EP1739377A1 - Heat transfer fin for heat exchanger - Google Patents
Heat transfer fin for heat exchanger Download PDFInfo
- Publication number
- EP1739377A1 EP1739377A1 EP05730153A EP05730153A EP1739377A1 EP 1739377 A1 EP1739377 A1 EP 1739377A1 EP 05730153 A EP05730153 A EP 05730153A EP 05730153 A EP05730153 A EP 05730153A EP 1739377 A1 EP1739377 A1 EP 1739377A1
- Authority
- EP
- European Patent Office
- Prior art keywords
- heat transfer
- fin
- heat
- foam metal
- heat exchanger
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
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- 238000012546 transfer Methods 0.000 title claims abstract description 219
- 239000011148 porous material Substances 0.000 claims abstract description 54
- 239000012530 fluid Substances 0.000 claims abstract description 9
- 229910052751 metal Inorganic materials 0.000 claims description 57
- 239000002184 metal Substances 0.000 claims description 57
- 239000006260 foam Substances 0.000 claims description 56
- 239000006262 metallic foam Substances 0.000 abstract 1
- XLYOFNOQVPJJNP-UHFFFAOYSA-N water Substances O XLYOFNOQVPJJNP-UHFFFAOYSA-N 0.000 description 37
- 239000003570 air Substances 0.000 description 32
- 239000003507 refrigerant Substances 0.000 description 24
- 230000007423 decrease Effects 0.000 description 20
- 238000001816 cooling Methods 0.000 description 19
- 238000002474 experimental method Methods 0.000 description 12
- 229910052782 aluminium Inorganic materials 0.000 description 9
- XAGFODPZIPBFFR-UHFFFAOYSA-N aluminium Chemical compound [Al] XAGFODPZIPBFFR-UHFFFAOYSA-N 0.000 description 9
- 238000012360 testing method Methods 0.000 description 8
- 238000004458 analytical method Methods 0.000 description 7
- 239000000463 material Substances 0.000 description 7
- 239000000126 substance Substances 0.000 description 7
- 230000000694 effects Effects 0.000 description 5
- 238000005259 measurement Methods 0.000 description 5
- 229910000838 Al alloy Inorganic materials 0.000 description 4
- 238000010792 warming Methods 0.000 description 4
- 238000011144 upstream manufacturing Methods 0.000 description 3
- RYGMFSIKBFXOCR-UHFFFAOYSA-N Copper Chemical compound [Cu] RYGMFSIKBFXOCR-UHFFFAOYSA-N 0.000 description 2
- 238000005452 bending Methods 0.000 description 2
- 230000000052 comparative effect Effects 0.000 description 2
- 230000005494 condensation Effects 0.000 description 2
- 238000009833 condensation Methods 0.000 description 2
- 229910052802 copper Inorganic materials 0.000 description 2
- 239000010949 copper Substances 0.000 description 2
- 238000005187 foaming Methods 0.000 description 2
- 238000000034 method Methods 0.000 description 2
- 238000005192 partition Methods 0.000 description 2
- 238000002834 transmittance Methods 0.000 description 2
- 239000012080 ambient air Substances 0.000 description 1
- 239000003795 chemical substances by application Substances 0.000 description 1
- 238000004891 communication Methods 0.000 description 1
- 238000011156 evaluation Methods 0.000 description 1
- 238000010438 heat treatment Methods 0.000 description 1
- 238000000465 moulding Methods 0.000 description 1
- 230000035699 permeability Effects 0.000 description 1
- 230000001737 promoting effect Effects 0.000 description 1
- 239000008400 supply water Substances 0.000 description 1
- 238000009423 ventilation Methods 0.000 description 1
Images
Classifications
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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
- F28F13/00—Arrangements for modifying heat-transfer, e.g. increasing, decreasing
- F28F13/003—Arrangements for modifying heat-transfer, e.g. increasing, decreasing by using permeable mass, perforated or porous materials
-
- 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
- F28D1/00—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators
- F28D1/02—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid
- F28D1/03—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with plate-like or laminated conduits
- F28D1/0308—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with plate-like or laminated conduits the conduits being formed by paired plates touching each other
- F28D1/0325—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with plate-like or laminated conduits the conduits being formed by paired plates touching each other the plates having lateral openings therein for circulation of the heat-exchange medium from one conduit to another
-
- 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
- F28D1/00—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators
- F28D1/02—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid
- F28D1/04—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits
- F28D1/047—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits the conduits being bent, e.g. in a serpentine or zig-zag
- F28D1/0477—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits the conduits being bent, e.g. in a serpentine or zig-zag the conduits being bent in a serpentine or zig-zag
- F28D1/0478—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits the conduits being bent, e.g. in a serpentine or zig-zag the conduits being bent in a serpentine or zig-zag the conduits having a non-circular cross-section
-
- 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
- F28D1/00—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators
- F28D1/02—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid
- F28D1/04—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits
- F28D1/053—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits the conduits being straight
- F28D1/0535—Heat-exchange apparatus having stationary conduit assemblies for one heat-exchange medium only, the media being in contact with different sides of the conduit wall, in which the other heat-exchange medium is a large body of fluid, e.g. domestic or motor car radiators with heat-exchange conduits immersed in the body of fluid with tubular conduits the conduits being straight the conduits having a non-circular cross-section
- F28D1/05366—Assemblies of conduits connected to common headers, e.g. core type radiators
Definitions
- the present invention relates to a heat transfer fin appropriate for use in a heat exchanger for air conditioners and other types of heat exchangers.
- Improvement in the heat transfer performance of an air side heat transfer fin for a heat exchanger or the like used in an air conditioner is an essential factor when miniaturizing the heat exchanger for saving energy in the system itself.
- a cross fin type heat exchanger including slit fins or louver fins has been proposed.
- Japanese Laid-Open Patent Publication No. 4-93595 and Japanese Laid-Open Patent Publication No. 9-26279 disclose such heat exchangers.
- slits or louvers are arranged for heat transfer fins made of thin plates having satisfactory heat transmittance such as aluminum plates, their front edges function to improve the heat transfer performance (heat transfer coefficient) with air.
- Japanese Laid-Open Patent Publication No. 2002-195774 proposes the use of a stacked type heat exchanger including flat heat transfer tubes and corrugated fins in an air conditioner.
- the entire heat exchanger including the flat heat transfer tubes and the corrugated fins as well as each part of the heat exchanger is shown in Figs. 31 and 32.
- a heat exchanger 10 includes pipe-shaped upper and lower headers 12A and 12B through which a refrigerant is flows in and out.
- the heat transfer tube 1 includes a plurality of refrigerant passages 2 having a square cross-section partitioned by partitions.
- the refrigerant flowing through the headers 12A and 12B from external refrigerant pipes 7 and 8 flows uniformly into each refrigerant passage 2 so that efficient heat exchange is immediately performed between the inside refrigerant and the outside air by the wide heat transfer area of the flat surface of the heat transfer tubes 1 and the corrugated fin 11.
- a plurality of louvers 11a and 11b for increasing the heat transfer efficiency with air are formed in the corrugated fins 11 along a plane from the upstream side to the downstream side and configured so that the front edges function to immediately improve the heat exchange performance between the refrigerant and the air.
- the present invention is configured as below to achieve the above object.
- a first solution according to the present invention is a heat transfer fin arranged on a heat transfer tube through which fluid flows for exchanging heat with air.
- the heat transfer fin contacts air to exchange heat.
- the heat transfer fin is made of foam metal having a pore density of 20 PIP or greater.
- the foam metal has an open cell type porous structure with fine linear grooves continuously connected to one another enabling fluid to flow therethrough.
- the surface area per unit volume is large. Therefore, the heat transfer area of the foam metal is large. The heat transfer is enhanced by disturbance in the fluid since the foam metal has a complex passage.
- the foam metal includes linear grooves, a temperature boundary layer can be easily renewed, and a high heat transfer coefficient can be obtained.
- the heat transfer performance of the heat transfer fin thus becomes extremely high.
- the heat exchange performance of a heat exchanger is greatly improved when employing the heat transfer fin made of foam metal.
- the passage configuration is complex in the foam metal and the pressure loss is large.
- an optimal pore density must be determined when employing foam metal as the material for the heat transfer fin. According to the results of various analyses and experiments, the more preferable pore density for maximizing the heat transfer property of the foam metal is 20 PIP or greater.
- the heat transfer tube through which fluid flows for exchanging heat with air is provided as a plurality, and the plurality of heat transfer tubes are set at an interval of 12 mm or less.
- the heat transfer fin made of foam metal has a superior heat transfer performance since the surface area per unit volume is large and the heat transfer area is large.
- the fin efficiency is low compared to a louver fin or the like because the foam metal has fine linear grooves. Therefore, the interval of the plurality of heat transfer tubes must be optimized. According to the results of various analyses, the interval of the plurality of heat transfer tubes is effective when it is 12 mm or less, and the heat transfer performance is sufficiently improved especially when the pore density is 20 PIP or greater.
- a heat exchanger is a stacked type heat exchanger.
- the stacked type heat exchanger is configured so that the heat transfer tube is flat and extends in the air flowing direction, and so that the heat transfer fin arranged in between is sufficiently long in the air flowing direction. Therefore, the stacked type heat exchanger itself has high heat transfer performance.
- the heat transfer performance further improves when employing the heat transfer fin made of foam metal in the stacked air-heat exchanger.
- a heat exchanger having a high heat exchange capability and appropriate for use in an air conditioner is formed with low cost, and miniaturization of the heat exchanger is achieved.
- the pore density is 20 PIP or greater and 60 PIP or less.
- the optimal upper limit value of the bore density is obtained through various analyses and experiments.
- an interval H between the plurality of heat transfer tubes is set at 4 mm or greater and 12 mm or less.
- the foam metal has superior heat transfer performance since the surface area per unit volume is large and the heat transfer area is large.
- the fin efficiency is low. Therefore, the interval for the plurality of heat transfer tubes must be optimized. From the results of various experiments, the heat transfer performance improved when the interval of the heat transfer tubes was 4 mm or greater and 12 mm or less.
- Figs. 1 and 2 show the configuration for an entire heat exchanger and the main part thereof according to preferred embodiment 1 of the present invention.
- the heat exchanger 10 is configured so that a plurality of parallel flat heat transfer tubes 1 extend between pipe shaped upper and lower headers 12A and 12b through which refrigerant flows in and out.
- the heat transfer tubes 1 are connected to the headers 12A and 12B and extend in a direction orthogonal to the headers 12A and 12B.
- the heat transfer fin 13 is not the corrugated fin of the prior art and is made of foam metal of open cell type having a porous structure, as shown in Fig. 2, that allows fluid to flow therethrough.
- the heat transfer tube 1 has a plurality of refrigerant passage grooves with square cross-sections partitioned by partitions in the same manner as the prior art shown in Fig. 32.
- Refrigerant which is drawn into and distributed by the external refrigerant piping 7 or 8, flows uniformly through each refrigerant passage groove from the upper side toward the lower side or from the lower side toward the upper side via the upper header 12A or the lower header 12B.
- Heat exchange is immediately and efficiently performed with the ambient air through the heat transfer surface of the refrigerant flowing through each refrigerant passage and the fin surface with the porous structure in the heat transfer fin 13 made of foam metal.
- the foam metal forming the heat transfer fin 13 is a porous substance. Therefore, the foam metal has a large heat transfer area since the surface area per unit volume is large and complex passages are formed therein. Thus, effective heat transfer promotion may be expected due to the disturbance of the fluid. Since foam metal has many fine linear groves connected to each other (see structure Fig. 2), a temperature boundary layer may easily be renewed, and an extremely high heat transfer coefficient can be obtained.
- the heat exchanger 10 having such a configuration is used, for example, as a condenser, the refrigerant introduced from the external refrigerant piping 7 via the upper header 12A is uniformly distributed to flowed from the upper side to the lower side of the heat transfer tube 1, and discharged from the outer refrigerant pipe 8 through the lower header 12B. If the heat exchanger 10 is used as an evaporator, the refrigerant flows in the opposite direction.
- the heat transfer tubes 1 are flat and elongated in the air flow direction, and the heat transfer fins 13 arranged in between also extend in the air flowing direction.
- the heat transfer fins 13 are easily formed by foaming and molding a metal having a high heat transfer coefficient, such as aluminum or copper, into a shape that can be brazed.
- the heat exchanger suitable for use in an air conditioner may be formed with a reduced size at a lower cost and with a high heat exchange performance.
- the foam metal forming the heat transfer fin 13 of the present embodiment is a porous substance, in which the surface area per unit area expands as the pore density PPI becomes higher as in (A) 10 PPI, (B) 20 PPI and (C) 40 PPI shown in Figs. 3(A), (B) and (C).
- the PPI pore per inch
- the PPI represents the density of gas bubbles per cubic inch.
- the pore density PPI of the foam metal was found to be generally preferable at 20 PPI (Figs. 3(B), (C)) or greater and 60 PPI or less.
- the interval (fin width) H of the heat transfer tube 1 must have an optimal value. According to the results of various analyses and experiments, the interval H of the heat transfer tube 1 was found to be optimal in the range of 4 mm or greater and 12 mm or less.
- the interval H of the heat transfer tube 1 is effective when it is 4 mm or greater and 12 mm or less, as described below.
- the heat transfer performance QN (W/m 3 ) per unit volume with respect to the front surface wind velocity V f greatly improves compared to that for the louver fin, as shown in the graph of Fig. 4, if the interval H of the heat transfer tube 1 is 4 mm or greater and 12 mm or less when the pore density of the heat transfer fin 13 made of foam metal is greater than or equal to 20 PPI and less than or equal to 60 PPI.
- the dimensions 5 mm, 8 mm, 12 mm in the examples of Figs. 4 and 5 are three sets of dimensional data used to analyze the interval (hereinafter considered as width of heat transfer fin 13) H of the heat transfer tube 1.
- a foam metal made of aluminum (aluminum alloy 6010) was used for the heat transfer fin as the open cell type tested foam metal.
- Three types of the foam metal having a pore density PPI of, for example, No. 1, 10 PPI shown in Fig. 3(A); No. 2, 20 PPI shown in Fig. 3(B); and No. 3, 40 PPI shown in Fig. 3(C) were prepared.
- three heat transfer tubes having different width dimensions (i.e., interval of heat transfer tube 1) H of 5 mm, 8 mm, and 12 mm were prepared for each of the three types of bore density PPI.
- heat exchange was performed between the refrigerant (warm water as one example) at the heat transfer tube side 1 in the configuration of Fig. 1 and the air flowing outside.
- the pressure loss and the heat transfer coefficient of the heat transfer fin for each case were experimentally obtained, and analysis was performed to clarify the basic heat transfer property and the influence of the wall surface of the heat transfer tube 1 serving as the heat source.
- the experiments were conducted under the conditions of 20°C for air temperature and 50% for relative humidity.
- the measurement of the pressure loss was performed in a non-load condition in which no warm water was supplied to the heat transfer tube 1, and the measurement of the heat transfer coefficient was performed by supplying warm water of 50°C as a warm heat source.
- the wind velocity range was 0.5 to 2.3 m/s in terms of the wind velocity V f at the front surface side (upstream side) of the heat transfer fin 13.
- the specific material of the aluminum foam metal used in this experiment was aluminum alloy 6101, as mentioned above.
- the detailed specification is shown in [Table 1].
- fins of the three widths H of 5 mm, 8 mm, and 12 mm were prepared.
- the height L of the foam metal was 89 mm
- the depth D was 13 mm
- the surface area per unit volume was ⁇ .
- the graph of Fig. 6 shows the relationship of the pressure loss P(Pa) with respect to the front surface wind velocity V f (m/s).
- the pressure loss ⁇ P increases as the pore density PIP increases, or as the pore size d pore decreases. Further, the pressure loss ⁇ P increases as the fin width H decreases. This is because the surface area (including wall surface) per unit volume increases as the pore size d pore decreases and the fin width H decreases.
- the pressure loss property of the foam metal is expressed as follows using the permeability (K) and the Ergun coefficient (C E ).
- K permeability
- C E Ergun coefficient
- the graph of Fig. 7 is obtained, for example, when K and CE obtained through the least-square method using equation (1) are shown in relation with the pore size d pore .
- K increases as the pore size d pore increase
- the fin width H increases as the friction caused by the wall surface decreases.
- C E slightly decreases as the pore size d pore increases.
- a definite tendency cannot be found in the influence of the fin width H.
- Fig. 8 shows change in the friction loss coefficient f with respect to Re K .
- the graph of Fig. 9 shows the product of the measured heat transfer coefficient ha and the surface area ⁇ per unit area.
- Qa is the heat transfer amount
- At is the total heat transfer area combining the surface area of the foam metal and the area of the wall surface
- ⁇ T LMTD is the logarithmic mean temperature difference.
- ha ⁇ represents the heat transfer performance per unit volume. It is apparent here that the heat transfer performance increases as the pore density PPI increases and the smaller the fin width H decreases. In particular, the heat transfer performance becomes higher than the conventional louver fin in the 20 PIP No. 2 fin sample and the 40 PIP No. 3 fin sample. This suggesting the possibility of sufficient miniaturization of the heat exchanger.
- the heat transfer coefficient h o should not include the fin efficiency for optimal designing.
- the heat transfer coefficient h o cannot be easily obtained due to the complex passage shape resulting from foam metal. Accordingly, h o is obtained through the following approximation method.
- the pressure loss and the heat transfer coefficient in this case were experimentally obtained and analyzed to clarify the basic heat transfer property and the influence of the wall surface of the heat transfer tube 1 serving as the heat source.
- the experiments were conducted under the conditions of 20°C for air temperature and 50% for relative humidity.
- the measurement of the pressure loss was performed in a non-load condition in which no cold water was supplied to the heat transfer tube 1, and the measurement of the heat transfer coefficient was performed by supplying cold water of 50°C as a cold heat source.
- the wind velocity range was 0.5 to 2.3 m/s in terms of the wind velocity V f at the front surface side (upstream side) of the heat transfer fin 13.
- the specific material for the aluminum foam metal used in this experiment is aluminum alloy 6101, as mentioned above.
- fins of the three widths H of 5 mm, 8 mm, and 12 mm were prepared.
- the height L of the foam metal in the vertical direction was 89 mm
- the depth D was 13 mm
- the surface area per unit volume was ⁇ .
- the graph of Fig. 12 shows the pressure loss ⁇ P/D(Pa/m) with respect to the front surface wind velocity V f (m/s) in the dry state.
- the pressure loss is calculated by including the air flow direction length D (m) of at the fins ( ⁇ P/D). In the following description, this is simply referred to as ⁇ P.
- Fig. 13 shows the pressure loss ⁇ P (Pa) with respect to the front surface wind velocity V f (m/s) in the wet state in when the temperature of the water serving as the refrigerant is 5°C
- Fig. 14 shows the same state when the temperature of the water serving as the refrigerant is 10°C.
- the influence of the pore density PPI and the fin width H is substantially the same as the tendency in the dry state.
- the value of the pressure loss ⁇ P greatly increases compared to that in the dry state (see Fig. 12).
- the pressure loss ⁇ P greatly increases because the condensed water accumulated on the fin surface becomes a ventilation resistance, and thus can be predicted that water drainage becomes an essential factor in comparison to the dry state.
- the ratio of the pressure loss ⁇ P between the wet state and the dry state is shown in Fig. 15 (for a water temperature of 5°C) and Fig. 16 (for a water temperature of 10°C). In the case of water temperature of 5°C in Fig. 15, the ratio of the pressure loss gradually increases as a whole when the air flow rate increases but the pressure loss decreases when the air flow rate increases in the cases of 8 mm, 12 mm at 10°C (Fig. 16).
- the ratio ⁇ P wet / ⁇ P dry of the pressure loss ⁇ P in a dry state and in a wet state increases as the pore density PPI increases.
- the ratio is greater for the foam metal fin than for the louver fin. That is, the water drainage is poorer in the foam metal fin than the louver fin.
- the foam metal fin used in the present test is in an experimental level in which the fin surface is not processed.
- the problem of water drainage is sufficiently improved, for example, by applying a hydrophilic agent.
- Fig. 17 shows the relationship of the heat transfer coefficient h dry with respect to the front surface wind velocity V f (m/s) in a dry state.
- the heat transfer coefficient h dry increases as the pore density PIP decreases, and the heat transfer coefficient dry decreases as the fin width H increases.
- Fig. 18 shows the relationship between h dry ⁇ and the front surface wind velocity V f (m/s) in the dry state. It is apparent that the heat transfer performance increases as the pore density PIP increases and the fin width H decreases.
- Figs. 20 and 21 show the relationship of the heat transfer coefficient h wet in the wet state with respect to the front surface wind velocity V f (m/s) in the wet state when the temperature of cold water is 5°C and 10°C.
- V f m/s
- h wet is slightly smaller than h dry .
- this is because the fin efficiency decreases in a dry state due to the heat transfer of latent heat resulting from condensation of the moisture in air.
- Fig. 22 water temperature 5°C
- Fig. 23 water temperature 10°C
- Fig. 24 water temperature 5°C
- Fig. 25 water temperature 10°C
- the substance transfer coefficient h mass of Fig. 24 increases as the entire air flow rate and fin width H increase and as the pore density PIP decreases.
- the foam metal fin of the present embodiment has higher pressure loss and higher heat transfer coefficient per volume compared to the existing louver fin.
- the pressure loss and the heat transfer coefficient must be comprehensively analyzed.
- V is the volume
- a c is the flow cross-sectional area
- Figs. 26 and 27 show the heat transfer coefficients h dry ⁇ , h wet ⁇ per unit volume for the dry state and the wet state in relation to the pump power EBA necessary per unit volume.
- the wet state of Fig. 27, for 40 PPI, H 5 mm, 8 mm, and 12 mm, the heat transfer performance improvement effect is higher by about 28%.
- Fig. 29 shows the structure for an air-heat exchanger according to a second embodiment of the present invention.
- the present embodiment relates to a serpentine heat exchanger in which a flat heat transfer tube 21 is bent into a serpentine shape as a single continuous structure.
- the high heat transfer performance is also achieved in the heat exchanger of such configuration in the same manner as the heat exchanger of the first embodiment described above.
- Fig. 30 shows the structure for a heat exchanger according to a third embodiment of the present invention.
- a plurality of plate shaped heat transfer tubes 31 extending in the horizontal direction are connected in a stack form by left and right connecting members 22, and the heat transfer fin 13 made of foam metal is arranged between the connecting members 22 of each layer.
- Each heat transfer tube 31 includes refrigerant inlet and outlet holes 23.
- the holes 23 connected through the connecting members 22 form a refrigerant passage.
- the connecting members 22 are used for a structure similar to that of the first embodiment in a stacked plate type air-heat exchanger.
- the heat transfer fin of the present invention is not limited to the structure of the heat exchanger in each embodiment and may obviously be applied to a heat transfer fin for performing heat exchange with air, such as a cross fin type or the like.
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- Physics & Mathematics (AREA)
- Thermal Sciences (AREA)
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- General Engineering & Computer Science (AREA)
- Chemical & Material Sciences (AREA)
- Dispersion Chemistry (AREA)
- Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
Abstract
A heat transfer fin (13) is provided on a heat transfer tube (1) wherein a fluid for exchanging heat with air flows. The heat transfer fin which transfers heat by having contact with air is composed of a metal foam having a pore density of 20 PIP or more.
Description
- The present invention relates to a heat transfer fin appropriate for use in a heat exchanger for air conditioners and other types of heat exchangers.
- Improvement in the heat transfer performance of an air side heat transfer fin for a heat exchanger or the like used in an air conditioner is an essential factor when miniaturizing the heat exchanger for saving energy in the system itself.
- To improve the heat transfer performance, a cross fin type heat exchanger including slit fins or louver fins has been proposed.
Japanese Laid-Open Patent Publication No. 4-93595 Japanese Laid-Open Patent Publication No. 9-26279 -
Japanese Laid-Open Patent Publication No. 2002-195774 - A
heat exchanger 10 includes pipe-shaped upper andlower headers 12A and 12B through which a refrigerant is flows in and out. A plurality of parallel flatheat transfer tubes 1, which extend between theheaders 12A and 12B in a direction orthogonal to theheaders headers 12A and 12B.Corrugated fins 11, which are formed by continuously bending a flat aluminum plate or the like, is arranged between theheat transfer tubes 1 to join the adjacentheat transfer tubes 1. - As shown in Fig. 32, the
heat transfer tube 1 includes a plurality ofrefrigerant passages 2 having a square cross-section partitioned by partitions. The refrigerant flowing through theheaders 12A and 12B fromexternal refrigerant pipes refrigerant passage 2 so that efficient heat exchange is immediately performed between the inside refrigerant and the outside air by the wide heat transfer area of the flat surface of theheat transfer tubes 1 and thecorrugated fin 11. - A plurality of
louvers corrugated fins 11 along a plane from the upstream side to the downstream side and configured so that the front edges function to immediately improve the heat exchange performance between the refrigerant and the air. - However, improvement in the heat transfer performance has limits when just forming slits and louvers in heat transfer fins made of thin aluminum plates regardless of the form of the heat transfer fin used. Selecting the material for the heat transfer fin and improving the heat transfer performance of the heat exchanger are thus being considered.
- It is an object of the present invention to provide heat transfer fins for a heat exchanger that significantly improves the heat transfer performance by forming air side heat transfer fins for a heat exchanger from a foam metal manufactured by foaming copper, aluminum, or the like, which have high heat transmittance.
- The present invention is configured as below to achieve the above object.
- A first solution according to the present invention is a heat transfer fin arranged on a heat transfer tube through which fluid flows for exchanging heat with air. The heat transfer fin contacts air to exchange heat. The heat transfer fin is made of foam metal having a pore density of 20 PIP or greater.
- The foam metal has an open cell type porous structure with fine linear grooves continuously connected to one another enabling fluid to flow therethrough. Thus, the surface area per unit volume is large. Therefore, the heat transfer area of the foam metal is large. The heat transfer is enhanced by disturbance in the fluid since the foam metal has a complex passage.
- Furthermore, since the foam metal includes linear grooves, a temperature boundary layer can be easily renewed, and a high heat transfer coefficient can be obtained. The heat transfer performance of the heat transfer fin thus becomes extremely high.
- Therefore, the heat exchange performance of a heat exchanger is greatly improved when employing the heat transfer fin made of foam metal.
- However, the passage configuration is complex in the foam metal and the pressure loss is large. Thus, an optimal pore density must be determined when employing foam metal as the material for the heat transfer fin. According to the results of various analyses and experiments, the more preferable pore density for maximizing the heat transfer property of the foam metal is 20 PIP or greater.
- In a second solution according to the present invention, the heat transfer tube through which fluid flows for exchanging heat with air is provided as a plurality, and the plurality of heat transfer tubes are set at an interval of 12 mm or less.
- As described above, the heat transfer fin made of foam metal has a superior heat transfer performance since the surface area per unit volume is large and the heat transfer area is large. However, the fin efficiency is low compared to a louver fin or the like because the foam metal has fine linear grooves. Therefore, the interval of the plurality of heat transfer tubes must be optimized. According to the results of various analyses, the interval of the plurality of heat transfer tubes is effective when it is 12 mm or less, and the heat transfer performance is sufficiently improved especially when the pore density is 20 PIP or greater.
- In a third solution according to the present invention, a heat exchanger is a stacked type heat exchanger.
- The stacked type heat exchanger is configured so that the heat transfer tube is flat and extends in the air flowing direction, and so that the heat transfer fin arranged in between is sufficiently long in the air flowing direction. Therefore, the stacked type heat exchanger itself has high heat transfer performance.
- The heat transfer performance further improves when employing the heat transfer fin made of foam metal in the stacked air-heat exchanger.
- According to the above structure, a heat exchanger having a high heat exchange capability and appropriate for use in an air conditioner is formed with low cost, and miniaturization of the heat exchanger is achieved.
- In a fourth solution according to the present invention, the pore density is 20 PIP or greater and 60 PIP or less.
- In addition to the first solution, the optimal upper limit value of the bore density is obtained through various analyses and experiments.
- In a fifth solution according to the present invention, an interval H between the plurality of heat transfer tubes is set at 4 mm or greater and 12 mm or less.
- As described above, the foam metal has superior heat transfer performance since the surface area per unit volume is large and the heat transfer area is large. However, the fin efficiency is low. Therefore, the interval for the plurality of heat transfer tubes must be optimized. From the results of various experiments, the heat transfer performance improved when the interval of the heat transfer tubes was 4 mm or greater and 12 mm or less.
-
- Fig. 1 is a perspective view showing the configuration of a heat exchanger according to a first embodiment of the present invention;
- Fig. 2 is an enlarged view showing the configuration of the material structure of a heat transfer fin made of foam metal;
- Fig. 3 is an enlarged view showing samples having different structure densities for the material structure of the heat transfer fin made of foam metal in which (A) shows a first (No. 1) sample, (B) shows a second (No. 2) sample, and (C) shows a third (No. 3) sample (pore ratios of 10 PPI, 20 PPI, an 40 PPI);
- Fig. 4 is a graph showing the relationship between the front surface side wind velocity Vf and the heat transfer performance QN per unit volume in the heat transfer fin;
- Fig. 5 is a graph showing the relationship between the heat transfer performance per unit volume and the pump power in the heat transfer fin;
- Fig. 6 is a graph showing the relationship of the pressure loss ΔP with respect to the front surface side wind velocity Vf in the heat transfer fin;
- Fig. 7 is a graph showing the pressure loss property when warming the foam metal in the heat transfer fin;
- Fig. 8 is a graph showing changes in a friction loss coefficient f when warming the heat transfer fin;
- Fig. 9 is a graph showing the relationship between the heat transfer coefficient and the surface area β per unit volume when warming the heat transfer fin;
- Fig. 10 is a graph showing the relationship between ho and the front surface side wind velocity Vf when warming the heat transfer fin;
- Fig. 11 is a graph showing the relationship between the calculation result obtained from equation (10) in the description of the operation of the heat transfer fin and the experiment result;
- Fig. 12 is a graph showing the relationship between the pressure loss (ΔP/D) with respect to the front surface side wind velocity (Vf) when dry cooling the heat transfer fin;
- Fig. 13 is a graph showing the relationship between the pressure loss (ΔP/D) with respect to the front surface side wind velocity (Vf) when wet cooling (
water temperature 5°C) the heat transfer fin; - Fig. 14 is a graph showing the relationship between the pressure loss (ΔP/D) with respect to the front surface side wind velocity (Vf) when wet cooling (
water temperature 10°C) the heat transfer fin; - Fig. 15 is a graph showing the ratio ΔPwet/ΔPdry of the pressure loss ΔPwet during wet cooling (
water temperature 5°C) with respect to the pressure loss ΔPdry during dry cooling (5°C) in the heat transfer fin; - Fig. 16 is a graph showing the ratio ΔPwet/ΔPdry of the pressure loss ΔPwet during wet cooling (
water temperature 10°C) with respect to the pressure loss ΔPdry during dry cooling (15°C) in the heat transfer fin; - Fig. 17 is a graph showing the relationship between the heat transfer coefficient hdry and the front surface side wind velocity Vf when dry cooling the heat transfer fin;
- Fig. 18 is a graph showing the relationship between the heat transfer performance hdryβ per unit volume and the front surface wind velocity (Vf) when dry cooling the heat transfer fin;
- Fig. 19 is a graph showing the heat transfer performance hdryβ per unit volume when dry cooling the heat transfer fin in comparison with the louver fin;
- Fig. 20 is a graph showing the heat transfer coefficient hwet when wet cooling (
water temperature 5°C) the heat transfer fin; - Fig. 21 is a graph showing the heat transfer coefficient hwet when wet cooling (
water temperature 10°C) the heat transfer fin; - Fig. 22 is a graph showing the heat transfer performance hwet per unit volume when wet cooling (
water temperature 5°C) the heat transfer fin; - Fig. 23 is a graph showing the heat transfer performance hwet per unit volume when wet cooling (
water temperature 10°C) heat transfer fin; - Fig. 24 is a graph showing the relationship between the substance transfer coefficient hmass and the front surface side wind velocity Vf when wet cooling (
water temperature 5°C) the heat transfer fin; - Fig. 25 is a graph showing the relationship between the substance transfer coefficient hmass and the front surface side wind velocity Vf when wet cooling (
water temperature 10°C) the heat transfer fin; - Fig. 26 is a graph showing the relationship between the heat transfer performance hdryβ per unit volume and the necessary power Eβ per unit volume when dry cooling the heat transfer fin;
- Fig. 27 is a graph showing the relationship between the heat transfer performance hdryβ per unit volume and the necessary power Eβ per unit volume when wet cooling the heat transfer fin;
- Fig. 28 is a graph showing the relationship between the mass transfer performance hmassβ per unit volume and the necessary power Eβ per unit volume when cooling the heat transfer fin;
- Fig. 29 is a perspective view showing the structure configuration of a heat exchanger according to a second embodiment of the present invention;
- Fig. 30 is a perspective view showing a configuration of a heat exchanger according to a third embodiment of the present invention;
- Fig. 31 is a perspective view showing a configuration of an air-heat exchanger in the prior art; and
- Fig. 32 is a partially cut-away perspective view showing the configuration of the main part of the conventional air-heat exchanger.
- Figs. 1 and 2 show the configuration for an entire heat exchanger and the main part thereof according to
preferred embodiment 1 of the present invention. As shown in Fig. 1, theheat exchanger 10 is configured so that a plurality of parallel flatheat transfer tubes 1 extend between pipe shaped upper andlower headers 12A and 12b through which refrigerant flows in and out. Theheat transfer tubes 1 are connected to theheaders 12A and 12B and extend in a direction orthogonal to theheaders 12A and 12B.Corrugated fins 11, which are formed by continuously bending a flat aluminum plate or the like, is joined to the adjacentheat transfer tube 1 between theheat transfer tubes 1. - In this embodiment, the
heat transfer fin 13 is not the corrugated fin of the prior art and is made of foam metal of open cell type having a porous structure, as shown in Fig. 2, that allows fluid to flow therethrough. - The
heat transfer tube 1 has a plurality of refrigerant passage grooves with square cross-sections partitioned by partitions in the same manner as the prior art shown in Fig. 32. Refrigerant, which is drawn into and distributed by the externalrefrigerant piping upper header 12A or the lower header 12B. Heat exchange is immediately and efficiently performed with the ambient air through the heat transfer surface of the refrigerant flowing through each refrigerant passage and the fin surface with the porous structure in theheat transfer fin 13 made of foam metal. - The foam metal forming the
heat transfer fin 13 is a porous substance. Therefore, the foam metal has a large heat transfer area since the surface area per unit volume is large and complex passages are formed therein. Thus, effective heat transfer promotion may be expected due to the disturbance of the fluid. Since foam metal has many fine linear groves connected to each other (see structure Fig. 2), a temperature boundary layer may easily be renewed, and an extremely high heat transfer coefficient can be obtained. - When the
heat exchanger 10 having such a configuration is used, for example, as a condenser, the refrigerant introduced from the externalrefrigerant piping 7 via theupper header 12A is uniformly distributed to flowed from the upper side to the lower side of theheat transfer tube 1, and discharged from the outerrefrigerant pipe 8 through the lower header 12B. If theheat exchanger 10 is used as an evaporator, the refrigerant flows in the opposite direction. - In such a stacked type heat exchanger, the
heat transfer tubes 1 are flat and elongated in the air flow direction, and theheat transfer fins 13 arranged in between also extend in the air flowing direction. Thus, the heat transfer performance is large. Theheat transfer fins 13 are easily formed by foaming and molding a metal having a high heat transfer coefficient, such as aluminum or copper, into a shape that can be brazed. - Therefore, the heat exchanger suitable for use in an air conditioner may be formed with a reduced size at a lower cost and with a high heat exchange performance.
- The foam metal forming the
heat transfer fin 13 of the present embodiment is a porous substance, in which the surface area per unit area expands as the pore density PPI becomes higher as in (A) 10 PPI, (B) 20 PPI and (C) 40 PPI shown in Figs. 3(A), (B) and (C). However, the pressure loss increases. Therefore, the determination of an optimal range for the pore density PPI is important when using foam metal for theheat transfer fin 13 of the air-heat exchanger described above. The PPI (pore per inch) represents the density of gas bubbles per cubic inch. - Through the results of various analyses and experiments, the pore density PPI of the foam metal was found to be generally preferable at 20 PPI (Figs. 3(B), (C)) or greater and 60 PPI or less.
- A problem in which the fin efficiency is low compared to the louver fin of the prior art arises for the
heat transfer fin 13 made of foam metal since the diameter of the linear grooves connected to one another is small. Therefore, the interval (fin width) H of theheat transfer tube 1 must have an optimal value. According to the results of various analyses and experiments, the interval H of theheat transfer tube 1 was found to be optimal in the range of 4 mm or greater and 12 mm or less. - That is, according to the experiment results, the interval H of the
heat transfer tube 1 is effective when it is 4 mm or greater and 12 mm or less, as described below. In particular, the heat transfer performance QN (W/m3) per unit volume with respect to the front surface wind velocity Vf greatly improves compared to that for the louver fin, as shown in the graph of Fig. 4, if the interval H of theheat transfer tube 1 is 4 mm or greater and 12 mm or less when the pore density of theheat transfer fin 13 made of foam metal is greater than or equal to 20 PPI and less than or equal to 60 PPI. Furthermore, when the interval H of theheat transfer tube 1 and the bore density of the heat transfer fin are as described above, the heat transfer amount per unit volume at the same power increases by about 25% compared to that for the louver fin, as shown in the graph of Fig. 5. This obtains the improvement effect for the effective heat transfer performance. - As will be described below, the
dimensions 5 mm, 8 mm, 12 mm in the examples of Figs. 4 and 5 are three sets of dimensional data used to analyze the interval (hereinafter considered as width of heat transfer fin 13) H of theheat transfer tube 1. - Several experiments were performed to check the improvement effect for the effective heat transfer performance of the
heat transfer fin 13 made of foam metal. - A foam metal made of aluminum (aluminum alloy 6010) was used for the heat transfer fin as the open cell type tested foam metal. Three types of the foam metal having a pore density PPI of, for example, No. 1, 10 PPI shown in Fig. 3(A); No. 2, 20 PPI shown in Fig. 3(B); and No. 3, 40 PPI shown in Fig. 3(C) were prepared. Furthermore, three heat transfer tubes having different width dimensions (i.e., interval of heat transfer tube 1) H of 5 mm, 8 mm, and 12 mm were prepared for each of the three types of bore density PPI. For the total of nine types of samples, heat exchange was performed between the refrigerant (warm water as one example) at the heat
transfer tube side 1 in the configuration of Fig. 1 and the air flowing outside. - The pressure loss and the heat transfer coefficient of the heat transfer fin for each case were experimentally obtained, and analysis was performed to clarify the basic heat transfer property and the influence of the wall surface of the
heat transfer tube 1 serving as the heat source. - The experiments were conducted under the conditions of 20°C for air temperature and 50% for relative humidity. The measurement of the pressure loss was performed in a non-load condition in which no warm water was supplied to the
heat transfer tube 1, and the measurement of the heat transfer coefficient was performed by supplying warm water of 50°C as a warm heat source. The wind velocity range was 0.5 to 2.3 m/s in terms of the wind velocity Vf at the front surface side (upstream side) of theheat transfer fin 13. - The specific material of the aluminum foam metal used in this experiment was aluminum alloy 6101, as mentioned above. The detailed specification is shown in [Table 1]. In order to check the influence of the interval H between the wall surfaces for the foam metal No. 1 to No. 3 (10, 20, 40 PPI) of the three types of pore densities described above, fins of the three widths H of 5 mm, 8 mm, and 12 mm were prepared. Thus, a total of nine types of samples were prepared. The height L of the foam metal was 89 mm, the depth D was 13 mm, and the surface area per unit volume was β.
[Table 1] Sample Pore Ratio PPI Porosity Surface Area β (m2/m3) Pore Size dpore (mm) Width H (mm) No. 1 10 PPI 0.914 791 2.5 5 8 12 No. 2 20 PPI 0.915 1744 1.25 5 8 12 No. 3 40 PPI 0.908 2740 0.65 5 8 12 - The graph of Fig. 6 shows the relationship of the pressure loss P(Pa) with respect to the front surface wind velocity Vf(m/s). The pressure loss ΔP increases as the pore density PIP increases, or as the pore size dpore decreases. Further, the pressure loss ΔP increases as the fin width H decreases. This is because the surface area (including wall surface) per unit volume increases as the pore size dpore decreases and the fin width H decreases. The foam metal has a higher pressure loss ΔP than a louver fin (width H=7.9 mm, D=13.6 mm, fin pitch=1.5 mm) of a comparative example.
-
- The graph of Fig. 7 is obtained, for example, when K and CE obtained through the least-square method using equation (1) are shown in relation with the pore size dpore. K increases as the pore size dpore increase, and the fin width H increases as the friction caused by the wall surface decreases. CE slightly decreases as the pore size dpore increases. However, a definite tendency cannot be found in the influence of the fin width H.
-
- Fig. 8 shows change in the friction loss coefficient f with respect to ReK.
-
- Qa is the heat transfer amount, At is the total heat transfer area combining the surface area of the foam metal and the area of the wall surface, and ΔTLMTD is the logarithmic mean temperature difference. Further, haβ represents the heat transfer performance per unit volume. It is apparent here that the heat transfer performance increases as the pore density PPI increases and the smaller the fin width H decreases. In particular, the heat transfer performance becomes higher than the conventional louver fin in the 20 PIP No. 2 fin sample and the 40 PIP No. 3 fin sample. This suggesting the possibility of sufficient miniaturization of the heat exchanger.
- Although the above heat transfer coefficient ha includes the fin efficiency, the heat transfer coefficient ho should not include the fin efficiency for optimal designing. However, the heat transfer coefficient ho cannot be easily obtained due to the complex passage shape resulting from foam metal. Accordingly, ho is obtained through the following approximation method.
-
- P is the perimeter, A is the cross-sectional area, and k is the heat conductivity. Since P/A of equation (6) is unknown for the foam metal, the value of m is assumed as in equation (7), ho is assumed as being constant due to dimension H, and the values of C are obtained for each pore density PPI (10 PPI, 20 PPI, 40 PPI) of No. 1 to No. 3 from experimental data in which the dimension H is changed. The results are shown in [Table 2].
[Table 2] 10 PPI 20 PPI 40 PPI C 30900 68000 94800 - From the above results, it is apparent that C increases as the pore density PPI increases. This is because the line diameter of line material increases as the pore density PPI increases and thereby increases P/A. The result indicating ho with respect to the front surface wind velocity Vf is the graph of Fig. 10. As apparent from the graph, ho increases as the pore density PPI decreases. However, this effect is barely seen when the
pore density PIP 20 PIP or less. Further, dpore and √K are considered as the characteristic lengths of the porous material. However, since K changes in accordance with the value of the interval H between the wall surfaces when influenced by the wall surface, as shown in Fig. 7, the value of K at H=12 mm (K12) in which the influence of the wall surface is minimum is used for the characteristic length, and the non-dimensional variable is defined as in the following equation. -
- The contrast of the correlation equation of equation (10) and the actual experimental results are shown in Fig. 11. As a result, it is apparent that they frequently match, and 90% of the data is within an error of ±6%.
- In this case, in the same manner as in the first test example described above, aluminum alloy 6010 was used as the foam metal forming the heat transfer fin. Three types of the foam metal having a pore density PIP of No. 1, 10 PIP; No. 2, 20 PIP; and No. 3, 40 PIP were prepared. Furthermore, three heat transfer tubes having different width dimensions (i.e., interval of heat transfer tube 1) H of 5 mm, 8 mm, and 12 mm were prepared for each of the three types of bore density PIP. For the total of nine types of samples, heat exchange was performed between the refrigerant (cold water as one example) at the heat
transfer tube side 1 in the configuration of Fig. 1 and the air flowing outside. - The pressure loss and the heat transfer coefficient in this case were experimentally obtained and analyzed to clarify the basic heat transfer property and the influence of the wall surface of the
heat transfer tube 1 serving as the heat source. - The experiments were conducted under the conditions of 20°C for air temperature and 50% for relative humidity. The measurement of the pressure loss was performed in a non-load condition in which no cold water was supplied to the
heat transfer tube 1, and the measurement of the heat transfer coefficient was performed by supplying cold water of 50°C as a cold heat source. The wind velocity range was 0.5 to 2.3 m/s in terms of the wind velocity Vf at the front surface side (upstream side) of theheat transfer fin 13. - The specific material for the aluminum foam metal used in this experiment is aluminum alloy 6101, as mentioned above. In the same manner as in the first test example, in order to check the influence of the interval H between the wall surfaces for the foam metal No. 1 to No. 3 (10, 20, 40 PIP) of the three types of pore densities described above, fins of the three widths H of 5 mm, 8 mm, and 12 mm were prepared. Thus, a total of nine types of samples were prepared. The height L of the foam metal in the vertical direction was 89 mm, the depth D was 13 mm, and the surface area per unit volume was β.
- In this case, the pressure loss ΔP under a non-load condition in which there is no flow of cold water is the same as in the first test example described above (refer to the graph of Fig. 6).
- However, when cold water flows under a non-load condition, two states must be considered, one being when the fin surface is dry and the other being when the fin surface is wet.
- 1-1) Pressure Loss In Dry State (State In Which Fin Surface Is Dry)
- First, the graph of Fig. 12 shows the pressure loss ΔP/D(Pa/m) with respect to the front surface wind velocity Vf(m/s) in the dry state. In the second test example, since the difference in both dry and wet states are considered, to increase the measurement accuracy of the pressure loss, the pressure loss is calculated by including the air flow direction length D (m) of at the fins (ΔP/D). In the following description, this is simply referred to as ΔP.
- In this case as well, the pressure loss ΔP increases as the pore density PIP increases, that is, as the pore size dpore decreases, and the pressure loss ΔP increases as the fin width H decreases. This is because the surface area (including wall surface) per unit volume increases as the pore size dpore decreases and the fin width H decreases. It is apparent that foam metal has a high pressure loss ΔP compared to the louver fin (width H=7.9 mm, D=13.6 mm, fin pitch=1.5 mm) of the comparative example.
- 1-2) Pressure Loss In Wet State (State In Which Fin Surface Is Wet)
- Fig. 13 shows the pressure loss ΔP (Pa) with respect to the front surface wind velocity Vf(m/s) in the wet state in when the temperature of the water serving as the refrigerant is 5°C, and Fig. 14 shows the same state when the temperature of the water serving as the refrigerant is 10°C.
- The influence of the pore density PPI and the fin width H is substantially the same as the tendency in the dry state. However, the value of the pressure loss ΔP greatly increases compared to that in the dry state (see Fig. 12). The pressure loss ΔP greatly increases because the condensed water accumulated on the fin surface becomes a ventilation resistance, and thus can be predicted that water drainage becomes an essential factor in comparison to the dry state. The ratio of the pressure loss ΔP between the wet state and the dry state is shown in Fig. 15 (for a water temperature of 5°C) and Fig. 16 (for a water temperature of 10°C). In the case of water temperature of 5°C in Fig. 15, the ratio of the pressure loss gradually increases as a whole when the air flow rate increases but the pressure loss decreases when the air flow rate increases in the cases of 8 mm, 12 mm at 10°C (Fig. 16).
- This is because the fin efficiency decreases when the supply water temperature is high, the fin width is large, and the wind velocity is fast. Thus, the temperature of the fin surface distant from the wall surface becomes higher than the dew point temperature of the air, the moisture in air cannot condense, and the increasing rate of the pressure loss ΔP decreases. That is, moisture condensation occurs in only parts of the fin.
- In the case of Fig. 15 (water temperature of 5°C), the ratio ΔPwet/ΔPdry of the pressure loss ΔP in a dry state and in a wet state increases as the pore density PPI increases. The ratio is greater for the foam metal fin than for the louver fin. That is, the water drainage is poorer in the foam metal fin than the louver fin.
- However, the foam metal fin used in the present test is in an experimental level in which the fin surface is not processed. Thus, the problem of water drainage is sufficiently improved, for example, by applying a hydrophilic agent.
- In this case as well, two states of the fin surface, a dry state and a wet state, must be considered.
- Fig. 17 shows the relationship of the heat transfer coefficient hdry with respect to the front surface wind velocity Vf(m/s) in a dry state. The heat transfer coefficient hdry increases as the pore density PIP decreases, and the heat transfer coefficient dry decreases as the fin width H increases. The heat transfer coefficient hdry for a foam metal fin of 10 PPI, H=5 mm is about the same as the louver fin but becomes inferior to the louver fin if the pore density PPI increases.
- However, since the foam metal fin has a large surface area β per unit volume, to evaluate the heat transfer performance described above, evaluation is performed with the heat transfer performance hdryβ per unit volume. Fig. 18 shows the relationship between hdryβ and the front surface wind velocity Vf(m/s) in the dry state. It is apparent that the heat transfer performance increases as the pore density PIP increases and the fin width H decreases. Fig. 19 shows the ratio of the heat transfer performance of the foam metal fin with respect to the louver fin. In the foam metal fin of
pore density 40 PIP and fin width H=5 mm, the heat transfer performance is greater than the louver fin by 1.5 times. This suggests the possibility of effective miniaturization of the heat exchanger. - Figs. 20 and 21 show the relationship of the heat transfer coefficient hwet in the wet state with respect to the front surface wind velocity Vf(m/s) in the wet state when the temperature of cold water is 5°C and 10°C. When comparing Figs. 20 and Fig. 21, it can be considered that the influence of the change in temperature (5°C to 10°C) of water is not large. Compared with the heat transfer coefficient hdry (see Fig. 17) in the dry state, hwet is slightly smaller than hdry. In addition to the heat transfer of sensible heat in the wet state, this is because the fin efficiency decreases in a dry state due to the heat transfer of latent heat resulting from condensation of the moisture in air.
- Fig. 22 (
water temperature 5°C) and Fig. 23 (water temperature 10°C) show the heat transfer performance hwetβ per unit volume in the wet state. As can be seen, the increase in hwetβ of the foam metal fin with respect to the louver fin is greater than in the dry state and is about 1.8 times greater for the pore density of 40 PPI and the fin width of H=5 mm. That is, the heat transfer promoting effect in the wet state is greater than in the dry state. - Fig. 24 (
water temperature 5°C) and Fig. 25 (water temperature 10°C) show the relationship between substance transfer coefficient hmass and air flow rate Vf(m/s). In the same manner as the heat transfer coefficient hwet of Fig. 20 (water temperature 5°C), the substance transfer coefficient hmass of Fig. 24 (water temperature 5°C) increases as the entire air flow rate and fin width H increase and as the pore density PIP decreases. However, when H=12 mm and PPI=20, 40, it does not increase even if the wind velocity rises from around the speed of 1.0m/s. Similar to the explanation for the pressure loss, this is because the temperature of part of the fin surface becomes higher than or equal to the dew point temperature and the moisture does not condense. This tendency is more significant for the water temperature of 10°C in Fig. 25. This is also seen in the substance transfer coefficient hmassβ(kg/m3s)-Vf(m/s) per unit volume (not shown). - As described above, the foam metal fin of the present embodiment has higher pressure loss and higher heat transfer coefficient per volume compared to the existing louver fin. However, in order to be configured as an air conditioner heat exchanger, the pressure loss and the heat transfer coefficient must be comprehensively analyzed. The pump power necessary for unit volume is expressed by the following equation.
- V is the volume, and Ac is the flow cross-sectional area.
- Figs. 26 and 27 show the heat transfer coefficients hdryβ, hwetβ per unit volume for the dry state and the wet state in relation to the pump power EBA necessary per unit volume. In the dry state of Fig. 26, the highest heat transfer coefficient is obtained with 40 PIP and H=5 mm, and the heat transfer performance improvement effect of about 24% with respect to the louver fin is expected. In the wet state of Fig. 27, for 40 PPI, H=5 mm, 8 mm, and 12 mm, the heat transfer performance improvement effect is higher by about 28%.
- As shown in Fig. 28, this is also seen in the mass heat transfer coefficient hmassβ(kg/m3s)-Eβw/m3) per unit volume.
- Fig. 29 shows the structure for an air-heat exchanger according to a second embodiment of the present invention.
- The present embodiment relates to a serpentine heat exchanger in which a flat
heat transfer tube 21 is bent into a serpentine shape as a single continuous structure. - The high heat transfer performance is also achieved in the heat exchanger of such configuration in the same manner as the heat exchanger of the first embodiment described above.
- Fig. 30 shows the structure for a heat exchanger according to a third embodiment of the present invention.
- In this embodiment, a plurality of plate shaped
heat transfer tubes 31 extending in the horizontal direction are connected in a stack form by left and right connectingmembers 22, and theheat transfer fin 13 made of foam metal is arranged between the connectingmembers 22 of each layer. Eachheat transfer tube 31 includes refrigerant inlet and outlet holes 23. Theholes 23 connected through the connectingmembers 22 form a refrigerant passage. The connectingmembers 22 are used for a structure similar to that of the first embodiment in a stacked plate type air-heat exchanger. - In such a configuration, high heat transfer performance is realized in the same manner as in the first embodiment.
- The heat transfer fin of the present invention is not limited to the structure of the heat exchanger in each embodiment and may obviously be applied to a heat transfer fin for performing heat exchange with air, such as a cross fin type or the like.
Claims (5)
- A heat transfer fin for a heat exchanger, wherein a heat transfer fin (13) is arranged on a heat transfer tube (1) through which fluid flows for exchanging heat with air, the heat transfer fin (13) contacts air to exchange heat, and the heat transfer fin (13) is made of foam metal having a pore density of 20 PIP or greater.
- The heat transfer fin for a heat exchanger according to claim 1, wherein a heat transfer tube (1) through which fluid flows for exchanging heat with air is provided as in a plurality, and the plurality of heat transfer tubes (1) are set at an interval H of 12 mm or less.
- The heat transfer fin for a heat exchanger according to claim 1 or 2, wherein the heat exchanger is a stacked type heat exchanger.
- The heat transfer fin for a heat exchanger according to any one of claims 1 to 3, wherein the pore density is 20 PIP or greater and 60 PIP or less.
- The heat transfer fin for a heat exchanger according to claim 2, wherein an interval H between the plurality of heat transfer tubes (1) is set at 4 mm or greater and 12 mm or less.
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JP2005053087A JP2005326136A (en) | 2004-04-16 | 2005-02-28 | Heat transfer fin for air heat exchanger |
PCT/JP2005/007383 WO2005100898A1 (en) | 2004-04-16 | 2005-04-18 | Heat transfer fin for heat exchanger |
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US (1) | US20080296008A1 (en) |
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Cited By (3)
Publication number | Priority date | Publication date | Assignee | Title |
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WO2014029465A1 (en) * | 2012-08-18 | 2014-02-27 | Audi Ag | Heat exchanger |
CN104896968A (en) * | 2015-06-16 | 2015-09-09 | 中国石油大学(华东) | Metal foam finned tube heat exchanger |
CN108317774A (en) * | 2018-01-31 | 2018-07-24 | 天津商业大学 | A kind of CO based on foam metal2Cooling evaporator |
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JPS52134153A (en) * | 1976-05-04 | 1977-11-10 | Nippon Denso Co Ltd | Layer-built heat exchanger |
JPS56155391A (en) * | 1980-04-30 | 1981-12-01 | Nippon Denso Co Ltd | Corrugated fin type heat exchanger |
JPS60294A (en) * | 1983-06-16 | 1985-01-05 | Matsushita Seiko Co Ltd | Heat exchanger utilizing foamed metal |
KR19990085965A (en) * | 1998-05-23 | 1999-12-15 | 박호군 | Porous Fin Plate Heat Exchanger |
NL1016713C2 (en) * | 2000-11-27 | 2002-05-29 | Stork Screens Bv | Heat exchanger and such a heat exchanger comprising thermo-acoustic conversion device. |
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- 2005-02-28 JP JP2005053087A patent/JP2005326136A/en active Pending
- 2005-04-18 WO PCT/JP2005/007383 patent/WO2005100898A1/en not_active Application Discontinuation
- 2005-04-18 EP EP05730153A patent/EP1739377A4/en not_active Withdrawn
- 2005-04-18 US US11/578,046 patent/US20080296008A1/en not_active Abandoned
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See also references of WO2005100898A1 * |
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WO2014029465A1 (en) * | 2012-08-18 | 2014-02-27 | Audi Ag | Heat exchanger |
US9664459B2 (en) | 2012-08-18 | 2017-05-30 | Audi Ag | Heat exchanger with a porous metal structure having manifolds and tubes |
CN104896968A (en) * | 2015-06-16 | 2015-09-09 | 中国石油大学(华东) | Metal foam finned tube heat exchanger |
CN108317774A (en) * | 2018-01-31 | 2018-07-24 | 天津商业大学 | A kind of CO based on foam metal2Cooling evaporator |
Also Published As
Publication number | Publication date |
---|---|
WO2005100898A1 (en) | 2005-10-27 |
EP1739377A4 (en) | 2009-12-02 |
JP2005326136A (en) | 2005-11-24 |
US20080296008A1 (en) | 2008-12-04 |
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