more and 60° or less. Figure 1 , Figure 3 As shown, the spiral recess 32 is formed as a curved surface that bulges inward toward the inner circumference.

[0061] The inner tube 3, which has a protrusion 31 and a spiral recess 32, can be manufactured, for example, as follows. First, prepare as follows: Figure 6 The tube blank 30 has a smooth tube shape with a circular cross-section, as shown by the double-dotted line. Then, by passing the tube blank 30 through the area surrounded by the four pressing discs 8 (described later), the portion of the tube blank 30 that forms the spiral recess 32 is deformed.

[0062] Here, four pressing discs 8 are arranged at equal intervals in the circumferential direction. Each pressing disc 8 has a disc shape and is configured to rotate freely in the circumferential direction of the pressing disc 8. Each pressing disc 8 is inclined relative to the length direction of the tube blank 30. That is, each pressing disc 8 is inclined in order to form a spiral recess 32 with the aforementioned torsion angle. Furthermore, since each pressing disc 8 presses the tube blank 30 towards the inner circumferential side when the tube blank 30 is passed through the area surrounded by them, a portion of each pressing disc 8 is configured to protrude towards the inner circumferential side from the outer circumferential surface of the tube blank 30.

[0063] Then, by passing the tube blank 30 through the space surrounded by four pressing discs 8, the side portion 81 of each pressing disc 8 presses the tube blank 30 inwards while the discs 8 rotate. This forms a spiral recess 32 in the tube blank 30. Furthermore, the portions where the spiral recess 32 is not formed, i.e., the portions between adjacent spiral recesses 32 in the circumferential direction within the inner tube 3, become protrusions 31 formed along the spiral recesses 32. Additionally, while forming the spiral recesses 32, the protrusions 31 also deform in a manner that follows the deformation of the spiral recesses 32, and the radius of curvature of the protrusions 31 is formed to be smaller than the radius of curvature of the tube blank 30. The deeper the spiral recesses 32 are formed, the smaller the radius of curvature of the protrusions 31 becomes. As described above, the inner tube 3 can be manufactured.

[0064] like Figure 1 , Figure 3 As shown, the inner tube 3, in a cross-sectional shape orthogonal to the X direction, has protrusions 31 at four equally spaced locations in the circumferential direction. Each protrusion 31 is a portion that protrudes outwards from the circumferential side in the cross-sectional shape of the inner tube 3, compared to its adjacent portion in the circumferential direction. The torsion angle of the protrusions 31 is also the same as the torsion angle of the spiral recesses 32, being 20° or more and 60° or less. The protrusions 31 of the inner tube 3 are formed as curved surfaces that bend outwards from the circumferential side.

[0065] like Figure 3 As shown, the boundary 33 between the protrusion 31 and the spiral recess 32 on the outer circumferential surface of the inner tube 3 is formed in an angular shape. This angular boundary 33 is formed, for example, by applying sufficient pressure from the pressing plate 8 to the tube blank 30 during the aforementioned manufacturing process of the inner tube 3. That is, when the boundary 33 is angular, the protrusion 31 tends to protrude sufficiently outward relative to the spiral recess 32, and the radius of curvature tends to be smaller. Therefore, with the angular boundary 33 formed, the protrusion 31 can be reliably and tightly fitted to the outer tube 2, enabling reliable circumferential separation of the outer flow path 4.

[0066] like Figure 3 As shown, all the protrusions 31 abut against the inner circumferential surface of the outer tube 2. All the protrusions 31 abut against the inner circumferential surface of the outer tube 2 at least in the entire X-direction region of the straight section 11 of the double tube 1. In this example, all the protrusions 31 abut against the inner circumferential surface of the outer tube 2 in the entire region where the protrusions 31 exist in the X-direction. That is, even if a portion of the double tube 1 is bent, all the protrusions 31 abut against the inner circumferential surface of the outer tube 2 in that bent portion.

[0067] like Figure 3As shown, the protrusion 31 is formed in a curved manner along the inner circumferential surface of the outer tube 2 (i.e., the difference between the radius of curvature of the protrusion 31 and the inner circumferential surface of the outer tube 2 is not greater than a given value), thereby easily ensuring the contact width between the protrusion 31 and the outer tube 2. Furthermore, in the cross-section of the straight tube portion 11 orthogonal to the X direction, the total length of the contact width between each protrusion 31 and the outer tube 2 is preferably less than 80% of the circumference C of the virtual circumcircle of the inner tube 3. This ensures the circumferential length of each outer flow path 4, enabling improved heat exchange performance and reduced pressure loss of the refrigerant flowing in the outer flow path 4. Moreover, by having four circumferential protrusions 31 abut against the outer tube 2, four outer flow paths 4, formed by circumferentially dividing the protrusions 31, are formed between the outer tube 2 and the inner tube 3.

[0068] like Figure 5 As shown, in the cross-section of the straight pipe section 11 orthogonal to the X direction, when the maximum depth of each outer flow path 4 is set to D [mm] and the circumferential arc length is set to L [mm], the average value of D / L for all outer flow paths 4 (average D / L value) is greater than 0.09 and less than 0.20. That is, in the cross-section orthogonal to the X direction, each outer flow path 4 is formed to a certain extent to be shallower in the radial direction and wider in the circumferential direction.

[0069] The maximum depth D of each outer flow path 4 in the cross section orthogonal to the X direction of the straight pipe section 11 is obtained by the maximum radial length between the virtual circumscribed circle 7 of the inner pipe 3 and the inner pipe 3. That is, in the cross section orthogonal to the X direction of the straight pipe section 11, the radial length between the outer flow path 4 and the circumscribed circle 7 varies depending on the circumferential position. The maximum value among these radial lengths is set as the maximum depth D of the outer flow path 4. For example, when a pipe with a diameter of approximately 19 mm is used as the tube blank (refer to...), Figure 6 When the symbol 30 is processed to form the inner tube 3, the maximum depth D of each outer flow path 4 can be set to, for example, 0.7 mm to 1.6 mm.

[0070] The circumferential arc length L of each outer flow path 4 in the cross section orthogonal to the X direction of the straight pipe section 11 is obtained by the following method. First, the contact portion 5 between each protrusion 31 and the outer pipe 2 is determined. Next, on the circumcircle 7 of the inner pipe 3, the length of the arc from any contact portion 5 to the adjacent contact portion 5 is measured. This arc length is set as the arc length L of the outer flow path 4 located between these contact portions 5. Here, the protrusion 31 is formed as a curved surface that bends along the inner circumferential surface of the outer pipe 2, therefore... Figure 4 As shown in the enlarged view of the boundary periphery between the contact portion 5 and the outer flow path 4, the gap g between the protrusion 31 and the outer tube 2 gradually increases radially as it approaches the center of the outer flow path 4 in the circumferential direction.

[0071] On the other hand, the size of the gap g between the protrusion 31 and the outer tube 2 gradually narrows as it moves away from the center of the outer flow path 4 in the circumferential direction. Therefore, it is difficult to specifically determine the position where the gap g between the protrusion 31 and the outer tube 2 completely disappears, i.e., the true endpoint of the abutment 5. Therefore, in this example, the radial length of the gap g around the boundary between the abutment 5 and the outer flow path 4 is 4 μm, and the position that can be clearly identified as the gap is set as the endpoint 51 of the abutment 5. Furthermore, the circumferential length of the circumferential circle 7 between the endpoint 51 of the abutment 5 on one side of the outer flow path 4 and the endpoint 51 of the abutment 5 on the other side is set as the circumferential arc length L of the outer flow path 4.

[0072] In the cross section of the straight pipe section 11 orthogonal to the X direction, when the radius of the circumscribed circle 7 of the inner pipe 3 is set to r [mm], and the angle between the virtual straight line V connecting the endpoints 51 of the adjacent abutment portions 5 on both sides of the outer flow path 4 and the center of the circumscribed circle 7 is set to θ [°], the circumferential arc length L of the outer flow path 4 can be obtained by L = 2πr × θ / 360. For example, when a pipe with a diameter of approximately 19 mm is used as the pipe blank (refer to...), Figure 6 When the inner tube 3 is formed by processing the symbol 30), the circumferential arc length L of each outer flow path 4 in the cross section orthogonal to the X direction of the straight tube 11 is, for example, 7 mm to 14 mm.

[0073] like Figure 5 As shown, the total circumferential arc length L of the outer flow path 4 in the cross section orthogonal to the X direction of the straight pipe section 11 is 60% or more of the circumference C of the circumference circle 7 of the inner pipe 3. This facilitates improved heat exchange performance and reduced pressure loss in the straight pipe section 11. From the viewpoint of ensuring sufficient contact width of the contact portion 5 between the inner pipe 3 and the outer pipe 2, in this example, it is preferable to set the total circumferential arc length L of the outer flow path 4 in the cross section orthogonal to the X direction of the straight pipe section 11 to 90% or less of the circumference C of the circumference circle 7 of the inner pipe 3. Furthermore, from the viewpoint of ensuring sufficient contact width of the contact portion 5 and improving heat exchange performance and reducing pressure loss, it is preferable to set the total circumferential arc length L of the outer flow path 4 in the cross section orthogonal to the X direction of the straight pipe section 11 to 63% or more and 77% or less of the circumference C of the circumference circle 7 of the inner pipe 3.

[0074] In the cross-section of the straight pipe section 11 orthogonal to the X direction, the total cross-sectional area of ​​the outer flow paths 4 is at least 5% of the cross-sectional area of ​​the inner flow path 6. This improves the heat exchange performance of the straight pipe section 11 and reduces the pressure loss of the refrigerant flowing in the outer flow paths 4. Furthermore, from the viewpoint of improving the productivity of the double pipe 1 and reducing the pressure loss of the refrigerant in the inner flow path 6, the total cross-sectional area of ​​the outer flow paths 4 in the cross-section of the straight pipe section 11 orthogonal to the X direction is preferably 30% or less of the cross-sectional area of ​​the inner flow path 6. In the cross-section of the straight pipe section 11 orthogonal to the X direction, to increase the total cross-sectional area of ​​the outer flow paths 4, the depth of the outer flow paths 4 in the inner pipe 3 must be considerably large. However, increasing the total cross-sectional area of ​​the outer flow path 4 while maintaining the average D / L value at 0.09 to 0.20 may lead to a deterioration in the productivity of the inner tube 3 and an increase in the pressure loss of the refrigerant flowing within the inner tube 3. These problems can be more easily avoided by setting the total cross-sectional area of ​​the outer flow path 4 to 30% or less of the cross-sectional area of ​​the inner flow path 6. Furthermore, from the viewpoints of improving the productivity of the double tube 1, improving the heat exchange performance of the straight tube section 11, reducing the pressure loss of the refrigerant in the outer flow path 4, and reducing the pressure loss of the refrigerant in the inner flow path 6, the total cross-sectional area of ​​the outer flow path 4 in the cross-section of the straight tube section 11 orthogonal to the X direction is preferably 9% or more and 20% or less of the cross-sectional area of ​​the inner flow path 6.

[0075] Furthermore, as described above, the cross-sectional shape of the inner tube 3 in the straight tube section 11, orthogonal to the X direction, is approximately the same at any position in the X direction. Therefore, at any position in the X direction, the maximum depth D and the circumferential arc length L of each outer flow path 4 are also approximately constant. Moreover, in the cross-sections at each position where the protrusion 31 exists in the straight tube section 11, the average value of D / L for all outer flow paths 4 is greater than 0.09 and less than 0.20.

[0076] In this example, the outer diameter of the outer tube 2 is set to be in the range of 15mm to 30mm, and the outer diameter of the circumscribed circle of the inner tube 3 is set to be in the range of 10mm to 29mm. In addition, the thickness of the outer tube 2 and the thickness of the inner tube 3 are equal.

[0077] The dual-pipe 1 in this example is used in air conditioning systems such as those for automobiles. The air conditioning system includes a condenser, evaporator, compressor, expansion valve, and a circulation path connecting them; the dual-pipe 1 is positioned within this circulation path. The dual-pipe 1 allows the low-temperature, low-pressure gaseous refrigerant exiting the evaporator to flow along the inner flow path 6... Figure 2 The high-temperature, high-pressure liquid refrigerant flowing from the condenser along the direction of the dashed arrow in the outer flow path 4... Figure 2The solid arrow indicates that the refrigerant flows in the direction opposite to that of the low-temperature, low-pressure gaseous refrigerant flowing in the inner tube 3, and is assembled in the circulation path. As a refrigerant, for example, Freon-based refrigerants such as R-134a and R-1234yf, or CO2 refrigerant, can be used.

[0078] Next, the effect of this example will be explained.

[0079] In this example, in the straight pipe section 11 of the double pipe 1, the inner circumferential surface of the outer pipe 2 is connected to the protrusion 31 of the inner pipe 3, and an outer flow path 4 divided into multiple circumferential points is formed between the outer pipe 2 and the inner pipe 3. In this way, by forming the outer flow path 4 to be divided in the circumferential direction in the straight pipe section 11, the flow rate of the refrigerant flowing in the outer flow path 4 in the straight pipe section 11 can be ensured, thereby improving the heat exchange performance in the straight pipe section 11.

[0080] Furthermore, in the cross-section of the straight pipe section 11 orthogonal to the X direction, when the maximum depth of each outer flow path 4 is set to D [mm] and the circumferential arc length is set to L [mm], the average value of D / L for all outer flow paths 4 is greater than 0.09 and less than 0.20. That is, in the straight pipe section 11, the outer flow paths 4 are formed with a certain degree of reduced depth and increased circumferential arc length. Therefore, in the straight pipe section 11, the contact area between the refrigerant flowing in the outer flow paths 4 and the outer pipe 2 and inner pipe 3 can be increased. As a result, the heat exchange performance in the straight pipe section 11 can also be improved.

[0081] Furthermore, by setting the average D / L value of all outer flow paths 4 in the cross-section of the straight pipe section 11 to a value greater than 0.09, it is possible to suppress the increase in pressure loss of the refrigerant flowing in each outer flow path 4 in the straight pipe section 11. Moreover, by setting the average value to a value less than 0.20 in the cross-section of the straight pipe section 11, it is possible to suppress the increase in pressure loss of the refrigerant flowing in the inner flow path 6 in the straight pipe section 11.

[0082] Furthermore, by pressing the outer tube 2 toward the inner tube 3, the outer tube 2 comes into contact with the protrusion 31. As a result, the outer tube 2 can be easily brought into direct (mechanical) contact with the protrusion 31, which can improve the productivity of the double tube 1 having multiple outer flow paths 4 divided into circumferential sections.

[0083] Furthermore, in the straight pipe section 11, the inner pipe 3 has four protrusions 31 in the circumferential direction. Therefore, compared to the case where the number of protrusions 31 in the straight pipe section 11 is three or less, it is easier to increase the flow velocity in the outer flow path 4 in the straight pipe section 11, and the heat exchange performance can be further improved. In addition, by having four protrusions 31 in the straight pipe section 11, it is possible to prevent the pressure loss of the refrigerant flowing in the outer flow path 4 from becoming too large.

[0084] Furthermore, in the cross-sectional shape of the inner tube 3, multiple protrusions 31 are formed at equal intervals. Therefore, the refrigerant flow rate in the outer flow path 4 is equal in the circumferential direction. As a result, uneven heat exchange between the refrigerant flowing in the outer flow path 4 and the refrigerant flowing in the flow path inside the inner tube 3 is easily suppressed.

[0085] Furthermore, in the straight pipe section 11, the torsion angle of each protrusion 31 is 20° or more and 60° or less. By setting the torsion angle to 20° or more, it is easy to increase the contact area between each outer flow path 4 and the inner pipe 3, thereby easily improving heat exchange performance. By setting the torsion angle to 60° or less, it is possible to prevent the pressure loss of the refrigerant flowing in the outer flow path 4 from becoming too large.

[0086] As described above, according to this example, a dual tube for a heat exchanger can be provided, which optimizes the flow path cross-sectional shape of the outer flow path in the straight tube section, thereby improving the heat exchange performance compared to the past.

[0087] (Example 2)

[0088] like Figure 7 As shown, this example illustrates a case where protrusions 31 are formed at 6 locations on the inner tube 3 within the straight tube section 11. Although not shown, all 6 protrusions 31 abut against the inner circumferential surface of the outer tube 2. In this example, when the maximum depth of each outer flow path 4 is set to D [mm] and the circumferential arc length is set to L [mm], the average value of D / L for all outer flow paths 4 is greater than 0.09 and less than 0.20.

[0089] Everything else is the same as in Example 1.

[0090] Furthermore, unless otherwise specified, any symbols used in this example that are the same as those used in the described embodiments represent the same constituent elements as those in the described embodiments.

[0091] In this example, a protrusion 31 is formed at 6 points in the inner tube 3 within the straight pipe section 11, thus the outer flow path 4 is formed at 6 points in the circumferential direction. Therefore, it is easy to further increase the flow rate of the refrigerant flowing in the outer flow path 4 within the straight pipe section 11. Furthermore, by setting the protrusion 31 at 6 points or less in the straight pipe section 11, it is possible to suppress the situation where the pressure loss of the refrigerant flowing in the outer flow path 4 within the straight pipe section 11 becomes excessive.

[0092] Furthermore, the dual tube in this example has the same effect as in Example 1.

[0093] (Experimental Example 1)

[0094] This example uses simulation to evaluate the degree of heat exchange performance between the refrigerant flowing through the outer flow path 4 and the refrigerant flowing through the inner flow path 6 under various variations of the average D / L value (average D / L value) of all outer flow paths 4 in a dual-tube system 1. In this simulation, Dassault Systèmes' SolidWorks FlowSimulation software was used as the analytical software.

[0095] In this example, samples 1-7, with the same basic structure as in Example 1 but with various modifications to the average D / L value, and sample 8 for comparison were prepared. Samples 1-7 are models, similar to those in Example 1, having protrusions 31 at four locations in the circumferential direction, with each protrusion 31 abutting against the inner circumferential surface of the outer tube 2. The average D / L value of sample 1 is 0.08, that of sample 2 is 0.09, that of sample 3 is 0.10, that of sample 4 is 0.14, that of sample 5 is 0.15, that of sample 6 is 0.16, and that of sample 7 is 0.20. Figure 8 The image shows sample 1 with an average D / L ratio of 0.08. Figure 9 The image shows sample 5, with an average D / L ratio of 0.15. Figure 10 Sample 7, with an average D / L ratio of 0.20, is shown in the figure.

[0096] On the other hand, such as Figure 11 As shown, sample 8 is a model with protrusions 31 at 3 points in the circumferential direction, and the diameter of the outer circle of the inner tube 3 is smaller than the diameter of the inner circle of the outer tube 2, and the inner tube 3 does not abut against the outer tube 2. The maximum radial length M between each spiral recess 32 in sample 8 and the outer circle 97 of the inner tube 3 is assumed to be 1.7 mm. In addition, sample 8 is assumed to be a double tube that has been normally in use.

[0097] In each sample, the outer tube 2 has a smooth tube shape with a circular cross-section, and the inner tube 3 has a protrusion 31 with a twist angle of 40°. Furthermore, as... Figure 12 As shown in the example of sample 5 with an average D / L ratio of 0.15, each sample is formed as a straight line, and the protrusion 31 in the inner tube 3 of each sample is formed throughout the entire X direction. In addition, the diameter of the outer circumference of the outer tube 2 is about 21 mm, the thickness of the outer tube 2 and the inner tube 3 is 1.2 mm, and the length in the X direction is 500 mm.

[0098] And, as Figure 12As shown in the example, a simulation was conducted where each sample was placed in an environment with an external air temperature of 23°C. Refrigerant F1 was introduced into the inner flow path 6 from one side in the X direction, and refrigerant F2 was introduced into the outer flow path 4 from the other side in the X direction. In this simulation, the initial temperature of refrigerant F2 flowing in the outer flow path 4 (temperature at the inlet of the outer flow path 4) was 50°C, and the initial temperature of refrigerant F1 flowing in the inner flow path 6 was 11°C. The flow rates of refrigerant F1 and F2 flowing in the outer flow path 4 and inner flow path 6, respectively, were 234 kg / h. Furthermore, the pressure of refrigerant F2 at the inlet of the outer flow path 4 was 1499 kPa, and the pressure of refrigerant F1 at the inlet of the inner flow path 6 was 199 kPa. The refrigerants F1 and F2 flowing in the outer flow path 4 and inner flow path 6 were assumed to be HFC-134a.

[0099] Then, in each sample, the heat absorbed by the refrigerant F1 flowing in the inner flow path 6 was calculated as an indicator of heat exchange performance. The heat absorbed by the refrigerant F1 flowing in the inner flow path 6 was obtained by multiplying the enthalpy difference by the flow rate of the refrigerant flowing in the inner flow path 6. This enthalpy difference is the difference between the enthalpy calculated based on the pressure and temperature of the refrigerant flowing through the outlet of the inner flow path 6 and the enthalpy calculated based on the pressure and temperature of the refrigerant flowing through the inlet of the inner flow path 6.

[0100] The results are shown in Figure 13 Additionally, in Figure 13 In the diagram, the results for samples 1-7 are plotted with the symbol "〇", and the results for sample 8, which serves as a comparison sample, are plotted with the symbol "◇". Regarding sample 8, the flow path between the outer tube 2 and the inner tube 3 is connected over the entire circumference. Although the average D / L value cannot be defined and calculated using the same method as for samples 1-7, in... Figure 13 For convenience, the results for sample 8 are plotted at the position where the average D / L is 0.15.

[0101] according to Figure 13 It can be seen that, in this example, compared with sample 8, samples 1-7 have a higher heat absorption of the refrigerant flowing in the inner flow path 6, that is, a higher heat exchange performance. Based on these results, it can be seen that in the straight pipe section 11, protrusions 31 are formed at four circumferential locations, and each protrusion 31 abuts against the outer pipe 2, thus forming an outer flow path 4 at four circumferential locations. This results in improved heat exchange performance compared to sample 8, which has protrusions 31 formed at three circumferential locations and does not abut against the outer pipe 2. Furthermore, according to... Figure 13 It can be seen that by setting the average D / L value to above 0.14, the heat exchange performance is further improved.

[0102] (Experimental Example 2)

[0103] This example uses simulation to evaluate the pressure loss of the refrigerant flowing in the outer flow path 4 of samples 1-8 in Experimental Example 1. Similar to Experimental Example 1, the simulation in this example uses "SolidWorks FlowSimulation" manufactured by Dassault Systèmes Ltd. as the analytical software. The construction and calculation conditions of samples 1-8 are the same as in Experimental Example 1.

[0104] In this example, the pressure difference between the refrigerant pressure flowing through the inlet portion of the outer flow path 4 and the refrigerant pressure flowing through the outlet portion of the outer flow path 4 is defined as the pressure loss of the refrigerant in the outer flow path 4.

[0105] The result is Figure 14 As shown in the image.

[0106] from Figure 14 It is known that by setting the average D / L value to a value greater than 0.09, the pressure loss of the refrigerant in the outer flow path 4 can be suppressed to less than 20 kPa. On the other hand, it is known that when the average D / L value is greater than or equal to 0.09, the pressure loss of the refrigerant in the outer flow path 4 increases sharply. Therefore, from the viewpoint of reducing the pressure loss of the refrigerant in the outer flow path 4, it is known that the average D / L value is preferably set to a value greater than 0.10.

[0107] (Experimental Example 3)

[0108] This example evaluates the pressure loss of the refrigerant flowing in the inner flow path 6 of samples 1-8 in Experimental Example 1 through simulation. Similar to Experimental Examples 1 and 2, Dassault Systèmes' SolidWorks FlowSimulation software was used as the analytical software in this simulation. The construction and calculation conditions of samples 1-8 were the same as in Experimental Example 1.

[0109] In this example, the pressure difference between the refrigerant pressure flowing through the inlet portion of the inner flow path 6 and the refrigerant pressure flowing through the outlet portion of the inner flow path 6 is defined as the pressure loss of the refrigerant in the inner flow path 6.

[0110] The result is Figure 15 As shown in the image.

[0111] according to Figure 15 It is known that by setting the average D / L value to less than 0.20, the pressure loss of the refrigerant in the inner flow path 6 can be reduced. Furthermore, it is known that by setting the average D / L value to less than 0.20, the pressure loss of the refrigerant in the inner flow path 6 can be reduced compared to the comparative sample 8. Therefore, from the viewpoint of reducing the pressure loss of the refrigerant in the inner flow path 6, it is preferable to set the average D / L value to less than 0.16.

[0112] This invention is not limited to the embodiments described above, and can be applied to various implementations without departing from its spirit. For example, the method of manufacturing the inner tube 3 having the protrusion 31 and the spiral recess 32 is not limited to the method described in Embodiment 1. For example, the inner tube 3 having the protrusion 31 and the spiral recess 32 can be manufactured by linearly drawing a smooth tube with a circular cross-section as raw material while rotating a mold having an inner hole with a shape corresponding to the spiral recess 32. In this case, the portion where the spiral recess 32 is not formed by drawing, that is, the portion between the circumferential spiral recesses 32 in the inner tube 3, becomes the protrusion 31 formed along the spiral recess 32.

Claims

1. A double-tube heat exchanger having a double-tube structure in which an inner tube is disposed inside an outer tube, for heat exchange between a fluid flowing inside the inner tube and a fluid flowing between the inner tube and the outer tube, wherein, In the straight section of the double tube for the heat exchanger, the inner tube has a plurality of protrusions that bend outwards and extend along the length direction. The protrusion is helically displaced along its length. In the cross-section of the straight tube section orthogonal to its length, the inner circumferential surface of the outer tube is circular. In the straight tube section, the inner circumferential surface of the outer tube is in contact with the protrusion of the inner tube, and an outer flow path divided into multiple circumferential points is formed between the outer tube and the inner tube. In the cross-section of the straight pipe section orthogonal to the length direction, when the maximum depth of each of the outer flow paths is set as D, the circumferential arc length is set as L, and the units of the maximum depth D and the circumferential arc length L of the outer flow paths are mm, the average value of D / L for all the outer flow paths is greater than 0.11 and less than 0.

20. At the boundary between the abutment portion of the protrusion and the outer tube and the outer flow path, the position where the radial length of the gap between the protrusion and the outer tube is 4 μm is defined as the endpoint of the abutment portion. The circumferential length of the circumferential circle between the endpoint of the abutment portion on one side of the outer flow path and the endpoint of the abutment portion on the other side is defined as the circumferential arc length L of the outer flow path. The protrusion is formed as a curved surface that bends along the inner circumferential surface of the outer tube. Around the boundary between the protrusion and the outer tube at the point of contact with the outer tube and the outer flow path, the gap between the protrusion and the outer tube gradually increases radially as it approaches the center of the outer flow path in the circumferential direction. The inner tube has a plurality of protrusions and a plurality of spiral recesses disposed between the protrusions and curved in a manner that protrudes toward the inner circumference. The spiral recesses are spirally displaced in the length direction, and the boundary between the protrusions and the spiral recesses is formed into an angle.

2. The double-tube heat exchanger according to claim 1, wherein, The outer tube is pressed toward the inner tube, thereby abutting the protrusion.

3. The double-tube heat exchanger according to claim 1 or 2, wherein, In the cross-section of the straight pipe section orthogonal to the length direction, the total arc length L of the outer flow path has a length of more than 60% of the circumference C of the virtual circumcircle of the inner pipe.

4. The double-tube heat exchanger according to claim 1 or 2, wherein, In the straight tube section, the number of protrusions provided on the inner tube is 4 or more and 6 or less.

5. The double-tube heat exchanger according to claim 1 or 2, wherein, In the straight tube section, the torsion angle of each of the protrusions is 20° or more and 60° or less.

6. The double-tube heat exchanger according to claim 1 or 2, wherein, In the cross-section of the straight pipe section orthogonal to the length direction, the total arc length L of the outer flow path has a length of less than 90% of the circumference C of the virtual circumcircle of the inner pipe.

7. The double-tube heat exchanger according to claim 1 or 2, wherein, In the cross-section of the straight pipe section orthogonal to the length direction, the total cross-sectional area of ​​the outer flow paths is more than 5% and less than 30% of the cross-sectional area of ​​the inner pipe.

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