US20170284714A1 - Heat exchanger and refrigeration cycle apparatus including the same - Google Patents
Heat exchanger and refrigeration cycle apparatus including the same Download PDFInfo
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- US20170284714A1 US20170284714A1 US15/512,657 US201515512657A US2017284714A1 US 20170284714 A1 US20170284714 A1 US 20170284714A1 US 201515512657 A US201515512657 A US 201515512657A US 2017284714 A1 US2017284714 A1 US 2017284714A1
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- line segments
- heat transfer
- heat exchanger
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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
- F25B30/00—Heat pumps
- F25B30/06—Heat pumps characterised by the source of low potential heat
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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
- F25B39/00—Evaporators; Condensers
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- F24J3/086—
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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
- F25D—REFRIGERATORS; COLD ROOMS; ICE-BOXES; COOLING OR FREEZING APPARATUS NOT OTHERWISE PROVIDED FOR
- F25D23/00—General constructional features
- F25D23/12—Arrangements of compartments additional to cooling compartments; Combinations of refrigerators with other equipment, e.g. stove
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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
- F28F1/02—Tubular elements of cross-section which is non-circular
- F28F1/022—Tubular elements of cross-section which is non-circular with multiple channels
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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
- F25B13/00—Compression machines, plants or systems, with reversible cycle
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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
- F25B23/00—Machines, plants or systems, with a single mode of operation not covered by groups F25B1/00 - F25B21/00, e.g. using selective radiation effect
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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
- F25B2313/00—Compression machines, plants or systems with reversible cycle not otherwise provided for
- F25B2313/002—Compression machines, plants or systems with reversible cycle not otherwise provided for geothermal
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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
- F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
- F25B2400/12—Inflammable refrigerants
- F25B2400/121—Inflammable refrigerants using R1234
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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
- F25B2500/00—Problems to be solved
- F25B2500/01—Geometry problems, e.g. for reducing size
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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
- F25B2500/00—Problems to be solved
- F25B2500/09—Improving heat transfers
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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
- F28D21/00—Heat-exchange apparatus not covered by any of the groups F28D1/00 - F28D20/00
- F28D2021/0019—Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for
- F28D2021/0068—Other heat exchangers for particular applications; Heat exchange systems not otherwise provided for for refrigerant cycles
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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
- F28F2255/00—Heat exchanger elements made of materials having special features or resulting from particular manufacturing processes
- F28F2255/16—Heat exchanger elements made of materials having special features or resulting from particular manufacturing processes extruded
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- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
- Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
- Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
- Y02E10/00—Energy generation through renewable energy sources
- Y02E10/10—Geothermal energy
Definitions
- the present invention relates to a heat exchanger including a flat tube and to a refrigeration cycle apparatus including the heat exchanger.
- a liquid film rich in high-boiling-point refrigerant is formed on the wall of a heat transfer tube, a concentration boundary layer rich in low-boiling-point refrigerant is formed between a vapor phase and the liquid film, and the concentration boundary layer acts as mass diffusion resistance R (thermal resistance) to each component vapor and reduces heat transfer in the heat transfer tube.
- Patent Literature 1 Japanese Unexamined Patent Application Publication No. 8-075384
- Patent Literature 2 Japanese Unexamined Patent Application Publication No. 2002-081881
- the present invention solves such problems and aims to reduce the mass diffusion resistance R (thermal resistance) in a concentration boundary layer in the flow path of a flat tube of a heat exchanger with a zeotropic refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf and thereby improve the heat transfer coefficient (intratubular heat transfer coefficient K) in the heat transfer tube.
- a heat exchanger includes at least one flat tube configured to allow a refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf to flow therethrough as a heat medium, wherein the flat tube includes a plurality of flow paths for the heat medium, and wherein each of the plurality of flow paths has a round rectangular shape in cross section, the round rectangular shape being defined by a plurality of portions including a pair of longitudinal line segments opposed to each other, a pair of lateral line segments opposed to each other, and a set of four rounded corners, each of the rounded corners being a segment of a circumference of a circle, the pair of longitudinal line segments being intersects with the pair of lateral line segments at the rounded corners. and the round rectangular shape being configured to satisfy 0.005 ⁇ rid ⁇ 0.8 where r is a radius of the circle, and d is a distance between the pair of longitudinal line segments opposed to each other.
- the use of a zeotropic refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf can reduce the mass diffusion resistance R (thermal resistance) in a concentration boundary layer in the flow path of a flat tube and thereby improve the heat transfer coefficient (intratubular heat transfer coefficient K) in the heat transfer tube.
- FIG. 1 is a cross-sectional view of a flat tube for use in a heat exchanger according to Embodiment 1 of the present invention.
- FIG. 2 is an enlarged cross-sectional view of the flow path of the flat tube according to Embodiment 1 of the present invention.
- FIG. 3 is an explanatory view of thermal resistance in a heat transfer tube of the heat exchanger according to Embodiment 1 of the present invention.
- FIG. 4 is a schematic view illustrating the state of a liquid film when the flow path according to Embodiment 1 of the present invention has a small r/d (where d is the distance between longitudinal line segments 2 c of the flow path, and r is the radius of a rounded corner 2 b of the flow path).
- FIG. 5 is a schematic view illustrating the state of a liquid film when the flow path according to Embodiment 1 of the present invention has a large r/d (where d is the distance between the longitudinal line segments 2 c of the flow path, and r is the radius of the rounded corner 2 b of the flow path).
- FIG. 6 is a gas-liquid phase equilibrium diagram of a zeotropic refrigerant mixture in the heat transfer tube according to Embodiment 1 of the present invention.
- FIG. 7 is a graph of the relationship between the intratubular heat transfer coefficient K and the ratio rid of the radius r of a rounded corner 2 b to the distance d between longitudinal line segments 2 c in a cross section of the flow path of a flat tube 1 when the heat exchanger according to Embodiment 1 of the present invention is used as a condenser.
- FIG. 8 is a schematic view illustrating the state of a liquid film in the flow path of a flat tube according to Embodiment 2 of the present invention.
- FIG. 9 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiments 1 and 2 of the present invention.
- FIG. 1 is a cross-sectional view of a flat tube for use in a heat exchanger according to Embodiment 1 of the present invention.
- FIG. 2 is an enlarged cross-sectional view of the flow path of the flat tube according to Embodiment 1 of the present invention.
- a flat tube 1 is a heat transfer tube of the heat exchanger and can be used in finned tube heat exchangers and other devices,
- the flat tube 1 is manufactured by extrusion of aluminum or another material.
- the flat tube 1 includes a plurality of flow paths 2 through which a heat medium, such as refrigerant, flows.
- the flow paths 2 have a generally rectangular cross-section as illustrated in FIG. 2 and are aligned unidirectionally such that refrigerant flows in parallel through the flat tube 1 .
- the cross section of the flow path 2 is defined by a pair of parallel lateral line segments 2 a, a pair of parallel longitudinal line segments 2 c , and a set of four rounded corners 2 b, each of the rounded corners 2 b being a segment of a circumference of a circle having a radius r.
- the flow path 2 has a rectangular cross section with four rounded corners (a rounded rectangle).
- Each of the rounded corners 2 b is a quarter circle, and both ends thereof are smoothly connected to the lateral line segments 2 a and the longitudinal line segments 2 c.
- the refrigerant is a zeotropic refrigerant mixture, for example, a three-refrigerant mixture of HFO1123, R32, and HFO1234yf.
- FIG. 3 is an explanatory view of thermal resistance in a heat transfer tube of the heat exchanger according to Embodiment 1 of the present invention.
- the heat transfer capability of a heat transfer tube is represented by the intratubular heat transfer coefficient K [W/m 2 ⁇ K].
- the intratubular heat transfer coefficient K is the heat transfer coefficient between refrigerant flowing through the heat transfer tube and the inner surface of the heat transfer tube and is an indicator of the heat exchange performance of the heat exchanger.
- a higher intratubular heat transfer coefficient K indicates lower thermal resistance of the heat transfer tube and higher heat exchange performance of the heat exchanger.
- the intratubular heat transfer coefficient K depends on the state of refrigerant in the flow path of the heat transfer tube.
- a vapor phase 4 (vapor bulk) flows through the center of the heat transfer tube, and heat transfer proceeds while a liquid film 5 is formed on a tube wall 3 .
- the intratubular heat transfer coefficient K is a function of the mass diffusion resistance R (thermal resistance) between the vapor phase 4 and the gas-liquid interface 6 and the heat transfer coefficient ⁇ L of the liquid film 5 .
- the mass diffusion resistance R (thermal resistance) between the vapor phase 4 and the gas-liquid interface 6 is a function of the mass transfer coefficient ⁇ v between the vapor phase 4 and the gas-liquid interface 6 .
- the mass diffusion resistance R decreases with increasing mass transfer coefficient ⁇ v.
- a temperature difference between the vapor phase 4 of refrigerant and the gas-liquid interface 6 results from the mass diffusion resistance R (thermal resistance) between the vapor phase 4 and the gas-liquid interface 6 .
- An increased mass diffusion resistance R results in an increased thermal resistance and an increased temperature difference between the vapor phase 4 of refrigerant and the gas-liquid interface 6 (an increased difference in the concentration of the zeotropic refrigerant mixture).
- the temperature Tvb of the state point of vapor bulk which is the temperature of the vapor phase 4 in the flow path 2
- the temperature Ti of the gas-liquid interface 6 depend on the mass transfer coefficient ⁇ v, the disturbance ⁇ v of the interphase gas-liquid interface, and the amount m of condensed refrigerant.
- the temperature Ti of the gas-liquid interface and the wall temperature Tw of the flow path 2 depend on the heat conduction of the liquid phase and provide the following relationship: the heat transfer coefficient ⁇ L of the liquid film ⁇ (thermal conductivity)/ ⁇ (the thickness of the liquid film).
- FIG. 4 is a schematic view illustrating the state of a liquid film when the flow path according to Embodiment 1 of the present invention has a small rid (where d is the distance between longitudinal line segments 2 c of the flow path, and r is the radius of the rounded corners 2 b of the flow path).
- FIG. 5 is a schematic view illustrating the state of a liquid film when the flow path according to Embodiment 1 of the present invention has a large r/d (where d is the distance between the longitudinal line segments 2 c of the flow path, and r is the radius of the rounded corners 2 b of the flow path).
- the thickness ⁇ of the liquid film is substantially uniform throughout the line segments than the state illustrated in FIG. 4 , thus resulting in greater disturbance ⁇ v of the gas-liquid interface.
- FIG. 6 is a gas-liquid phase equilibrium diagram of a zeotropic refrigerant mixture in the heat transfer tube according to Embodiment 1 of the present invention.
- the zeotropic refrigerant mixture according to Embodiment 1 is a refrigerant mixture of low-boiling-point refrigerants HFO1123+R32 and a high-boiling-point refrigerant HFO1234y1.
- the state point of vapor bulk of the zeotropic refrigerant mixture is indicated by a filled circle and moves along the saturated vapor line.
- the temperature of the state point of vapor bulk is Tvb
- the gas-liquid interface reaches the gas-liquid equilibrium state and has a temperature Ti
- the liquid phase and the wall temperature Tw are in the subcooling region.
- condensation in the heat exchanger begins when the state point of vapor bulk reaches the state of the point A from superheated vapor.
- the vapor bulk at the point A condenses into saturated liquid in the state of the point A′.
- the temperature Tvb of the state point of vapor bulk decreases to the condensing temperature of the point B, and the resulting saturated liquid has the refrigerant composition yb of the point B′.
- the component ratio of the low-boiling-point refrigerants HFO1123 and R32 ranges from approximately 50:50 to approximately 40:60, and the low-boiling-point refrigerants form a near-azeotropic refrigerant.
- the refrigerant mixture of HFO1123 and R32 is construed to have the characteristics of an almost single refrigerant.
- FIG. 7 is a graph of relationship between the intratubular heat transfer coefficient K and the ratio rid of the radius r of the rounded corners 2 b to the distance d between the longitudinal line segments 2 c in a cross section of the flow path of the flat tube 1 when the heat exchanger according to Embodiment 1 is used as a condenser.
- the intratubular heat transfer coefficient K [W/m 2 ⁇ K] is in its maximum range within 10% or less of the maximum value, and these are optimum values in terms of efficiency.
- the intratubular heat transfer coefficient K [W/m 2 ⁇ K] increases with decreasing r/d.
- the thickness of the liquid film 5 increases at the four corners of the flow path 2 and decreases along the longitudinal line segments 2 c and the lateral line segments 2 a. This decreases the thickness ⁇ of the liquid film between the gas-liquid interface 6 and the heat transfer wall of the heat transfer tube, increases the heat transfer coefficient ⁇ L of the liquid film 5 , and improves the intratubular heat transfer coefficient K.
- the liquid film at the rounded corners 2 b has disturbance ⁇ v at the gas-liquid interface, a small amount m of condensed refrigerant results in a low mass transfer coefficient ⁇ v, a high mass diffusion resistance R, and a low intratubular heat transfer coefficient K. Furthermore, the large thickness ⁇ of the liquid film results in a low heat transfer coefficient ⁇ L of the liquid film 5 and a low intratubular heat transfer coefficient K.
- the liquid film 5 is not so thick at the four corners, and the thickness ⁇ of the liquid film along the longitudinal line segments 2 c and the lateral line segments 2 a is greater than that in a region of 0.005>r/d. This results in a decreased heat transfer coefficient ⁇ L of the liquid film 5 and a decreased intratubular heat transfer coefficient K.
- the disturbance ⁇ v of the gas-liquid interface increases with the thickness ⁇ of the liquid film, and the mass diffusion resistance R decreases.
- the total intratubular heat transfer coefficient K is higher than that in the region of 0.005>r/d.
- the liquid film 5 is negligibly thickened at the four corners, and the thickness ⁇ of the liquid film along the longitudinal line segments 2 c and the lateral line segments 2 a is greater than that in the region of 0.005 ⁇ r/d ⁇ 0.8. This results in a low heat transfer coefficient al of the liquid film 5 and a lower total intratubular heat transfer coefficient K than that in the region of 0.005 ⁇ r/d ⁇ 0.8.
- the use of the zeotropic refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf can decrease the mass diffusion resistance R (thermal resistance) in a concentration boundary layer in the flow path of the flat tube and optimize the intratubular heat transfer coefficient K [W/m 2 ⁇ K] in terms of efficiency.
- the heat exchanger according to Embodiment 1 functions as a condenser in the above description, the same effects on the intratubular heat transfer coefficient K [W/m 2 ⁇ K] can be achieved when the heat exchanger according to Embodiment 1 functions as an evaporator.
- the flow path 2 of the flat tube 1 has a generally rectangular cross-section.
- Embodiment 2 is different from Embodiment 1 in that in addition to the shape of the flow path 2 according to Embodiment 1, protrusions 2 d having a projected shape are provided to the longitudinal line segments 2 c in a cross section of the flow path 2 .
- FIG. 8 is a schematic view illustrating the state of a liquid film 5 in the flow path of a flat tube according to Embodiment 2 of the present invention.
- the liquid film 5 is almost uniformly formed on the inner circumference of the flow path 2 . This is because liquid disperses around both the rounded corners 2 b and the protrusions 2 d, and the liquid film 5 is rarely thickened in a particular portion.
- FIG. 9 is a refrigerant circuit diagram of a refrigeration cycle apparatus according to Embodiments 1 and 2 of the present invention.
- the refrigerant circuit diagram of FIG. 9 includes a compressor 10 , a condenser heat exchanger 11 , an expansion valve 12 , an evaporator heat exchanger 13 , and fans 14 and 15 .
- the heat exchanger including the flat tube 1 according to Embodiments 1 or 2 of the present invention is applicable to the condenser heat exchanger 11 , the evaporator heat exchanger 13 , or both.
- the employment thereof provides a refrigeration cycle apparatus having high energy efficiency.
- the energy efficiency includes the following formulae.
- Heating energy efficiency indoor heat exchanger (condenser) capacity/total energy input
- Cooling energy efficiency indoor heat exchanger (evaporator) capacity/total energy input
- any refrigerating machine oil such as mineral oil, alkylbenzene oil, ester oil, ether oil, or fluorine oil, can offer the advantages irrespective of the solubility of the oil in refrigerant.
Abstract
A heat exchanger includes at least one flat tube configured to allow a refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf to flow therethrough as a heat medium. The flat tube includes a plurality of flow paths for the heat medium. Each of the plurality of flow paths has a round rectangular shape in cross section, the round rectangular shape being defined by a pair of longitudinal line segments opposed to each other, a pair of lateral line segments opposed to each other, and a set of four rounded corners, each of the rounded corners being a segment of a circumference of a circle. The pair of longitudinal line segments are intersects with the pair of lateral line segments at the rounded corners. The round rectangular shape is configured to satisfy 0.005≦r/d≦0.8 where r is a radius of the circle, and d is a distance between the pair of longitudinal line segments opposed to each other.
Description
- The present invention relates to a heat exchanger including a flat tube and to a refrigeration cycle apparatus including the heat exchanger.
- In a heat exchanger of a refrigeration cycle apparatus including a zeotropic refrigerant mixture, it is known that a liquid film rich in high-boiling-point refrigerant is formed on the wall of a heat transfer tube, a concentration boundary layer rich in low-boiling-point refrigerant is formed between a vapor phase and the liquid film, and the concentration boundary layer acts as mass diffusion resistance R (thermal resistance) to each component vapor and reduces heat transfer in the heat transfer tube.
- It is known that to agitate such a concentration boundary layer and reduce the mass diffusion resistance R, a plurality of projections are formed on the inner surface of a heat transfer tube to improve the mass transfer coefficientβv on the vapor side, reduce the mass diffusion resistance R (thermal resistance), and improve the heat transfer coefficient within the heat transfer tube (intratubular heat transfer coefficient K) (see
Patent Literature 1 andPatent Literature 2, for example). - Patent Literature 1: Japanese Unexamined Patent Application Publication No. 8-075384
- Patent Literature 2: Japanese Unexamined Patent Application Publication No. 2002-081881
- In a heat exchanger including projections in a heat transfer tube as described in
Patent Literature 1 andPatent Literature 2, although the mass diffusion resistance R (thermal resistance) of a zeotropic refrigerant mixture can be reduced to improve the heat transfer coefficient within the heat transfer tube (intratubular heat transfer coefficient K), the projections increase the pressure drop of the refrigerant flowing therethrough. Furthermore, although projections can be formed by various forming methods, it is difficult to form projections, for example, when a flat tube is extruded. - The present invention solves such problems and aims to reduce the mass diffusion resistance R (thermal resistance) in a concentration boundary layer in the flow path of a flat tube of a heat exchanger with a zeotropic refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf and thereby improve the heat transfer coefficient (intratubular heat transfer coefficient K) in the heat transfer tube.
- A heat exchanger according to an embodiment of the present invention includes at least one flat tube configured to allow a refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf to flow therethrough as a heat medium, wherein the flat tube includes a plurality of flow paths for the heat medium, and wherein each of the plurality of flow paths has a round rectangular shape in cross section, the round rectangular shape being defined by a plurality of portions including a pair of longitudinal line segments opposed to each other, a pair of lateral line segments opposed to each other, and a set of four rounded corners, each of the rounded corners being a segment of a circumference of a circle, the pair of longitudinal line segments being intersects with the pair of lateral line segments at the rounded corners. and the round rectangular shape being configured to satisfy 0.005≦rid≦0.8 where r is a radius of the circle, and d is a distance between the pair of longitudinal line segments opposed to each other.
- In a heat exchanger according to an embodiment of the present invention and a refrigeration cycle apparatus including the heat exchanger, the use of a zeotropic refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf can reduce the mass diffusion resistance R (thermal resistance) in a concentration boundary layer in the flow path of a flat tube and thereby improve the heat transfer coefficient (intratubular heat transfer coefficient K) in the heat transfer tube.
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FIG. 1 is a cross-sectional view of a flat tube for use in a heat exchanger according toEmbodiment 1 of the present invention. -
FIG. 2 is an enlarged cross-sectional view of the flow path of the flat tube according toEmbodiment 1 of the present invention. -
FIG. 3 is an explanatory view of thermal resistance in a heat transfer tube of the heat exchanger according toEmbodiment 1 of the present invention. -
FIG. 4 is a schematic view illustrating the state of a liquid film when the flow path according toEmbodiment 1 of the present invention has a small r/d (where d is the distance betweenlongitudinal line segments 2 c of the flow path, and r is the radius of arounded corner 2 b of the flow path). -
FIG. 5 is a schematic view illustrating the state of a liquid film when the flow path according toEmbodiment 1 of the present invention has a large r/d (where d is the distance between thelongitudinal line segments 2 c of the flow path, and r is the radius of therounded corner 2 b of the flow path). -
FIG. 6 is a gas-liquid phase equilibrium diagram of a zeotropic refrigerant mixture in the heat transfer tube according toEmbodiment 1 of the present invention. -
FIG. 7 is a graph of the relationship between the intratubular heat transfer coefficient K and the ratio rid of the radius r of arounded corner 2 b to the distance d betweenlongitudinal line segments 2 c in a cross section of the flow path of aflat tube 1 when the heat exchanger according toEmbodiment 1 of the present invention is used as a condenser. -
FIG. 8 is a schematic view illustrating the state of a liquid film in the flow path of a flat tube according toEmbodiment 2 of the present invention, -
FIG. 9 is a refrigerant circuit diagram of a refrigeration cycle apparatus according toEmbodiments - Embodiments of the present invention will be described below with reference to the accompanying drawings. However, the present invention is not limited to these embodiments.
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FIG. 1 is a cross-sectional view of a flat tube for use in a heat exchanger according toEmbodiment 1 of the present invention. -
FIG. 2 is an enlarged cross-sectional view of the flow path of the flat tube according toEmbodiment 1 of the present invention. - In
FIG. 1 , aflat tube 1 is a heat transfer tube of the heat exchanger and can be used in finned tube heat exchangers and other devices, Theflat tube 1 is manufactured by extrusion of aluminum or another material. Theflat tube 1 includes a plurality offlow paths 2 through which a heat medium, such as refrigerant, flows. Theflow paths 2 have a generally rectangular cross-section as illustrated inFIG. 2 and are aligned unidirectionally such that refrigerant flows in parallel through theflat tube 1. - As illustrated in
FIG. 2 , the cross section of theflow path 2 is defined by a pair of parallellateral line segments 2 a, a pair of parallellongitudinal line segments 2 c, and a set of fourrounded corners 2 b, each of therounded corners 2 b being a segment of a circumference of a circle having a radius r. Thus, theflow path 2 has a rectangular cross section with four rounded corners (a rounded rectangle). - Each of the
rounded corners 2 b is a quarter circle, and both ends thereof are smoothly connected to thelateral line segments 2 a and thelongitudinal line segments 2 c. - In the heat exchanger according to
Embodiment 1, the refrigerant is a zeotropic refrigerant mixture, for example, a three-refrigerant mixture of HFO1123, R32, and HFO1234yf. -
FIG. 3 is an explanatory view of thermal resistance in a heat transfer tube of the heat exchanger according toEmbodiment 1 of the present invention. - The heat transfer capability of a heat transfer tube is represented by the intratubular heat transfer coefficient K [W/m2·K]. The intratubular heat transfer coefficient K is the heat transfer coefficient between refrigerant flowing through the heat transfer tube and the inner surface of the heat transfer tube and is an indicator of the heat exchange performance of the heat exchanger. A higher intratubular heat transfer coefficient K indicates lower thermal resistance of the heat transfer tube and higher heat exchange performance of the heat exchanger.
- The intratubular heat transfer coefficient K depends on the state of refrigerant in the flow path of the heat transfer tube.
- When gas refrigerant flows through the heat transfer tube and condenses, a vapor phase 4 (vapor bulk) flows through the center of the heat transfer tube, and heat transfer proceeds while a liquid film 5 is formed on a
tube wall 3. - The intratubular heat transfer coefficient K is a function of the mass diffusion resistance R (thermal resistance) between the
vapor phase 4 and the gas-liquid interface 6 and the heat transfer coefficient αL of the liquid film 5. - First, the mass diffusion resistance R (thermal resistance) between the
vapor phase 4 and the gas-liquid interface 6 will be described below. - The mass diffusion resistance R (thermal resistance) between the
vapor phase 4 and the gas-liquid interface 6 is a function of the mass transfer coefficient βv between thevapor phase 4 and the gas-liquid interface 6. The mass diffusion resistance R decreases with increasing mass transfer coefficient βv. - A temperature difference between the
vapor phase 4 of refrigerant and the gas-liquid interface 6 results from the mass diffusion resistance R (thermal resistance) between thevapor phase 4 and the gas-liquid interface 6. An increased mass diffusion resistance R results in an increased thermal resistance and an increased temperature difference between thevapor phase 4 of refrigerant and the gas-liquid interface 6 (an increased difference in the concentration of the zeotropic refrigerant mixture). - The temperature Tvb of the state point of vapor bulk, which is the temperature of the
vapor phase 4 in theflow path 2, and the temperature Ti of the gas-liquid interface 6 depend on the mass transfer coefficient βv, the disturbance Φv of the interphase gas-liquid interface, and the amount m of condensed refrigerant. - There is the relationship βv (mass transfer coefficient)˜Φv (disturbance of the interphase gas-liquid interface)×m (the amount of condensed refrigerant), and βv (mass transfer coefficient) increases with Φv (disturbance of the interphase gas-liquid interface)×m (the amount of condensed refrigerant).
- Thus, a higher mass transfer coefficient βv results in a lower mass diffusion resistance R and a higher intratubular heat transfer coefficient K. This results in a smaller difference between the temperature Tvb of the state point of vapor bulk and the temperature Ti of the gas-
liquid interface 6. - Next, the heat transfer coefficient αL of the liquid film will be described below.
- The temperature Ti of the gas-liquid interface and the wall temperature Tw of the
flow path 2 depend on the heat conduction of the liquid phase and provide the following relationship: the heat transfer coefficient αL of the liquid film˜λ (thermal conductivity)/δ(the thickness of the liquid film). - Thus, a decreased thickness of the liquid film δ and an increased heat transfer coefficient αL result in an increased intratubular heat transfer coefficient K. Thus, to improve the intratubular heat transfer coefficient K, it is necessary to determine the maximum in the function of the mass diffusion resistance R between the
vapor phase 4 and the gas-liquid interface 6 and the heat transfer coefficient αL of the liquid film. -
FIG. 4 is a schematic view illustrating the state of a liquid film when the flow path according toEmbodiment 1 of the present invention has a small rid (where d is the distance betweenlongitudinal line segments 2 c of the flow path, and r is the radius of therounded corners 2 b of the flow path). -
FIG. 5 is a schematic view illustrating the state of a liquid film when the flow path according toEmbodiment 1 of the present invention has a large r/d (where d is the distance between thelongitudinal line segments 2 c of the flow path, and r is the radius of therounded corners 2 b of the flow path). - When gas refrigerant flows through the heat transfer tube and condenses, as illustrated in
FIGS. 4 and 5 , the gas refrigerant (vapor bulk) flows through the center of the heat transfer tube, and heat transfer proceeds while a liquid film 5 is formed on atube wall 3. InFIG. 4 where r/d is small (d is the distance between thelongitudinal line segments 2 c of theflow path 2, and r is the radius of therounded corners 2 b of the flow path 2), the thickness δ of a liquid film is large at the four corners of theflow path 2 and is small along the line segments. The disturbance Φv of the gas-liquid interface is small along the line segments due to the small thickness δ of the liquid film and is large at the four corners. - As illustrated in
FIG. 5 where r/d is greater than that in theflow path 2 illustrated inFIG. 4 , the thickness δ of the liquid film is substantially uniform throughout the line segments than the state illustrated inFIG. 4 , thus resulting in greater disturbance Φv of the gas-liquid interface. -
FIG. 6 is a gas-liquid phase equilibrium diagram of a zeotropic refrigerant mixture in the heat transfer tube according toEmbodiment 1 of the present invention, - The zeotropic refrigerant mixture according to
Embodiment 1 is a refrigerant mixture of low-boiling-point refrigerants HFO1123+R32 and a high-boiling-point refrigerant HFO1234y1. - In
FIG. 6 , the state point of vapor bulk of the zeotropic refrigerant mixture is indicated by a filled circle and moves along the saturated vapor line. The temperature of the state point of vapor bulk is Tvb, the gas-liquid interface reaches the gas-liquid equilibrium state and has a temperature Ti, and the liquid phase and the wall temperature Tw are in the subcooling region. - For a zeotropic refrigerant mixture having a refrigerant composition yb, condensation in the heat exchanger begins when the state point of vapor bulk reaches the state of the point A from superheated vapor. The vapor bulk at the point A condenses into saturated liquid in the state of the point A′.
- As condensation proceeds, a refrigerant mixture containing the high-boiling-point refrigerant HFO1234yf at a high component ratio condenses earlier, and the temperature Tvb of the state point of vapor bulk decreases gradually. The temperature Ti of the gas-liquid interface also decreases simultaneously,
- Finally, the temperature Tvb of the state point of vapor bulk decreases to the condensing temperature of the point B, and the resulting saturated liquid has the refrigerant composition yb of the point B′.
- When the charge composition of the high-boiling-point refrigerant HFO1234yf ranges from 50% to 90% of the zeotropic refrigerant mixture, the mass diffusion resistance R between the vapor bulk phase and the gas-
liquid interface 6 increases. This results in a large temperature difference between the temperature Tvb of the state point of vapor bulk and the temperature Ti of the gas-liquid interface (difference in the concentration of the zeotropic refrigerant mixture). In this state, to improve the intratubular heat transfer coefficient K, it is effective to cause the disturbance Φv of the gas-liquid interface as illustrated inFIG. 5 to increase the mass transfer coefficient βv and decrease the mass diffusion resistance R. - By contrast, when the charge composition of HFO1234yf is 50% or less of the zeotropic refrigerant mixture, the mass diffusion resistance R between the
vapor phase 4 and the gas-liquid interface 6 decreases, and the characteristics approaches those of a single refrigerant. This results in a small temperature difference between the temperature Tvb of the state point of vapor bulk and the temperature Ti of the gas-liquid interface (difference in the concentration of the zeotropic refrigerant mixture). Thus, the disturbance Φv of the gas-liquid interface contributes less to the improvement of the intratubular heat transfer coefficient K. - The component ratio of the low-boiling-point refrigerants HFO1123 and R32 ranges from approximately 50:50 to approximately 40:60, and the low-boiling-point refrigerants form a near-azeotropic refrigerant. Thus, the refrigerant mixture of HFO1123 and R32 is construed to have the characteristics of an almost single refrigerant.
- From the above, when the heat exchanger according to
Embodiment 1 is used as a condenser, the relationship between the intratubular heat transfer coefficient K and the ratio rid of the radius r of therounded corners 2 b to the distance d between thelongitudinal line segments 2 c in a cross section of theflow path 2 of theflat tube 1 will be described below with reference toFIG. 7 , -
FIG. 7 is a graph of relationship between the intratubular heat transfer coefficient K and the ratio rid of the radius r of therounded corners 2 b to the distance d between thelongitudinal line segments 2 c in a cross section of the flow path of theflat tube 1 when the heat exchanger according toEmbodiment 1 is used as a condenser. - When the mass velocity Gr of refrigerant in the
flow path 2 is 200 [kg/m2·s], which is the condition employed in general heat exchangers, and the ratio r/d of the radius r of therounded corners 2 b to the distance d between thelongitudinal line segments 2 c in a cross section of theflow path 2 is in the range of 0.005≦r/d≦0.8,the intratubular heat transfer coefficient K [W/m2·K] is in its maximum range within 10% or less of the maximum value, and these are optimum values in terms of efficiency. For a near-azeotropic refrigerant mixture R410A, the intratubular heat transfer coefficient K [W/m2·K] increases with decreasing r/d. - This will be described below.
- In
FIG. 7 , when rid decreases, the thickness of the liquid film 5 increases at the four corners of theflow path 2 and decreases along thelongitudinal line segments 2 c and thelateral line segments 2 a. This decreases the thickness δ of the liquid film between the gas-liquid interface 6 and the heat transfer wall of the heat transfer tube, increases the heat transfer coefficient αL of the liquid film 5, and improves the intratubular heat transfer coefficient K. - In a region with 0.005>r/d, however, the disturbance Φv of the gas-liquid interface in effective heat transfer portions of the
longitudinal line segments 2 c and thelateral line segments 2 a decreases, the mass diffusion resistance R on the vapor side increases, and the intratubular heat transfer coefficient K decreases significantly. With these contradictory effects, the total intratubular heat transfer coefficient K decreases. - Although the liquid film at the
rounded corners 2 b has disturbance Φv at the gas-liquid interface, a small amount m of condensed refrigerant results in a low mass transfer coefficient βv, a high mass diffusion resistance R, and a low intratubular heat transfer coefficient K. Furthermore, the large thickness δ of the liquid film results in a low heat transfer coefficient αL of the liquid film 5 and a low intratubular heat transfer coefficient K. - In a region of 0.0056≦r/d≦0.8, the liquid film 5 is not so thick at the four corners, and the thickness δ of the liquid film along the
longitudinal line segments 2 c and thelateral line segments 2 a is greater than that in a region of 0.005>r/d. This results in a decreased heat transfer coefficient αL of the liquid film 5 and a decreased intratubular heat transfer coefficient K. - However, the disturbance Φv of the gas-liquid interface increases with the thickness δ of the liquid film, and the mass diffusion resistance R decreases. Thus, the total intratubular heat transfer coefficient K is higher than that in the region of 0.005>r/d.
- In a region of 0.8<r/d, the liquid film 5 is negligibly thickened at the four corners, and the thickness δ of the liquid film along the
longitudinal line segments 2 c and thelateral line segments 2 a is greater than that in the region of 0.005≦r/d≦0.8. This results in a low heat transfer coefficient al of the liquid film 5 and a lower total intratubular heat transfer coefficient K than that in the region of 0.005≦r/d≦0.8. - Thus, when the ratio r/d of the radius r of the
rounded corners 2 b to the length d of thelateral line segments 2 a in a cross section of theflow path 2 is in the range of 0.005≦r/d≦0.8, the use of the zeotropic refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf can decrease the mass diffusion resistance R (thermal resistance) in a concentration boundary layer in the flow path of the flat tube and optimize the intratubular heat transfer coefficient K [W/m2·K] in terms of efficiency. - Although the heat exchanger according to
Embodiment 1 functions as a condenser in the above description, the same effects on the intratubular heat transfer coefficient K [W/m2·K] can be achieved when the heat exchanger according toEmbodiment 1 functions as an evaporator. - In
Embodiment 1, theflow path 2 of theflat tube 1 has a generally rectangular cross-section.Embodiment 2 is different fromEmbodiment 1 in that in addition to the shape of theflow path 2 according toEmbodiment 1,protrusions 2 d having a projected shape are provided to thelongitudinal line segments 2 c in a cross section of theflow path 2. -
FIG. 8 is a schematic view illustrating the state of a liquid film 5 in the flow path of a flat tube according toEmbodiment 2 of the present invention. - The liquid film 5 is almost uniformly formed on the inner circumference of the
flow path 2. This is because liquid disperses around both therounded corners 2 b and theprotrusions 2 d, and the liquid film 5 is rarely thickened in a particular portion. - This ensures a uniform thickness δ of the liquid film in the effective heat transfer portion and increases the disturbance Φv of the gas-liquid interface. This decreases the mass diffusion resistance R on the vapor side and improves the intratubular heat transfer coefficient K. This can also increase the intratubular heat transfer area and improve the heat exchanger efficiency.
-
FIG. 9 is a refrigerant circuit diagram of a refrigeration cycle apparatus according toEmbodiments - The refrigerant circuit diagram of
FIG. 9 includes acompressor 10, acondenser heat exchanger 11, anexpansion valve 12, anevaporator heat exchanger 13, andfans - The heat exchanger including the
flat tube 1 according toEmbodiments condenser heat exchanger 11, theevaporator heat exchanger 13, or both. The employment thereof provides a refrigeration cycle apparatus having high energy efficiency. - The energy efficiency includes the following formulae.
-
Heating energy efficiency=indoor heat exchanger (condenser) capacity/total energy input -
Cooling energy efficiency=indoor heat exchanger (evaporator) capacity/total energy input - In the heat exchanger described in
Embodiment 1 and a refrigeration cycle apparatus including the heat exchanger, any refrigerating machine oil, such as mineral oil, alkylbenzene oil, ester oil, ether oil, or fluorine oil, can offer the advantages irrespective of the solubility of the oil in refrigerant. - 1 flat tube, 2 flow path, 2 a lateral line segment, 2 b rounded corner, 2 c longitudinal line segment, 2 d protrusion, 3 tube wall, 4 vapor phase, 5 liquid film, 6 gas-liquid interface, 10 compressor, 11 condenser heat exchanger, 12 expansion valve, 13 evaporator heat exchanger, 14 fan, 15 fan, K intratubular heat transfer coefficient, R mass diffusion resistance, Ti gas-liquid interface temperature, Tvb temperature of state point of vapor bulk, Tw wall temperature, m amount of condensation, r radius of rounded corner, d distance between pair of longitudinal line segments, yb refrigerant composition, αL heat transfer coefficient of liquid film, βv mass transfer coefficient on vapor side, λ heat transfer coefficient of liquid film, δ thickness of liquid film, Φv disturbance of gas-liquid interface.
Claims (5)
1. A heat exchanger comprising:
a flat tube configured to allow a refrigerant mixture inclusive of HFO1123, R32, and HFO1234yf to flow therethrough as a heat medium,
the flat tube having a plurality of flow paths for the heat medium, and
each of the plurality of flow paths having a rectangular shape in cross section, the rectangular shape being defined by a pair of longitudinal line segments opposed to each other, a pair of lateral line segments opposed to each other, and a set of four rounded corners, each of the rounded corners being a segment of a circumference of a circle, the pair of longitudinal line segments being intersects with the pair of lateral line segments at the rounded corners,
the rectangular shape being configured to satisfy
0.005≦r/d≦0.8
0.005≦r/d≦0.8
where r is a radius of the circle, and d is a distance between the pair of longitudinal line segments opposed to each other.
2. The heat exchanger of claim 1 , wherein a mixture ratio of HFO1234yf in the refrigerant mixture ranges from 50% to 90% by weight.
3. The heat exchanger of claim 1 , wherein the longitudinal line segments have a protrusion extending in a longitudinal direction of the flat tube.
4. The heat exchanger of claim 1 , wherein the flat tube is formed by extrusion.
5. A refrigeration cycle apparatus comprising the heat exchanger of claim 1 .
Applications Claiming Priority (1)
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PCT/JP2015/050464 WO2016110997A1 (en) | 2015-01-09 | 2015-01-09 | Heat exchanger and refrigeration cycle apparatus having said heat exchanger |
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US20170284714A1 true US20170284714A1 (en) | 2017-10-05 |
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US15/512,657 Abandoned US20170284714A1 (en) | 2015-01-09 | 2015-01-09 | Heat exchanger and refrigeration cycle apparatus including the same |
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US (1) | US20170284714A1 (en) |
EP (1) | EP3244156B1 (en) |
JP (1) | JP6395863B2 (en) |
CN (1) | CN107003081A (en) |
WO (1) | WO2016110997A1 (en) |
Cited By (1)
Publication number | Priority date | Publication date | Assignee | Title |
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US20170003060A1 (en) * | 2014-03-17 | 2017-01-05 | Mitsubishi Electric Corporation | Air-conditioning apparatus |
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JP6878918B2 (en) * | 2017-01-30 | 2021-06-02 | 株式会社富士通ゼネラル | Refrigeration cycle equipment |
Citations (3)
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US5476141A (en) * | 1993-04-19 | 1995-12-19 | Sanden Corporation | Flat-type refrigerant tube having an improved pressure-resistant strength |
US8534349B2 (en) * | 2000-11-24 | 2013-09-17 | Keihin Thermal Technology Corporation | Heat exchanger tube and heat exchanger |
US20150376486A1 (en) * | 2013-04-30 | 2015-12-31 | Asahi Glass Company, Limited | Working fluid for heat cycle |
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JPH0875384A (en) * | 1994-07-01 | 1996-03-19 | Hitachi Ltd | Heat transfer tube for non-azeotrope refrigerant, heat exchanger using the same tube, assembling method and refrigerating air conditioner using the same exchanger |
JP3771433B2 (en) * | 2000-09-01 | 2006-04-26 | 住友軽金属工業株式会社 | Method for condensing non-azeotropic refrigerant mixture |
KR100906769B1 (en) * | 2002-01-31 | 2009-07-10 | 한라공조주식회사 | Heat exchanger tube with tumbling toy-shaped passages and heat exchanger using the same |
WO2009154149A1 (en) * | 2008-06-16 | 2009-12-23 | 三菱電機株式会社 | Non‑azeotropic refrigerant mixture and refrigeration cycle device |
JP5257102B2 (en) * | 2009-01-30 | 2013-08-07 | 三菱電機株式会社 | Heat exchanger and refrigeration air conditioner |
JP5595343B2 (en) * | 2011-06-29 | 2014-09-24 | 三菱電機株式会社 | Heat exchanger, refrigeration cycle circuit using the same, refrigerator using the refrigeration cycle circuit, and air conditioner |
-
2015
- 2015-01-09 JP JP2016568239A patent/JP6395863B2/en active Active
- 2015-01-09 EP EP15876873.9A patent/EP3244156B1/en active Active
- 2015-01-09 CN CN201580065041.1A patent/CN107003081A/en active Pending
- 2015-01-09 WO PCT/JP2015/050464 patent/WO2016110997A1/en active Application Filing
- 2015-01-09 US US15/512,657 patent/US20170284714A1/en not_active Abandoned
Patent Citations (3)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US5476141A (en) * | 1993-04-19 | 1995-12-19 | Sanden Corporation | Flat-type refrigerant tube having an improved pressure-resistant strength |
US8534349B2 (en) * | 2000-11-24 | 2013-09-17 | Keihin Thermal Technology Corporation | Heat exchanger tube and heat exchanger |
US20150376486A1 (en) * | 2013-04-30 | 2015-12-31 | Asahi Glass Company, Limited | Working fluid for heat cycle |
Cited By (2)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US20170003060A1 (en) * | 2014-03-17 | 2017-01-05 | Mitsubishi Electric Corporation | Air-conditioning apparatus |
US10001309B2 (en) * | 2014-03-17 | 2018-06-19 | Mitsubishi Electric Corporation | Air-conditioning apparatus |
Also Published As
Publication number | Publication date |
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EP3244156B1 (en) | 2021-02-24 |
JP6395863B2 (en) | 2018-09-26 |
JPWO2016110997A1 (en) | 2017-04-27 |
WO2016110997A1 (en) | 2016-07-14 |
CN107003081A (en) | 2017-08-01 |
EP3244156A4 (en) | 2018-08-29 |
EP3244156A1 (en) | 2017-11-15 |
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