US20090166019A1 - Double-wall-tube heat exchanger - Google Patents
Double-wall-tube heat exchanger Download PDFInfo
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- US20090166019A1 US20090166019A1 US12/314,763 US31476308A US2009166019A1 US 20090166019 A1 US20090166019 A1 US 20090166019A1 US 31476308 A US31476308 A US 31476308A US 2009166019 A1 US2009166019 A1 US 2009166019A1
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- tube
- inner tube
- heat exchanger
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- double
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28D—HEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA DO NOT COME INTO DIRECT CONTACT
- F28D7/00—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall
- F28D7/10—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall the conduits being arranged one within the other, e.g. concentrically
- F28D7/106—Heat-exchange apparatus having stationary tubular conduit assemblies for both heat-exchange media, the media being in contact with different sides of a conduit wall the conduits being arranged one within the other, e.g. concentrically consisting of two coaxial conduits or modules of two coaxial conduits
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
- F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
- F25B40/00—Subcoolers, desuperheaters or superheaters
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/006—Tubular elements; Assemblies of tubular elements with variable shape, e.g. with modified tube ends, with different geometrical features
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/12—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element
- F28F1/14—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element and extending longitudinally
- F28F1/16—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element and extending longitudinally the means being integral with the element, e.g. formed by extrusion
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/12—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element
- F28F1/34—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element and extending obliquely
- F28F1/36—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only outside the tubular element and extending obliquely the means being helically wound fins or wire spirals
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/40—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only inside the tubular element
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/42—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being both outside and inside the tubular element
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/42—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being both outside and inside the tubular element
- F28F1/422—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being both outside and inside the tubular element with outside means integral with the tubular element and inside means integral with the tubular element
Definitions
- the present invention relates to a double-wall-tube heat exchanger and, more particularly, to a double-wall-tube heat exchanger which has an outer tube, and an inner tube provided in and spaced apart from the outer tube.
- the term “condenser” refers to not only an ordinary condenser but also a subcooling condenser, which has a condensing section and a subcooling section.
- a conventionally proposed refrigeration system for use in a car air conditioner includes a compressor; a condenser having a condensing section and a subcooling section; an evaporator; an expansion valve serving as a pressure-reducing device; a vapor-liquid separator; and an intermediate heat exchanger disposed between the condenser and the evaporator and adapted to perform heat exchange between a high-temperature refrigerant from the subcooling section of the condenser and a low-temperature refrigerant from the evaporator (as disclosed in, for example, Japanese Patent Application Laid-Open (kokai) No. 2006-162241).
- the refrigerant which has been subcooled in the subcooling section of the condenser is further cooled in the intermediate heat exchanger by the low-temperature, low-pressure refrigerant from the evaporator.
- the cooling performance of the evaporator is improved.
- the intermediate heat exchanger used in the refrigeration system described in the above-mentioned publication has an outer tube, and an inner tube disposed in and spaced apart from the outer tube; the inner tube has grooves which are formed on its outer wall surface, by deforming its wall, in such a manner as to extend in its longitudinal direction; a clearance between the outer tube and the inner tube serves as a high-temperature refrigerant flow path through which the high-temperature refrigerant from the condenser flows; and the interior of the inner tube serves as a low-temperature refrigerant flow path through which the low-temperature refrigerant from the evaporator flows.
- the intermediate heat exchanger described in the above-mentioned publication involves the following problem: the area of heat transfer between the high-temperature refrigerant flow path and the low-temperature refrigerant flow path is small, resulting in insufficient heat exchange performance.
- An object of the present invention is to solve the above-mentioned problem and to provide a double-wall-tube heat exchanger exhibiting excellent heat exchange performance.
- the present invention comprises the following modes.
- a double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube.
- a clearance between the outer tube and the inner tube and the interior of the inner tube serve as respective refrigerant flow paths.
- the inner tube has a plurality of interior fins formed on the inner circumferential surface thereof. The interior fins project radially inward; extend in the longitudinal direction of the inner tube; and are arranged at circumferential intervals.
- the inner tube has a plurality of elongated projections formed on the outer circumferential surface thereof. The elongated projections project radially outward; extend in the longitudinal direction; and are arranged at circumferential intervals.
- the fin height of the interior fins is greater than the projecting height of the elongated projections.
- Par. 2 specifies a radial clearance of 0.4 mm to 1.2 mm inclusive between the inner circumferential surface of the outer tube and the portion of the outer circumferential surface of the inner tube where the elongated projections are not formed, for the following reason. If the radial clearance is excessively small, pressure loss increases sharply in the refrigerant flow path formed between the inner tube and the outer tube. On the other hand, if the radial clearance is excessively large, the flow velocity of a refrigerant drops in the refrigerant flow path formed between the inner tube and the outer tube, potentially resulting in a drop in heat transfer coefficient.
- Par. 3 specifies a clearance of 0 mm to 0.5 mm inclusive between the projecting ends of the elongated projections of the inner tube and the inner circumferential surface of the outer tube, for the following reason. If the clearance is excessively large, in the case where the double-wall-tube heat exchanger has a bend(s), wrinkles are apt to be formed on the outer tube in a bending process.
- a double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube.
- a clearance between the outer tube and the inner tube and the interior of the inner tube serves as respective refrigerant flow paths.
- the inner tube has a plurality of interior fins formed on the inner circumferential surface thereof. The interior fins project radially inward; extend in the longitudinal direction of the inner tube; and are arranged at circumferential intervals.
- the outer tube has a plurality of elongated projections formed on the inner circumferential surface thereof. The elongated projections project radially inward; extend in the longitudinal direction of the outer tube; and are arranged at circumferential intervals.
- Par. 5 specifies a radial clearance of 0.4 mm to 1.2 mm inclusive between the portion of the inner circumferential surface of the outer tube where the elongated projections are not formed and the outer circumferential surface of the inner tube, for the following reason. If the radial clearance is excessively small, pressure loss increases sharply in the refrigerant flow path formed between the inner tube and the outer tube. On the other hand, if the radial clearance is excessively large, the flow velocity of a refrigerant drops in the refrigerant flow path formed between the inner tube and the outer tube, potentially resulting in a drop in heat transfer coefficient.
- Par. 6 specifies a clearance of 0 mm to 0.5 mm inclusive between the projecting ends of the elongated projections of the outer tube and the outer circumferential surface of the inner tube, for the following reason. If the clearance is excessively large, in the case where the double-wall-tube heat exchanger has a bend(s), wrinkles are apt to be formed on the outer tube in a bending process.
- Par. 7 specifies a fin thickness of 0.2 mm to 2.0 mm inclusive for the interior fins of the inner tube, for the following reason. If the fin thickness is excessively thin, the fin efficiency of the interior fins drops, and working may become difficult. If the fin thickness is in excess of 2.0 mm, the effect of improving the fin efficiency of the interior fins is impaired, and working may become difficult.
- the fin thickness of the interior fins of the inner tube is more preferably 0.3 mm to 0.7 mm inclusive.
- Par. 8 specifies a fin height of 1.0 mm to 3.0 mm inclusive for the interior fins of the inner tube, for the following reason. If the fin height is excessively low, the area of heat transfer between the inner tube and a refrigerant which flows through the refrigerant flow path in the inner tube fails to become sufficiently large; as a result, heat transfer performance fails to be sufficiently improved. If the fin height is excessively high, in the case where the double-wall-tube heat exchanger has a bend(s), the interior fins may be buckled in a bending process, potentially blocking the refrigerant flow path in the inner tube.
- Par. 9 specifies an inside diameter of 12 mm or greater for the inner tube, for the following reason. If the inside diameter of the inner tube is excessively small, pressure loss in the refrigerant flow path in the inner tube increases sharply.
- the interior fins of the inner tube may be arranged at a fin pitch of 2 mm or greater as measured at roots of the interior fins, for the following reason. If the fin pitch is excessively small, pressure loss in the refrigerant flow path in the inner tube increases sharply. Particularly, in the case where the double-wall-tube heat exchanger has a bend(s), the interior fins come into contact with one another in a bending process, causing a sharp increase in pressure loss in the refrigerant flow path in the inner tube.
- the inner tube may have a wall thickness of 0.2 mm to 2.0 mm inclusive, for the following reason. If the wall thickness is excessively thin, strength becomes insufficient. If the wall thickness is excessively thick, weight increases, and cost increases as well.
- the inner tube has a plurality of the interior fins formed on the inner circumferential surface thereof, the interior fins projecting radially inward, extending in the longitudinal direction of the inner tube, and being arranged at circumferential intervals;
- the inner tube has a plurality of the elongated projections formed on the outer circumferential surface thereof, the elongated projections projecting radially outward, extending in the longitudinal direction, and being arranged at circumferential intervals; and the fin height of the interior fins is greater than the projecting height of the elongated projections.
- the area of heat transfer between the refrigerant flow path formed between the inner and outer tubes and the refrigerant flow path in the inner tube is greater, so that heat exchange performance is improved.
- the double-wall-tube heat exchanger of par. 1) is used as an intermediate heat exchanger of a refrigeration system described in the above-mentioned publication, a vapor-phase refrigerant, whose heat transfer coefficient is relatively low, flows through the refrigerant flow path in the inner tube.
- the interior fins function to increase the area of heat transfer in the refrigerant flow path in the inner tube through which the vapor-phase refrigerant flows, the performance of the double-wall-tube heat exchanger is improved.
- the elongated projections function to prevent crushing of the refrigerant flow path formed between the inner tube and the outer tube.
- the inner tube has a plurality of the interior fins formed on the inner circumferential surface thereof, the interior fins projecting radially inward, extending in the longitudinal direction of the inner tube, and being arranged at circumferential intervals; and the outer tube has a plurality of the elongated projections formed on the inner circumferential surface thereof, the elongated projections projecting radially inward, extending in the longitudinal direction of the outer tube, and being arranged at circumferential intervals.
- the area of heat transfer between the refrigerant flow path formed between the inner and outer tubes and the refrigerant flow path in the inner tube is greater, so that heat exchange performance is improved.
- the elongated projections function to prevent crushing of the refrigerant flow path formed between the inner tube and the outer tube.
- the outer surface of the inner tube assumes the form of a smooth cylindrical surface, joining work can be facilitated for a refrigerant inflow pipe, a refrigerant outflow pipe, a joint member, etc.
- an increase in pressure loss in the refrigerant flow path formed between the inner and outer tubes can be prevented, and the flow velocity of a refrigerant in the refrigerant flow path formed between the inner and outer tubes increases, whereby heat transfer coefficient is improved, resulting in improvement of the performance of the double-wall-tube heat exchanger.
- the fin efficiency of the interior fins of the inner tube is improved, whereby heat exchange performance is improved.
- the area of heat transfer between the inner tube and a refrigerant which flows through the refrigerant flow path in the inner tube becomes sufficiently large, whereby the performance of heat transfer between the refrigerant and the inner tube is sufficiently improved.
- the double-wall-tube heat exchanger has a bend(s)
- there is prevented blocking of the refrigerant flow path in the inner tube which could otherwise result from the interior fins being buckled in a bending process.
- an increase in pressure loss in the refrigerant flow path in the inner tube can be restrained.
- FIG. 1 is a partially cutaway front view showing the configuration of a double-wall-tube heat exchanger according to Embodiment 1 of the present invention with a longitudinally intermediate portion omitted;
- FIG. 2 is an enlarged fragmentary view of FIG. 1 ;
- FIG. 3 is a sectional view taken along line A-A of FIG. 2 ;
- FIG. 4 is an enlarged fragmentary view of FIG. 3 ;
- FIG. 5 is a sectional view taken along line BPB of FIG. 2 ;
- FIG. 7 is a diagram showing a refrigeration system which uses the double-wall-tube heat exchanger of Embodiment 1 as an intermediate heat exchanger;
- FIG. 8 is a graph showing the relation between the fin thickness and the fin efficiency of the interior fins
- FIG. 9 is a graph showing the relation of the number of fins and the fin pitch to exchanged heat quantity and pressure loss
- FIG. 10 is a graph showing the relation of the inside diameter of the inner tube to pressure loss and overall heat transfer coefficient
- FIG. 11 is a graph showing the relation of liquid flow path width to pressure loss and overall heat transfer coefficient
- FIG. 12 is a perspective, fragmentary view showing a modified inner tube of the double-wall-tube heat exchanger of Embodiment 1;
- FIG. 13 is an equivalent view of FIG. 2 , showing a double-wall-tube heat exchanger according to Embodiment 2 of the present invention
- FIG. 14 is a sectional view taken along line C-C of FIG. 13 ;
- FIG. 15 is an enlarged fragmentary view of FIG. 14 ;
- FIG. 16 is a sectional view taken along line D-D of FIG. 13 ;
- FIG. 17 is an enlarged fragmentary view of FIG. 16 ;
- FIG. 18 is an equivalent view of FIG. 13 , showing a modified inner tube of the double-wall-tube heat exchanger of Embodiment 2.
- aluminum encompasses aluminum alloys in addition to pure aluminum.
- FIGS. 1 to 7 The present embodiment is shown in FIGS. 1 to 7 .
- a double-wall-tube heat exchanger 1 includes an outer tube 2 and an inner tube 3 .
- the outer tube 2 is formed of an aluminum extrudate having a circular cross section.
- the inner tube 3 is formed of an aluminum extrudate having a circular cross section and is inserted concentrically into and spaced apart from the outer tube 2 .
- a clearance between the outer tube 2 and the inner tube 3 serves as a first refrigerant flow path 4 .
- the interior of the inner tube 3 serves as a second refrigerant flow path 5 .
- Portions of the outer tube 2 which are located longitudinally outward of the expanded portions 6 and 7 are subjected to roller working which is performed from the radial outside toward the radial inside along the entire circumference, thereby forming diameter-reduced portions 14 .
- the diameter-reduced portions 14 are brazed to the inner tube 3 at positions located toward the opposite ends of the inner tube 3 .
- the diameter-reduced portions 14 are formed after the inner tube 3 is placed within the outer tube 2 . In the course of formation of the diameter-reduced portions 14 , associated portions of the elongated projections 13 of the inner tube 3 are crushed and bite into the inner circumferential surfaces of the diameter-reduced portions 14 (see FIGS. 5 and 6 ).
- the expanded pipe portion 15 a of the vapor-phase refrigerant inflow pipe 15 and the expanded pipe portion 16 a of the vapor-phase refrigerant outflow pipe 16 may be pressed from the radial outside so as to establish a condition in which associated portions of the elongated projections 13 are crushed and bite into the inner circumferential surfaces of the expanded pipe portions 15 a and 16 a; as a result, the clearance between the inner circumferential surfaces of the expanded pipe portions 15 a and 16 a and a portion of the outer circumferential surface of the inner tube 3 where the elongated projections 13 are not formed is reduced to such an extent as to be filled with the brazing material.
- brazing between the inner tube 3 and each of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 is carried out simultaneously with brazing between the outer tube 2 and the inner tube 3 while an appropriate gap is maintained between the opposite ends of the outer tube 2 and the corresponding ends of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 .
- FIG. 7 shows a refrigeration system in which the above-described double-wall-tube heat exchanger 1 is used as an intermediate heat exchanger.
- the refrigeration system uses, for example, a chlorofluorocarbon-based refrigerant.
- the refrigeration system includes a compressor 20 ; a condenser 21 having a condensing section 22 , a liquid receiver 23 serving as a vapor-liquid separator, and a subcooling section 24 ; an evaporator 25 ; an expansion valve 26 serving as a pressure-reducing device; and the double-wall-tube heat exchanger 1 which serves as an intermediate heat exchanger for performing heat exchange between a refrigerant from the condenser 20 and a refrigerant from the evaporator 25 .
- Piping extending from the subcooling section 24 of the condenser 20 is connected to the liquid-phase refrigerant inflow pipe 9 connected to the outer tube 2 of the double-wall-tube heat exchanger 1 .
- piping extending to the expansion valve 26 is connected to the liquid-phase refrigerant outflow pipe 11 connected to the outer tube 2 .
- piping extending from the evaporator 25 is connected to the vapor-phase refrigerant inflow pipe 15 connected to the inner tube 3 of the double-wall-tube heat exchanger 1 .
- piping extending to the compressor 20 is connected to the vapor-phase refrigerant outflow pipe 16 connected to the inner pipe 3 .
- the refrigeration system is mounted in a vehicle; for example, an automobile, as a car air conditioner.
- a high-temperature, high-pressure vapor-liquid mixed-phase refrigerant which has undergone compression in the compressor 20 , is cooled and condensed in the condensing section 22 of the condenser 21 .
- the refrigerant flows into the liquid receiver 23 and is separated into two phases; namely, the vapor phase and the liquid phase.
- the resultant liquid-phase refrigerant flows into the subcooling section 24 and is subcooled.
- the subcooled liquid-phase refrigerant flows through the liquid-phase refrigerant inflow pipe 9 and flows into the first refrigerant flow path 4 of the double-wall-tube heat exchanger 1 .
- the liquid-phase refrigerant dividedly flows into all the channels formed between the adjacent elongated projections 13 in the first refrigerant flow path 4 .
- a vapor-phase refrigerant from the evaporator 25 passes through the vapor-phase refrigerant inflow pipe 15 and flows into the second refrigerant flow path 5 of the double-wall-tube heat exchanger 1 .
- the liquid-phase refrigerant is further cooled by the vapor-phase refrigerant whose temperature is relatively low and which flows through the second refrigerant flow path 5 .
- the liquid-phase refrigerant flows to the expansion valve 26 through the liquid-phase refrigerant outflow pipe 11 .
- the expansion valve 26 the liquid-phase refrigerant is adiabatically expanded and is thereby pressure-reduced.
- the two-phase refrigerant flows into the evaporator 25 and is evaporated in the evaporator 25 .
- the vapor-phase refrigerant flows to the compressor 20 through the vapor-phase refrigerant outflow pipe 16 .
- the fin thickness T 1 of the interior fins 12 of the inner tube 3 is preferably 0.2 mm to 2.0 mm inclusive. This has been obtained from the results of computer simulation calculation. Specifically, the relation between the fin thickness T 1 and fin efficiency of the interior fins 12 was obtained by computer simulation calculation which was conducted while the fin thickness T 1 and the fin height H 1 of the interior fins 12 were varied under the following condition: the heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 in the double-wall-tube heat exchanger 1 was set to 380 W/m 2 ⁇ K.
- FIG. 8 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 8 that a fin thickness T 1 of the interior fins 12 of 0.2 mm to 2.0 mm inclusive is preferred. As is apparent from FIG. 8 , when the fin thickness T 1 of the interior fins 12 is less than 0.2 mm, fin efficiency drops sharply. Also, when the fin thickness T 1 is in excess of 1.2 mm, the effect of improving fin efficiency is saturated.
- the fin height H 1 of the interior fins 12 of the inner tube 3 is 1.0 mm to 3.0 mm inclusive. This also has been derived from the results shown in FIG. 8 .
- FIG. 9 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 9 that a fin pitch P 1 of the interior fins 12 as measured at roots of the interior fins 12 of 2.0 mm or greater is preferred. As is apparent from FIG. 9 , when the fin pitch P 1 of the interior fins 12 is less than 2 mm, pressure loss in the second refrigerant flow path 5 increases. In FIG. 9 , exchanged heat quantity and pressure loss are expressed in percentage with those of the case where an inner tube not having the interior fins is used, being each taken as 100%.
- the inside diameter D of the inner tube 3 is 12 mm or greater. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the inside diameter D of the inner tube 3 to pressure loss and the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 was obtained by computer simulation calculation which was conducted while the inside diameter D of the inner tube 3 was varied under the following conditions: the fin thickness T 1 of the interior fins 12 of the inner tube 3 was set to 0.5 mm; the fin height H 1 of the interior fins 12 was set to 1.5 mm; the fin pitch P 1 of the interior fins 12 as measured at roots of the interior fins 12 was set to 2.5 mm; and the temperature and pressure of the vapor-phase refrigerant as measured at the inlet of the second refrigerant flow path 5 were set to 5.0° C.
- FIG. 10 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 10 that an inside diameter D of the inner tube 3 of 12 mm or greater is preferred. As is apparent from FIG. 10 , when the inside diameter D of the inner tube 3 is less than 12 mm, pressure loss increases sharply. Preferably, the upper limit of the inside diameter D of the inner tube 3 is 18 mm. This is because, as is apparent from FIG. 10 , when the inside diameter D of the inner tube 3 is in excess of 18 mm, the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 drops. In FIG.
- pressure loss and the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the second refrigerant flow path 5 to the inner circumferential surface of the inner tube 3 are expressed in percentage with those of the case where an inner tube having an inside diameter of 12 mm is used, being each taken as 100%.
- the liquid flow path width W is 0.4 mm to 1.2 mm inclusive. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the liquid flow path width W to pressure loss and the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first refrigerant flow path 4 to the outer circumferential surface of the inner tube 3 was obtained for the cases of the inside diameter D of the inner tube 3 being 12 mm and 18 mm by computer simulation calculation which was conducted while the liquid flow path width W was varied under the following conditions: the elongated-projection thickness T 2 of the elongated projections 13 of the inner tube 3 was set to 0.5 mm; the pitch P 2 of the elongated projections 13 as measured at roots of the elongated projections 13 was set to 3 mm; the inside diameter D of the inner tube 3 was set to 12 mm and 18 mm; the wall thickness of the inner tube 3 was set to 1.2 mm; and the temperature and pressure of the liquid-phase refrigerant as measured at the
- FIG. 11 shows the results of the computer simulation calculation. It has been derived from the results shown in FIG. 11 that the a liquid flow path width W of 0.4 mm to 1.2 mm inclusive is preferred. As is apparent from FIG. 11 , when the liquid flow path width W is less than 0.4 mm, pressure loss increases sharply. Also, when the liquid flow path width W is in excess of 1.2 mm, the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first refrigerant flow path 4 to the outer circumferential surface of the inner tube 3 drops. In FIG.
- pressure loss and the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first refrigerant flow path 4 to the outer circumferential surface of the inner tube 3 are expressed in percentage with those of the case where the inside diameter D of the inner tube 3 is 12 mm, and the liquid flow path width W is 0.4 mm, being each taken as 100%.
- FIG. 12 shows a modified inner tube of the double-wall-tube heat exchanger of Embodiment 1.
- An inner tube 30 shown in FIG. 12 is twisted about its axis, so that the interior fins 12 and the elongated projections 13 are spiral.
- FIGS. 13 to 17 The present embodiment is shown in FIGS. 13 to 17 .
- FIGS. 13 to 17 show the configurations of essential portions of a double-wall-tube heat exchanger according to Embodiment 2 of the present invention.
- the outer tube 2 has a plurality of elongated projections 32 formed integrally with the inner circumferential surface of the outer tube 2 .
- the elongated projections 32 project radially inward, extend in the longitudinal direction, and are arranged at equal circumferential intervals.
- the outer tube 2 does not have expanded tube portions at opposite end portions thereof.
- the inner tube 3 has diameter-reduced tube portions 33 located slightly longitudinally inward of the opposite ends thereof.
- a refrigerant inlet (not shown) is formed in the wall of the outer tube 2 at a position corresponding to one diameter-reduced tube portion (not shown).
- the refrigerant outlet 8 is formed in the wall of the outer tube 2 at a position corresponding to the other diameter-reduced tube portion 33 .
- An end portion of a liquid-phase refrigerant inflow pipe (not shown) made of aluminum is inserted into the refrigerant inlet and is brazed to the outer tube 2 .
- An end portion of the liquid-phase refrigerant outflow pipe 11 made of aluminum is inserted into the refrigerant outlet 8 and is brazed to the outer tube 2 .
- Elongated projections are not formed on the outer circumferential surface of the inner tube 3 .
- Portions of the outer tube 2 which are located longitudinally outward of the diameter-reduced tube portions 33 of the inner tube 3 are subjected to roller working which is performed from the radial outside toward the radial inside along the entire circumference, thereby forming the diameter-reduced portions 14 .
- the diameter-reduced portions 14 are brazed to portions of the inner tube 3 near the opposite ends thereof.
- the diameter-reduced portions 14 are formed after the inner tube 3 is placed within the outer tube 2 .
- associated portions of the elongated projections 32 of the outer tube 2 are crushed and bite into the outer circumferential surfaces of the inner tube 3 (see FIGS. 16 and 17 ).
- the double-wall-tube heat exchanger 31 of Embodiment 2 does not require an operation of cutting off relevant portions of the elongated projections 32 from the outer tube 2 and an operation of pressing the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 from the radial outside.
- brazing between the inner tube 3 and each of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 is carried out simultaneously with brazing between the outer tube 2 and the inner tube 3 while an appropriate gap is maintained between the opposite ends of the outer tube 2 and the corresponding ends of the expanded pipe portions 15 a and 16 a of the vapor-phase refrigerant inflow and outflow pipes 15 and 16 .
- the double-wall-tube heat exchanger 31 of Embodiment 2 is incorporated into the refrigeration system shown in FIG. 7 in a manner similar to that of the double-wall-tube heat exchanger 1 of Embodiment 1.
- the liquid-phase refrigerant flows through the liquid-phase refrigerant inflow pipe and flows into the first refrigerant flow path 4 of the double-wall-tube heat exchanger 31 , by the effect of the unillustrated one diameter-reduced tube portion of the inner tube 3 , the liquid-phase refrigerant dividedly flows into all the channels formed between the adjacent elongated projections 32 in the first refrigerant flow path 4 . Having passed through all the channels formed between the adjacent elongated projections 32 in the first refrigerant flow path 4 of the double-wall-tube heat exchanger 31 , flows of the liquid-phase refrigerant join together in the other diameter-reduced tube portion 33 . The resultant liquid-phase refrigerant flows to the expansion valve 26 through the liquid-phase refrigerant outflow pipe 11 .
- the double-wall-tube heat exchanger 31 of Embodiment 2 may also use the inner tube 3 which is twisted about its axis.
- FIG. 18 shows a modified inner tube of the double-wall-tube heat exchanger of Embodiment 2.
- Opposite end portions of an inner tube 35 shown in FIG. 18 are extended, so that the vapor-phase refrigerant inflow pipe and the vapor-phase refrigerant outflow pipe are not used.
- Piping extending from the evaporator 25 is connected to an end portion of the inner tube 35 which is located on a side toward the refrigerant outlet 8 .
- Piping extending to the compressor 20 is connected to the other end portion of the inner tube 35 .
- the elongated projections 13 are formed on the outer circumferential surface of the inner tube 3 .
- the elongated projections 32 are formed on the inner circumferential surface of the outer tube 2 .
- an embodiment may be such that the elongated projections 13 and the elongated projections 32 are formed on the outer circumferential surface of the inner tube 3 and the inner circumferential surface of the outer tube 2 , respectively.
- the elongated projections 13 of the inner tube and the elongated projections 32 of the outer tube 2 are arranged in a circumferentially staggered manner.
Abstract
Description
- The present invention relates to a double-wall-tube heat exchanger and, more particularly, to a double-wall-tube heat exchanger which has an outer tube, and an inner tube provided in and spaced apart from the outer tube.
- Herein, the term “condenser” refers to not only an ordinary condenser but also a subcooling condenser, which has a condensing section and a subcooling section.
- A conventionally proposed refrigeration system for use in a car air conditioner includes a compressor; a condenser having a condensing section and a subcooling section; an evaporator; an expansion valve serving as a pressure-reducing device; a vapor-liquid separator; and an intermediate heat exchanger disposed between the condenser and the evaporator and adapted to perform heat exchange between a high-temperature refrigerant from the subcooling section of the condenser and a low-temperature refrigerant from the evaporator (as disclosed in, for example, Japanese Patent Application Laid-Open (kokai) No. 2006-162241). In the refrigeration system described in the publication, the refrigerant which has been subcooled in the subcooling section of the condenser is further cooled in the intermediate heat exchanger by the low-temperature, low-pressure refrigerant from the evaporator. By this procedure, the cooling performance of the evaporator is improved.
- The intermediate heat exchanger used in the refrigeration system described in the above-mentioned publication has an outer tube, and an inner tube disposed in and spaced apart from the outer tube; the inner tube has grooves which are formed on its outer wall surface, by deforming its wall, in such a manner as to extend in its longitudinal direction; a clearance between the outer tube and the inner tube serves as a high-temperature refrigerant flow path through which the high-temperature refrigerant from the condenser flows; and the interior of the inner tube serves as a low-temperature refrigerant flow path through which the low-temperature refrigerant from the evaporator flows.
- However, the intermediate heat exchanger described in the above-mentioned publication involves the following problem: the area of heat transfer between the high-temperature refrigerant flow path and the low-temperature refrigerant flow path is small, resulting in insufficient heat exchange performance.
- An object of the present invention is to solve the above-mentioned problem and to provide a double-wall-tube heat exchanger exhibiting excellent heat exchange performance.
- To achieve the above object, the present invention comprises the following modes.
- 1) A double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube. A clearance between the outer tube and the inner tube and the interior of the inner tube serve as respective refrigerant flow paths. The inner tube has a plurality of interior fins formed on the inner circumferential surface thereof. The interior fins project radially inward; extend in the longitudinal direction of the inner tube; and are arranged at circumferential intervals. The inner tube has a plurality of elongated projections formed on the outer circumferential surface thereof. The elongated projections project radially outward; extend in the longitudinal direction; and are arranged at circumferential intervals. The fin height of the interior fins is greater than the projecting height of the elongated projections.
- 2) A double-wall-tube heat exchanger according to par. 1), wherein a radial clearance between the inner circumferential surface of the outer tube and a portion of the outer circumferential surface of the inner tube where the elongated projections are not formed is 0.4 mm to 1.2 mm inclusive.
- Par. 2) specifies a radial clearance of 0.4 mm to 1.2 mm inclusive between the inner circumferential surface of the outer tube and the portion of the outer circumferential surface of the inner tube where the elongated projections are not formed, for the following reason. If the radial clearance is excessively small, pressure loss increases sharply in the refrigerant flow path formed between the inner tube and the outer tube. On the other hand, if the radial clearance is excessively large, the flow velocity of a refrigerant drops in the refrigerant flow path formed between the inner tube and the outer tube, potentially resulting in a drop in heat transfer coefficient.
- 3) A double-wall-tube heat exchanger according to par. 1), wherein a clearance between projecting ends of the elongated projections of the inner tube and the inner circumferential surface of the outer tube is 0 mm to 0.5 mm inclusive.
- Par. 3) specifies a clearance of 0 mm to 0.5 mm inclusive between the projecting ends of the elongated projections of the inner tube and the inner circumferential surface of the outer tube, for the following reason. If the clearance is excessively large, in the case where the double-wall-tube heat exchanger has a bend(s), wrinkles are apt to be formed on the outer tube in a bending process.
- 4) A double-wall-tube heat exchanger comprising an outer tube, and an inner tube disposed in and spaced apart from the outer tube. A clearance between the outer tube and the inner tube and the interior of the inner tube serves as respective refrigerant flow paths. The inner tube has a plurality of interior fins formed on the inner circumferential surface thereof. The interior fins project radially inward; extend in the longitudinal direction of the inner tube; and are arranged at circumferential intervals. The outer tube has a plurality of elongated projections formed on the inner circumferential surface thereof. The elongated projections project radially inward; extend in the longitudinal direction of the outer tube; and are arranged at circumferential intervals.
- 5) A double-wall-tube heat exchanger according to par. 4), wherein a radial clearance between a portion of the inner circumferential surface of the outer tube where the elongated projections are not formed and the outer circumferential surface of the inner tube is 0.4 mm to 1.2 mm inclusive.
- Par. 5) specifies a radial clearance of 0.4 mm to 1.2 mm inclusive between the portion of the inner circumferential surface of the outer tube where the elongated projections are not formed and the outer circumferential surface of the inner tube, for the following reason. If the radial clearance is excessively small, pressure loss increases sharply in the refrigerant flow path formed between the inner tube and the outer tube. On the other hand, if the radial clearance is excessively large, the flow velocity of a refrigerant drops in the refrigerant flow path formed between the inner tube and the outer tube, potentially resulting in a drop in heat transfer coefficient.
- 6) A double-wall-tube heat exchanger according to par. 4), wherein a clearance between projecting ends of the elongated projections of the outer tube and the outer circumferential surface of the inner tube is 0 mm to 0.5 mm inclusive.
- Par. 6) specifies a clearance of 0 mm to 0.5 mm inclusive between the projecting ends of the elongated projections of the outer tube and the outer circumferential surface of the inner tube, for the following reason. If the clearance is excessively large, in the case where the double-wall-tube heat exchanger has a bend(s), wrinkles are apt to be formed on the outer tube in a bending process.
- 7) A double-wall-tube heat exchanger according to par. 1) or 4), wherein the interior fins of the inner tube have a fin thickness of 0.2 mm to 2.0 mm inclusive.
- Par. 7) specifies a fin thickness of 0.2 mm to 2.0 mm inclusive for the interior fins of the inner tube, for the following reason. If the fin thickness is excessively thin, the fin efficiency of the interior fins drops, and working may become difficult. If the fin thickness is in excess of 2.0 mm, the effect of improving the fin efficiency of the interior fins is impaired, and working may become difficult. In consideration of extrusion workability in forming the inner tube by extrusion, and bending workability in the case where the double-wall-tube heat exchanger has a bend(s), the fin thickness of the interior fins of the inner tube is more preferably 0.3 mm to 0.7 mm inclusive.
- 8) A double-wall-tube heat exchanger according to par. 1) or 4), wherein the interior fins have a fin height of 1.0 mm to 3.0 mm inclusive.
- Par. 8) specifies a fin height of 1.0 mm to 3.0 mm inclusive for the interior fins of the inner tube, for the following reason. If the fin height is excessively low, the area of heat transfer between the inner tube and a refrigerant which flows through the refrigerant flow path in the inner tube fails to become sufficiently large; as a result, heat transfer performance fails to be sufficiently improved. If the fin height is excessively high, in the case where the double-wall-tube heat exchanger has a bend(s), the interior fins may be buckled in a bending process, potentially blocking the refrigerant flow path in the inner tube.
- 9) A double-wall-tube heat exchanger according to par. 1) or 4), wherein the inner tube has an inside diameter of 12 mm or greater.
- Par. 9) specifies an inside diameter of 12 mm or greater for the inner tube, for the following reason. If the inside diameter of the inner tube is excessively small, pressure loss in the refrigerant flow path in the inner tube increases sharply.
- In the double-wall-tube heat exchanger according to any one of pars. 1) to 9), the interior fins of the inner tube may be arranged at a fin pitch of 2 mm or greater as measured at roots of the interior fins, for the following reason. If the fin pitch is excessively small, pressure loss in the refrigerant flow path in the inner tube increases sharply. Particularly, in the case where the double-wall-tube heat exchanger has a bend(s), the interior fins come into contact with one another in a bending process, causing a sharp increase in pressure loss in the refrigerant flow path in the inner tube.
- In the double-wall-tube heat exchanger according to any one of pars. 1) to 9), the inner tube may have a wall thickness of 0.2 mm to 2.0 mm inclusive, for the following reason. If the wall thickness is excessively thin, strength becomes insufficient. If the wall thickness is excessively thick, weight increases, and cost increases as well.
- According to the double-wall-tube heat exchanger of par. 1), the inner tube has a plurality of the interior fins formed on the inner circumferential surface thereof, the interior fins projecting radially inward, extending in the longitudinal direction of the inner tube, and being arranged at circumferential intervals; the inner tube has a plurality of the elongated projections formed on the outer circumferential surface thereof, the elongated projections projecting radially outward, extending in the longitudinal direction, and being arranged at circumferential intervals; and the fin height of the interior fins is greater than the projecting height of the elongated projections. Thus, as compared with the double-wall-tube heat exchanger described in the above-mentioned publication, the area of heat transfer between the refrigerant flow path formed between the inner and outer tubes and the refrigerant flow path in the inner tube is greater, so that heat exchange performance is improved. In the case where the double-wall-tube heat exchanger of par. 1) is used as an intermediate heat exchanger of a refrigeration system described in the above-mentioned publication, a vapor-phase refrigerant, whose heat transfer coefficient is relatively low, flows through the refrigerant flow path in the inner tube. Since the interior fins function to increase the area of heat transfer in the refrigerant flow path in the inner tube through which the vapor-phase refrigerant flows, the performance of the double-wall-tube heat exchanger is improved. In the case where the double-wall-tube heat exchanger has a bend(s), the elongated projections function to prevent crushing of the refrigerant flow path formed between the inner tube and the outer tube.
- According to the double-wall-tube heat exchanger of par. 2), an increase in pressure loss in the refrigerant flow path formed between the inner and outer tubes can be prevented, and the flow velocity of a refrigerant in the refrigerant flow path formed between the inner and outer tubes increases, whereby heat transfer coefficient is improved, resulting in improvement of the performance of the double-wall-tube heat exchanger.
- According to the double-wall-tube heat exchanger of par. 3), in the case where the double-wall-tube heat exchanger has a bend(s), formation of wrinkles on the outer tube in a bending process can be reliably prevented.
- According to the double-wall-tube heat exchanger of par. 4), the inner tube has a plurality of the interior fins formed on the inner circumferential surface thereof, the interior fins projecting radially inward, extending in the longitudinal direction of the inner tube, and being arranged at circumferential intervals; and the outer tube has a plurality of the elongated projections formed on the inner circumferential surface thereof, the elongated projections projecting radially inward, extending in the longitudinal direction of the outer tube, and being arranged at circumferential intervals. Thus, as compared with the double-wall-tube heat exchanger described in the above-mentioned publication, the area of heat transfer between the refrigerant flow path formed between the inner and outer tubes and the refrigerant flow path in the inner tube is greater, so that heat exchange performance is improved. Also, in the case where the double-wall-tube heat exchanger has a bend(s), the elongated projections function to prevent crushing of the refrigerant flow path formed between the inner tube and the outer tube. Furthermore, since the outer surface of the inner tube assumes the form of a smooth cylindrical surface, joining work can be facilitated for a refrigerant inflow pipe, a refrigerant outflow pipe, a joint member, etc.
- According to the double-wall-tube heat exchanger of par. 5), an increase in pressure loss in the refrigerant flow path formed between the inner and outer tubes can be prevented, and the flow velocity of a refrigerant in the refrigerant flow path formed between the inner and outer tubes increases, whereby heat transfer coefficient is improved, resulting in improvement of the performance of the double-wall-tube heat exchanger.
- According to the double-wall-tube heat exchanger of par. 6), in the case where the double-wall-tube heat exchanger has a bend(s), formation of wrinkles on the outer tube in a bending process can be reliably prevented.
- According to the double-wall-tube heat exchanger of par. 7), the fin efficiency of the interior fins of the inner tube is improved, whereby heat exchange performance is improved.
- According to the double-wall-tube heat exchanger of par. 8), the area of heat transfer between the inner tube and a refrigerant which flows through the refrigerant flow path in the inner tube becomes sufficiently large, whereby the performance of heat transfer between the refrigerant and the inner tube is sufficiently improved. In the case where the double-wall-tube heat exchanger has a bend(s), there is prevented blocking of the refrigerant flow path in the inner tube which could otherwise result from the interior fins being buckled in a bending process.
- According to the double-wall-tube heat exchanger of par. 9), an increase in pressure loss in the refrigerant flow path in the inner tube can be restrained.
-
FIG. 1 is a partially cutaway front view showing the configuration of a double-wall-tube heat exchanger according toEmbodiment 1 of the present invention with a longitudinally intermediate portion omitted; -
FIG. 2 is an enlarged fragmentary view ofFIG. 1 ; -
FIG. 3 is a sectional view taken along line A-A ofFIG. 2 ; -
FIG. 4 is an enlarged fragmentary view ofFIG. 3 ; -
FIG. 5 is a sectional view taken along line BPB ofFIG. 2 ; -
FIG. 6 is an enlarged fragmentary view ofFIG. 5 ; -
FIG. 7 is a diagram showing a refrigeration system which uses the double-wall-tube heat exchanger ofEmbodiment 1 as an intermediate heat exchanger; -
FIG. 8 is a graph showing the relation between the fin thickness and the fin efficiency of the interior fins; -
FIG. 9 is a graph showing the relation of the number of fins and the fin pitch to exchanged heat quantity and pressure loss; -
FIG. 10 is a graph showing the relation of the inside diameter of the inner tube to pressure loss and overall heat transfer coefficient; -
FIG. 11 is a graph showing the relation of liquid flow path width to pressure loss and overall heat transfer coefficient; -
FIG. 12 is a perspective, fragmentary view showing a modified inner tube of the double-wall-tube heat exchanger ofEmbodiment 1; -
FIG. 13 is an equivalent view ofFIG. 2 , showing a double-wall-tube heat exchanger according toEmbodiment 2 of the present invention; -
FIG. 14 is a sectional view taken along line C-C of FIG. 13; -
FIG. 15 is an enlarged fragmentary view ofFIG. 14 ; -
FIG. 16 is a sectional view taken along line D-D ofFIG. 13 ; -
FIG. 17 is an enlarged fragmentary view ofFIG. 16 ; and -
FIG. 18 is an equivalent view ofFIG. 13 , showing a modified inner tube of the double-wall-tube heat exchanger ofEmbodiment 2. - Embodiments of the present invention will next be described in detail with reference to the drawings.
- In the following description, the term “aluminum” encompasses aluminum alloys in addition to pure aluminum.
- In the drawings, like sections or components throughout the several views are denoted by like reference numerals, and repeated description thereof is omitted.
- The present embodiment is shown in
FIGS. 1 to 7 . -
FIG. 1 shows the configuration of a double-wall-tube heat exchanger according toEmbodiment 1 of the present invention.FIGS. 2 to 6 show the configurations of essential portions of the double-wall-tube heat exchanger.FIG. 7 shows a refrigeration system which uses the double-wall-tube heat exchanger ofEmbodiment 1 as an intermediate heat exchanger. - In
FIGS. 1 to 5 , a double-wall-tube heat exchanger 1 includes anouter tube 2 and aninner tube 3. Theouter tube 2 is formed of an aluminum extrudate having a circular cross section. Theinner tube 3 is formed of an aluminum extrudate having a circular cross section and is inserted concentrically into and spaced apart from theouter tube 2. A clearance between theouter tube 2 and theinner tube 3 serves as a firstrefrigerant flow path 4. The interior of theinner tube 3 serves as a secondrefrigerant flow path 5. - The
outer tube 2 has expandedtube portions tube portion 6 of theouter tube 2. Arefrigerant outlet 8 is formed in the tube wall of the expandedtube portion 7 of theouter tube 2. An end portion of a liquid-phaserefrigerant inflow pipe 9 made of aluminum is inserted into the refrigerant inlet and is brazed to the expandedtube portion 6. An end portion of a liquid-phaserefrigerant outflow pipe 11 made of aluminum is inserted into therefrigerant outlet 8 and is brazed to the expandedtube portion 7. Preferably, theouter tube 2 has an outside diameter of 25 mm or less and a tube wall thickness of 0.2 mm to 2.0 mm inclusive. - The
inner tube 3 has a plurality ofinterior fins 12 formed integrally with the inner circumferential surface thereof. Theinterior fins 12 project radially inward, extend in the longitudinal direction of theinner tube 3, and are arranged at equal circumferential intervals. Theinner tube 3 also has a plurality ofelongated projections 13 formed integrally with the outer circumferential surface thereof. Theelongated projections 13 project radially outward, extend in the longitudinal direction, and are arranged at equal circumferential intervals. The fin height of theinterior fins 12 is greater than the projecting height of theelongated projections 13. Portions of theouter tube 2 which are located longitudinally outward of the expandedportions portions 14. The diameter-reducedportions 14 are brazed to theinner tube 3 at positions located toward the opposite ends of theinner tube 3. The diameter-reducedportions 14 are formed after theinner tube 3 is placed within theouter tube 2. In the course of formation of the diameter-reducedportions 14, associated portions of theelongated projections 13 of theinner tube 3 are crushed and bite into the inner circumferential surfaces of the diameter-reduced portions 14 (seeFIGS. 5 and 6 ). This reduces the clearance between the inner circumferential surface of theouter tube 2 and a portion of the outer circumferential surface of theinner tube 3 where theelongated projections 13 are not formed, to such an extent as to be filled with a brazing material. In this condition, the diameter-reducedportions 14 of theouter tube 2 are brazed to theinner tube 3. Also, the clearance between the inner circumferential surfaces of the diameter-reducedportions 14 of theouter tube 2 and the portion of the outer circumferential surface of theinner tube 3 where theelongated projections 13 are not formed is filled with a brazing material 17 (seeFIG. 6 ). - An end portion; i.e., an expanded
pipe portion 15 a, of a vapor-phaserefrigerant inflow pipe 15 made of aluminum is fitted onto and brazed to an end portion of theinner tube 3 which is located on a side toward therefrigerant outlet 8. Similarly, an end portion; i.e., an expandedpipe portion 16 a, of a vapor-phaserefrigerant outflow pipe 16 made of aluminum is fitted onto and brazed to an end portion of theinner tube 3 which is located on a side toward the refrigerant inlet. The portions of theinner tube 3 onto which the expandedpipe portions elongated projections 13 cut off. Instead of cutting off the associated portions of theelongated projections 13, as in the case of brazing between theouter tube 2 and theinner tube 3, the expandedpipe portion 15 a of the vapor-phaserefrigerant inflow pipe 15 and the expandedpipe portion 16 a of the vapor-phaserefrigerant outflow pipe 16 may be pressed from the radial outside so as to establish a condition in which associated portions of theelongated projections 13 are crushed and bite into the inner circumferential surfaces of the expandedpipe portions pipe portions inner tube 3 where theelongated projections 13 are not formed is reduced to such an extent as to be filled with the brazing material. Preferably, brazing between theinner tube 3 and each of the expandedpipe portions outflow pipes outer tube 2 and theinner tube 3 while an appropriate gap is maintained between the opposite ends of theouter tube 2 and the corresponding ends of the expandedpipe portions outflow pipes -
FIG. 7 shows a refrigeration system in which the above-described double-wall-tube heat exchanger 1 is used as an intermediate heat exchanger. - In
FIG. 7 , the refrigeration system uses, for example, a chlorofluorocarbon-based refrigerant. The refrigeration system includes acompressor 20; acondenser 21 having a condensingsection 22, aliquid receiver 23 serving as a vapor-liquid separator, and asubcooling section 24; anevaporator 25; anexpansion valve 26 serving as a pressure-reducing device; and the double-wall-tube heat exchanger 1 which serves as an intermediate heat exchanger for performing heat exchange between a refrigerant from thecondenser 20 and a refrigerant from theevaporator 25. Piping extending from thesubcooling section 24 of thecondenser 20 is connected to the liquid-phaserefrigerant inflow pipe 9 connected to theouter tube 2 of the double-wall-tube heat exchanger 1. Similarly, piping extending to theexpansion valve 26 is connected to the liquid-phaserefrigerant outflow pipe 11 connected to theouter tube 2. Also, piping extending from theevaporator 25 is connected to the vapor-phaserefrigerant inflow pipe 15 connected to theinner tube 3 of the double-wall-tube heat exchanger 1. Similarly, piping extending to thecompressor 20 is connected to the vapor-phaserefrigerant outflow pipe 16 connected to theinner pipe 3. The refrigeration system is mounted in a vehicle; for example, an automobile, as a car air conditioner. - In operation of the refrigeration system, a high-temperature, high-pressure vapor-liquid mixed-phase refrigerant, which has undergone compression in the
compressor 20, is cooled and condensed in the condensingsection 22 of thecondenser 21. Subsequently, the refrigerant flows into theliquid receiver 23 and is separated into two phases; namely, the vapor phase and the liquid phase. The resultant liquid-phase refrigerant flows into thesubcooling section 24 and is subcooled. The subcooled liquid-phase refrigerant flows through the liquid-phaserefrigerant inflow pipe 9 and flows into the firstrefrigerant flow path 4 of the double-wall-tube heat exchanger 1. At this time, by the effect of the expandedtube portion 6, the liquid-phase refrigerant dividedly flows into all the channels formed between the adjacentelongated projections 13 in the firstrefrigerant flow path 4. Meanwhile, a vapor-phase refrigerant from the evaporator 25 passes through the vapor-phaserefrigerant inflow pipe 15 and flows into the secondrefrigerant flow path 5 of the double-wall-tube heat exchanger 1. While flowing through the firstrefrigerant flow path 4, the liquid-phase refrigerant is further cooled by the vapor-phase refrigerant whose temperature is relatively low and which flows through the secondrefrigerant flow path 5. Having passed through all the channels formed between the adjacentelongated projections 13 in the firstrefrigerant flow path 4 of the double-wall-tube heat exchanger 1, flows of the liquid-phase refrigerant join together in the expandedtube portion 7. The resultant liquid-phase refrigerant flows to theexpansion valve 26 through the liquid-phaserefrigerant outflow pipe 11. In theexpansion valve 26, the liquid-phase refrigerant is adiabatically expanded and is thereby pressure-reduced. Subsequently, the two-phase refrigerant flows into theevaporator 25 and is evaporated in theevaporator 25. Meanwhile, having passed through the secondrefrigerant flow path 5 of the double-wall-tube heat exchanger 1, the vapor-phase refrigerant flows to thecompressor 20 through the vapor-phaserefrigerant outflow pipe 16. - In the double-wall-
tube heat exchanger 1 in which theouter tube 2 and theinner tube 3 are formed of aluminum, the fin thickness T1 of theinterior fins 12 of theinner tube 3 is preferably 0.2 mm to 2.0 mm inclusive. This has been obtained from the results of computer simulation calculation. Specifically, the relation between the fin thickness T1 and fin efficiency of theinterior fins 12 was obtained by computer simulation calculation which was conducted while the fin thickness T1 and the fin height H1 of theinterior fins 12 were varied under the following condition: the heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the secondrefrigerant flow path 5 to the inner circumferential surface of theinner tube 3 in the double-wall-tube heat exchanger 1 was set to 380 W/m2·K.FIG. 8 shows the results of the computer simulation calculation. It has been derived from the results shown inFIG. 8 that a fin thickness T1 of theinterior fins 12 of 0.2 mm to 2.0 mm inclusive is preferred. As is apparent fromFIG. 8 , when the fin thickness T1 of theinterior fins 12 is less than 0.2 mm, fin efficiency drops sharply. Also, when the fin thickness T1 is in excess of 1.2 mm, the effect of improving fin efficiency is saturated. - Preferably, the fin height H1 of the
interior fins 12 of theinner tube 3 is 1.0 mm to 3.0 mm inclusive. This also has been derived from the results shown inFIG. 8 . - Preferably, the fin pitch P1 of the
interior fins 12 of theinner tube 3 as measured at roots of theinterior fins 12 is 2 mm or greater. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the number of the interior fins 12 to pressure loss and the exchanged heat quantity between the liquid-phase refrigerant and the vapor-phase refrigerant was obtained by computer simulation calculation which was conducted while the number of the interior fins 12 (the fin pitch P1 as measured at roots of the interior fins 12) was varied under the following conditions: the inside diameter D of the inner tube 3 was set to 13.5 mm; the fin thickness T1 of the interior fins 12 was set to 0.5 mm; the fin height H1 of the interior fins 12 was set to 1.5 mm; the elongated-projection thickness T2 of the elongated projections 13 was set to 0.5 mm; the projecting height H2 of the elongated projections 13 was set to 0.5 mm; the pitch P2 of the elongated projections 13 as measured at roots of the elongated projections 13 was set to 3.0 mm; the radial clearance W (hereinafter referred to as liquid flow path width) between the inner circumferential surface of the outer tube 2 and a portion of the outer circumferential surface of the inner tube 3 where the elongated projections 13 are not formed was set to 0.8 mm; the temperature and pressure of the liquid-phase refrigerant as measured at the inlet of the first refrigerant flow path 4 were set to 42.0° C. and 1.28 MPaG, respectively; and the temperature and pressure of the vapor-phase refrigerant as measured at the inlet of the second refrigerant flow path 5 were set to 8.0° C. and 0.21 MPaG, respectively.FIG. 9 shows the results of the computer simulation calculation. It has been derived from the results shown inFIG. 9 that a fin pitch P1 of theinterior fins 12 as measured at roots of theinterior fins 12 of 2.0 mm or greater is preferred. As is apparent fromFIG. 9 , when the fin pitch P1 of theinterior fins 12 is less than 2 mm, pressure loss in the secondrefrigerant flow path 5 increases. InFIG. 9 , exchanged heat quantity and pressure loss are expressed in percentage with those of the case where an inner tube not having the interior fins is used, being each taken as 100%. - Preferably, the inside diameter D of the
inner tube 3 is 12 mm or greater. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the inside diameter D of theinner tube 3 to pressure loss and the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the secondrefrigerant flow path 5 to the inner circumferential surface of theinner tube 3 was obtained by computer simulation calculation which was conducted while the inside diameter D of theinner tube 3 was varied under the following conditions: the fin thickness T1 of theinterior fins 12 of theinner tube 3 was set to 0.5 mm; the fin height H1 of theinterior fins 12 was set to 1.5 mm; the fin pitch P1 of theinterior fins 12 as measured at roots of theinterior fins 12 was set to 2.5 mm; and the temperature and pressure of the vapor-phase refrigerant as measured at the inlet of the secondrefrigerant flow path 5 were set to 5.0° C. and 0.30 MPaG, respectively.FIG. 10 shows the results of the computer simulation calculation. It has been derived from the results shown inFIG. 10 that an inside diameter D of theinner tube 3 of 12 mm or greater is preferred. As is apparent fromFIG. 10 , when the inside diameter D of theinner tube 3 is less than 12 mm, pressure loss increases sharply. Preferably, the upper limit of the inside diameter D of theinner tube 3 is 18 mm. This is because, as is apparent fromFIG. 10 , when the inside diameter D of theinner tube 3 is in excess of 18 mm, the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the secondrefrigerant flow path 5 to the inner circumferential surface of theinner tube 3 drops. InFIG. 10 , pressure loss and the overall heat transfer coefficient in heat transfer from the vapor-phase refrigerant flowing through the secondrefrigerant flow path 5 to the inner circumferential surface of theinner tube 3 are expressed in percentage with those of the case where an inner tube having an inside diameter of 12 mm is used, being each taken as 100%. - Further, preferably, the liquid flow path width W is 0.4 mm to 1.2 mm inclusive. This has been obtained from the results of computer simulation calculation. Specifically, the relation of the liquid flow path width W to pressure loss and the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the first
refrigerant flow path 4 to the outer circumferential surface of theinner tube 3 was obtained for the cases of the inside diameter D of theinner tube 3 being 12 mm and 18 mm by computer simulation calculation which was conducted while the liquid flow path width W was varied under the following conditions: the elongated-projection thickness T2 of theelongated projections 13 of theinner tube 3 was set to 0.5 mm; the pitch P2 of theelongated projections 13 as measured at roots of theelongated projections 13 was set to 3 mm; the inside diameter D of theinner tube 3 was set to 12 mm and 18 mm; the wall thickness of theinner tube 3 was set to 1.2 mm; and the temperature and pressure of the liquid-phase refrigerant as measured at the inlet of the firstrefrigerant flow path 4 were set to 40° C. and 1.38 MPaG, respectively.FIG. 11 shows the results of the computer simulation calculation. It has been derived from the results shown inFIG. 11 that the a liquid flow path width W of 0.4 mm to 1.2 mm inclusive is preferred. As is apparent fromFIG. 11 , when the liquid flow path width W is less than 0.4 mm, pressure loss increases sharply. Also, when the liquid flow path width W is in excess of 1.2 mm, the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the firstrefrigerant flow path 4 to the outer circumferential surface of theinner tube 3 drops. InFIG. 11 , pressure loss and the overall heat transfer coefficient in heat transfer from the liquid-phase refrigerant flowing through the firstrefrigerant flow path 4 to the outer circumferential surface of theinner tube 3 are expressed in percentage with those of the case where the inside diameter D of theinner tube 3 is 12 mm, and the liquid flow path width W is 0.4 mm, being each taken as 100%. -
FIG. 12 shows a modified inner tube of the double-wall-tube heat exchanger ofEmbodiment 1. - An
inner tube 30 shown inFIG. 12 is twisted about its axis, so that theinterior fins 12 and theelongated projections 13 are spiral. - The present embodiment is shown in
FIGS. 13 to 17 . -
FIGS. 13 to 17 show the configurations of essential portions of a double-wall-tube heat exchanger according toEmbodiment 2 of the present invention. - In the case of the double-wall-
tube heat exchanger 31 ofEmbodiment 2, theouter tube 2 has a plurality ofelongated projections 32 formed integrally with the inner circumferential surface of theouter tube 2. Theelongated projections 32 project radially inward, extend in the longitudinal direction, and are arranged at equal circumferential intervals. Also, theouter tube 2 does not have expanded tube portions at opposite end portions thereof. - The
inner tube 3 has diameter-reducedtube portions 33 located slightly longitudinally inward of the opposite ends thereof. A refrigerant inlet (not shown) is formed in the wall of theouter tube 2 at a position corresponding to one diameter-reduced tube portion (not shown). Therefrigerant outlet 8 is formed in the wall of theouter tube 2 at a position corresponding to the other diameter-reducedtube portion 33. An end portion of a liquid-phase refrigerant inflow pipe (not shown) made of aluminum is inserted into the refrigerant inlet and is brazed to theouter tube 2. An end portion of the liquid-phaserefrigerant outflow pipe 11 made of aluminum is inserted into therefrigerant outlet 8 and is brazed to theouter tube 2. Elongated projections are not formed on the outer circumferential surface of theinner tube 3. - Portions of the
outer tube 2 which are located longitudinally outward of the diameter-reducedtube portions 33 of theinner tube 3 are subjected to roller working which is performed from the radial outside toward the radial inside along the entire circumference, thereby forming the diameter-reducedportions 14. The diameter-reducedportions 14 are brazed to portions of theinner tube 3 near the opposite ends thereof. The diameter-reducedportions 14 are formed after theinner tube 3 is placed within theouter tube 2. In the course of formation of the diameter-reducedportions 14, associated portions of theelongated projections 32 of theouter tube 2 are crushed and bite into the outer circumferential surfaces of the inner tube 3 (seeFIGS. 16 and 17 ). This reduces the clearance between the outer circumferential surface of theinner tube 3 and a portion of the inner circumferential surface of theouter tube 2 where theelongated projections 32 are not formed, to such an extent as to be filled with a brazing material. In this condition, the diameter-reducedportions 14 of theouter tube 2 are brazed to theinner tube 3. Also, the clearance between the outer circumferential surface of theinner tube 3 and a portion of the inner circumferential surface of the diameter-reducedportion 14 of theouter tube 2 where theelongated projections 32 are not formed is filled with the brazing material 17 (seeFIG. 17 ). - In contrast to the double-wall-
tube heat exchanger 1 ofEmbodiment 1, when the expandedpipe portions outflow pipes inner tube 3, the double-wall-tube heat exchanger 31 ofEmbodiment 2 does not require an operation of cutting off relevant portions of theelongated projections 32 from theouter tube 2 and an operation of pressing the expandedpipe portions outflow pipes tube heat exchanger 31 ofEmbodiment 2, preferably, brazing between theinner tube 3 and each of the expandedpipe portions outflow pipes outer tube 2 and theinner tube 3 while an appropriate gap is maintained between the opposite ends of theouter tube 2 and the corresponding ends of the expandedpipe portions outflow pipes - Other configurational features are similar to those of the double-wall-
tube heat exchanger 1 ofEmbodiment 1. The double-wall-tube heat exchanger 31 ofEmbodiment 2 is incorporated into the refrigeration system shown inFIG. 7 in a manner similar to that of the double-wall-tube heat exchanger 1 ofEmbodiment 1. - When the subcooled liquid-phase refrigerant flows through the liquid-phase refrigerant inflow pipe and flows into the first
refrigerant flow path 4 of the double-wall-tube heat exchanger 31, by the effect of the unillustrated one diameter-reduced tube portion of theinner tube 3, the liquid-phase refrigerant dividedly flows into all the channels formed between the adjacentelongated projections 32 in the firstrefrigerant flow path 4. Having passed through all the channels formed between the adjacentelongated projections 32 in the firstrefrigerant flow path 4 of the double-wall-tube heat exchanger 31, flows of the liquid-phase refrigerant join together in the other diameter-reducedtube portion 33. The resultant liquid-phase refrigerant flows to theexpansion valve 26 through the liquid-phaserefrigerant outflow pipe 11. - The double-wall-
tube heat exchanger 31 ofEmbodiment 2 may also use theinner tube 3 which is twisted about its axis. -
FIG. 18 shows a modified inner tube of the double-wall-tube heat exchanger ofEmbodiment 2. - Opposite end portions of an
inner tube 35 shown inFIG. 18 are extended, so that the vapor-phase refrigerant inflow pipe and the vapor-phase refrigerant outflow pipe are not used. Piping extending from theevaporator 25 is connected to an end portion of theinner tube 35 which is located on a side toward therefrigerant outlet 8. Piping extending to thecompressor 20 is connected to the other end portion of theinner tube 35. - In
Embodiment 1 described above, theelongated projections 13 are formed on the outer circumferential surface of theinner tube 3. InEmbodiment 2 described above, theelongated projections 32 are formed on the inner circumferential surface of theouter tube 2. However, an embodiment may be such that theelongated projections 13 and theelongated projections 32 are formed on the outer circumferential surface of theinner tube 3 and the inner circumferential surface of theouter tube 2, respectively. In this case, theelongated projections 13 of the inner tube and theelongated projections 32 of theouter tube 2 are arranged in a circumferentially staggered manner.
Claims (12)
Applications Claiming Priority (2)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
JP2007339446A JP2009162395A (en) | 2007-12-28 | 2007-12-28 | Double-wall-tube heat exchanger |
JP2007-339446 | 2007-12-28 |
Publications (1)
Publication Number | Publication Date |
---|---|
US20090166019A1 true US20090166019A1 (en) | 2009-07-02 |
Family
ID=40690984
Family Applications (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
US12/314,763 Abandoned US20090166019A1 (en) | 2007-12-28 | 2008-12-16 | Double-wall-tube heat exchanger |
Country Status (4)
Country | Link |
---|---|
US (1) | US20090166019A1 (en) |
JP (1) | JP2009162395A (en) |
CN (1) | CN101469920A (en) |
DE (1) | DE102008062486A1 (en) |
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US20100313589A1 (en) * | 2009-06-13 | 2010-12-16 | Brent Alden Junge | Tubular element |
US20110203677A1 (en) * | 2010-02-23 | 2011-08-25 | HS R & A Co., Ltd | Tube-socket assembly and method of manufacturing the same |
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US20110214847A1 (en) * | 2010-03-05 | 2011-09-08 | HS R & A Co., Ltd | Double pipe and heat exchanger having the same |
US20110219814A1 (en) * | 2010-03-09 | 2011-09-15 | GM Global Technology Operations LLC | Tubular heat exchanger for motor vehicle air conditioners |
US20130319645A1 (en) * | 2011-01-06 | 2013-12-05 | Tetra Laval Holdings & Finance S.A. | Optimised surface for freezing cylinder |
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KR101120560B1 (en) * | 2011-04-18 | 2012-03-09 | 박지오 | Method for manufacturing multiplex cooling pipe with spiral fin |
US20140182825A1 (en) * | 2011-05-24 | 2014-07-03 | Pierburg Gmbh | Heat transfer device |
US20130146262A1 (en) * | 2011-12-12 | 2013-06-13 | Hs R & A Co., Ltd. | Double pipe heat exchanger having multi-directional connector and air conditioner for vehicle including the same |
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US20130192804A1 (en) * | 2012-02-01 | 2013-08-01 | Sumitomo Light Metal Industries, Ltd. | Double pipe for heat exchanger |
EP2917673A4 (en) * | 2012-11-12 | 2016-08-17 | Ceramatec Inc | A fixed bed reactor heat transfer structure |
US9446486B2 (en) | 2012-11-15 | 2016-09-20 | GM Global Technology Operations LLC | Internal heat exchanger for a motor vehicle air-conditioning system |
GB2510934B (en) * | 2012-11-15 | 2016-10-05 | Gm Global Tech Operations Llc | Internal heat exchanger for a motor vehicle air-conditioning system |
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US20150315937A1 (en) * | 2012-12-14 | 2015-11-05 | Renate Kintea | Heat engine |
DE102013007590A1 (en) * | 2013-05-02 | 2014-11-06 | GM Global Technology Operations LLC (n. d. Gesetzen des Staates Delaware) | Internal heat exchanger for a motor vehicle air conditioning system |
USD746416S1 (en) * | 2013-08-23 | 2015-12-29 | Penn Aluminum International LLC | End-fitting of a concentric-tube heat exchanger |
CN104565641A (en) * | 2013-10-25 | 2015-04-29 | 上海汽车空调配件有限公司 | Pipe body connecting part for connecting coaxial pipe and aluminum pipe and connecting method thereof |
US20160131040A1 (en) * | 2014-11-10 | 2016-05-12 | Rolls-Royce Plc | Heat exchanger |
US10221768B2 (en) * | 2014-11-10 | 2019-03-05 | Rolls-Royce Plc | Heat exchanger having a coaxial or concentric tube construction |
US20180172361A1 (en) * | 2016-12-16 | 2018-06-21 | Hs Marston Aerospace Limited | Heat exchanger |
US11835301B2 (en) | 2021-04-07 | 2023-12-05 | Ecoinnovation Technologies Incorporée | Modular heat exchanger and method of assembly thereof |
Also Published As
Publication number | Publication date |
---|---|
CN101469920A (en) | 2009-07-01 |
JP2009162395A (en) | 2009-07-23 |
DE102008062486A1 (en) | 2009-07-02 |
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