CA1302395C - Heat exchanger tube having minute internal fins - Google Patents
Heat exchanger tube having minute internal finsInfo
- Publication number
- CA1302395C CA1302395C CA000564999A CA564999A CA1302395C CA 1302395 C CA1302395 C CA 1302395C CA 000564999 A CA000564999 A CA 000564999A CA 564999 A CA564999 A CA 564999A CA 1302395 C CA1302395 C CA 1302395C
- Authority
- CA
- Canada
- Prior art keywords
- heat exchanger
- tube
- inches
- recited
- refrigerant
- 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.)
- Expired - Lifetime
Links
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/40—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only inside the tubular element
Landscapes
- Physics & Mathematics (AREA)
- Engineering & Computer Science (AREA)
- Geometry (AREA)
- Thermal Sciences (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
Abstract
Title HEAT EXCHANGER TUBE HAVING
MINUTE INTERNAL FINS
Inventor Carl Bergt Abstract A heat exchanger includes heat exchanger tubes with minute internal fins disposed at a minimal helix angle with respect to a longitudinal axis of the tube. When conveying a refrigerant, the minimal helix angle provides surprisingly better heat transfer rates than other tubes having greater helix angles.
MINUTE INTERNAL FINS
Inventor Carl Bergt Abstract A heat exchanger includes heat exchanger tubes with minute internal fins disposed at a minimal helix angle with respect to a longitudinal axis of the tube. When conveying a refrigerant, the minimal helix angle provides surprisingly better heat transfer rates than other tubes having greater helix angles.
Description
13023g~
Title HEAT EXCHANGER TUBE HAVING
MINUTE INTERNAL FINS
Technical Field ~ The subject invention generally pertains to heat exchanger tubes, and more specifically to tubes having internal fins.
Background of the Invention Refrigeration systems, such as air conditioners and heat pumps, typically include a compressor, one heat exchanger functioning as an evaporator, an expansion device, and a second heat exchanger functioning as a condenser, all of which are connected in series to circulate a refrigerant in a closed loop.
The two heat exchangers each include at least one heat exshanger tube for transferring heat either to or from the refrigerant.
The rate of heat transfer is enhanced by providing the heat exchanger tubes with internal fins.
In designing tubes having internal fins, many interrelated factors need to be considered such as fin height, fin spacing, helix angle, heat flux, flow rate, and various properties of the fluid being conveyed through the tube. Varying each of these factors provides an infinite combination of factors, making it difficult and costly to extensively study all the possibilities. As a result, some apparently conflicting conclusions have been drawn based on different lab tests.
Title HEAT EXCHANGER TUBE HAVING
MINUTE INTERNAL FINS
Technical Field ~ The subject invention generally pertains to heat exchanger tubes, and more specifically to tubes having internal fins.
Background of the Invention Refrigeration systems, such as air conditioners and heat pumps, typically include a compressor, one heat exchanger functioning as an evaporator, an expansion device, and a second heat exchanger functioning as a condenser, all of which are connected in series to circulate a refrigerant in a closed loop.
The two heat exchangers each include at least one heat exshanger tube for transferring heat either to or from the refrigerant.
The rate of heat transfer is enhanced by providing the heat exchanger tubes with internal fins.
In designing tubes having internal fins, many interrelated factors need to be considered such as fin height, fin spacing, helix angle, heat flux, flow rate, and various properties of the fluid being conveyed through the tube. Varying each of these factors provides an infinite combination of factors, making it difficult and costly to extensively study all the possibilities. As a result, some apparently conflicting conclusions have been drawn based on different lab tests.
U.S. Patent 4,044,797 suggests using an internal fin ` height of .02 to .2 mm (.0008 to .0079 inches) and a helix angle of 4- to 15-, while U.S. Patent 4,118,944 concludes that the most effective helix angles are greater than 20-. ~.S. Patent 4,545,428 suggests a fin height of .1 to .6 mm (.0039 to .0236 inches) and a helix angle of 16- to 35-. It should be noted, however, that none of the three patents recommends a helix angle of less than 4-.
In view of the above patents alone, it is difficult, if not impossible, to determine the optimum fin height and helix angle. Therefore, in developing the present invention, lab tests were performed to determine the best internal fin design. The unexpected results of the tests showed that straight internal fins, i.e., a zero degree helix angle, provide a surprisingly effective internal heat transfer surface. The results were so surprising that additional, more extensive tests were performed to check the validity of the earlier tests. The more extensive tests conflrmed the surprising results.
Since a heat exchanger tube having internal helical fins is more difficult and more costly to manufacture than a straight finned tube, it is an object of the invention to provide a heat exchanger tube with substantially straight internal fins that provides a greater heat transfer coefficient than that of a similar tube with helical fins.
It has now been discovered that it is possible to provide a heat exchanger tube with substantially straight internal fins having an optimum fin height.
It has also been found that it is possible to provide a tube with substantially straight internal fins that are circumferentially spaced apart at optimum intervals to avoid excessively thin, fragile fins and to avoid excess space between the fins.
In addition, it is possible to provide an effective heat exchanger tube for conveying refrigerant of varying quality.
- 1305~3~S
Furthermore, it is possible to provide a refrigerant flow rate that takes full advantage of substantially straight internal fins.
It is also possible to provide an internally finned S heat exchanger tube that provides heat transfer when conveying a fluorinated hydrocarbon refrigerant.
The invention will be more clearly understood with reference to the attached drawings and the description of the preferred embodiment which follows hereinbelow.
Summa~y of the Invention . ' , .
According to one aspect of the present invention there is provided a heat exchanger comprising a heat exchanger tube conveying a réfrigerant said tube having a plurality of longitudinally running internal fins disposed along an inner surface of said tube at a helix angle of less than 4 degrees with respect to a longitudinal axis of said tube, said fins having a fin height of .007 inches to .030 inches and being circumferentially spaced around said inner surface at intervals of .010 to .040 inches.
Brief ~escriotion of the Drawinos Figure 1 illustrates a half-section of a preferred embodiment of the invention.
Figure 2 illustrate a half-section of a heat exchanger tube having a helix angle that is greater than zero.
Figure 3 shows how a heat transfer multiplier varies as 30 a function of fin height and helix angle.
'i~
Figure 4 shows how a heat transfer multiplier varies as a function of helix angle for a given fin height.
Figure 5 shows how a heat transfer multiplier varies as a function of fin height for a given helix angle.
Figure 6 shows how a heat transfer multiplier varies as a function of mass flow rate for a preferred embodiment of the invention.
Descriptio of the Preferred Embodiment Heat exchanger tube 10, shown in Figure 1, is one embodiment of the subject invention. Tube 10 includes generally straight longitudinal running internal fins 12, i.e., its helix angle ~ is zero degrees with respect to a longitudinal axis 14 of the tube. Fins 12 have a height "h" of .0155 inches and are distributed circumferentially around the inner surface of tube 10 at .017 inch intervals 16. Interval 16 is defined as the distance between a center point of one fin to the center point of an adjacent fin.
Under equivalent test conditions, tube 10 proves to be superior to a variety of other internally finned tubes having various fin heights and helix angles, such as tube 18 shown in Figure 2. The variety of other tubes that have been tested and compared to tube 10 are generally suitable in refrigeration systems. This means that the tubes have a nominal outside diameter of approximately one inch or less and the internal surface of the tubes have a high concentration (intervals of .010 to .040 inches) of relatively minute fins (fin height below .035 inches) to enhance heat transfer while minimizing flow resistance, or pressure drop through the tube. Minimizing pressure drop is especially important when conveying refrigerant, because the refrigerant's temperature and vapor/liquid quality (ratio of vapor to liquid mass) changes significantly with pressure which, in turn, affects the rate of heat transfer across the tube. The fins are closely distributed around the tube's inner circumference at .013 to .033 inch intervals 16 to maximize fin surface area while avoiding the use of excessively thin, fragile fins.
Each tube that was tested had a nominal outer diameter of 3/8 inches and was tested by conveying a refrigerant through the interior of the tube. The specific refrigerant that was used in the tests was "FREON", which is a trademark for a fluorinated hydrocarbon. More specifically, the refrigerant was "FREON R-22"
which is a trademark for difluoromonochloromethane. The temperature of the external surface of the tube was controlled to provide a constant heat flux of 5,000 Btu/hr-ft2. The inlet and outlet temperature of the refrigerant was controllably changed to test each tube as i~ conveyed refrigerant of different vapor/liquid qualities. The quality was varied at five incremental values ranging from .15 to .85, and the results that are shown in Figures 3 through 6 are based on an average quality of .6 with the tubes functioning as an evaporator. Figures 3, 4 and 5 are based on a refrigerant mass flow rate of 200 lbs/hr.
The tests provided the heat transfer coefficient (Btu/hr-ft2-F) of various tubes having internal fins. The coefficients were compared to those of smooth tubes having no internal fins. From the comparison, a dimensionless improvement factor, referred to hereinbelow as a heat transfer multiplier or simply a multiplier, was determined by dividing the heat transfer coefficient of the finned tube by the coefficient a comparable smooth tube.
9~
The results of the tests are summarized in Figure 3.
Eight different internally finned tubes are represented by points "A-H" which have been plotted according to fin height h and helix angle e. As indicated at point B, each point A-H is accompanied by its heat transfer multiplier 20 as determined by actual tests.
Below each multiplier 20, in parentheses, is a calculated - multiplier "Z" based on an empirically derived equation 22 having a 96% coorelation with the measured multipliers 20. Equation 22 defines multiplier Z as a function of fin height h and helix angle e, with h being expressed in mils (1 mil = .001 inches) and e expressed in degrees with respect to a longitudinal axis of the tube.
Regions 24 and 26 of Figure 3 represent combinations of fin height h and helix angle e that provide a multiplier greater than two based on equation 22. In other words, the heat transfer rate of a smooth tube could be expected to double if it were modified to include internal fins having a combination height h and helix angle e that lies within regions 24 or 26. It is worth noting that U.S. Patents 4,545,428 and 4,118,944 have pointed out the importance of regions 28 and 30 respectively which generally coincides with region 26; however, the importance of the relatively narrow region 24 has not been appreciated. Region 24 identifies a specific set of tubes that represent the preferred embodiment of the invention.
Figure 4 illustrates how varying the helix angle affects heat transfer for given a fin height of approximately .008 inches (8 mils). Tubes represented by points C, D, G and H
have a fin height of .008, .0075, .0085, and .008 inches respectively. A V-shape curve 32 represents multiplier Z as a function of helix angle e based on equation 22 for a constant fin height of .008 inches. Curve 32 illustrates how the multiplier increases as the helix angle increases or decreases from a low point 34 of 12 degrees.
i;~O;~39~;
The effect of fin height for a given helix angle is illustrated in Figure 5. Curve 36 represents multiplier Z as a function of fin height based on equation 22 with a constant helix angle of 6. Internally finned tubes represented by points B, C, D and E have a helix angle of 7, 7, 5 and 6 respectively.
Point 38 represents a comparable smooth tube having no internal - fins. Figure 5 shows that for a helix angle of 6, optimum heat transfer is obtained at a fin height of at least .007 inches.
However, its best to limit the height to less than .030 inches to facilitate fin forming. This generally limits the fin height to no more than a nominal thickness 40 (Figure 2) of readily available tubing having a nominal pre-finned wall thickness ranging from .012 to .033 inches. Moreover, when using 3/ô inch O.D. tubing having a nominal wall thickness of .027 inches, an internal fin having a fin height of .020 inches, for example, will only project 7~ across the internal diameter of the tube to provide minimal flow restriction.
The tests also show that the heat transfer multiplier Z
is highest at lower flow rates. This is illustrated by curve 42, shown in Figure 6, which is based on actual data points 44 and 46 and empirically derived points 48. Curve 42 clearly shows that a mass flow rate below 400 lbs/hr is the optimum flow rate for tube 10 which represents one embodiment of the invention. Line 43 represents a smooth tube which, by definition, has a multiplier equal to one. Comparing curve 42 to line 43 shows that tube 10 is superior to a comparable smooth tube at flow rates below 650 lbs/hr. For a tube 10 having a nominal 3/8 inch ouside diameter and a .321 inch inside diameter, 650 lbs/hr provides a mass flow per cross-sectional area (mass flux) of 8,032 lbs/hr-in2.
Tube 10 was also tested in a condensing mode. The tests were similar to the tests performed in the evaporating mode, except the refrigerant being conveyed through the tube was cooled instead of heated. The tests showed that multiplier 20 of tube 10 only decreased from 2.12 in the evaporating mode to 2.02 in the condensing mode. The small 4.7% decrease indicates that tube 10 is suitable to function as both an evaporator and a condenser.
Although the invention is described wih respect to a preferred embodiment, modifications thereto will be apparent to those skilled in the art. Therefore, the scope of the invention is to be determined by reference to the claims which follow.
I claim:
In view of the above patents alone, it is difficult, if not impossible, to determine the optimum fin height and helix angle. Therefore, in developing the present invention, lab tests were performed to determine the best internal fin design. The unexpected results of the tests showed that straight internal fins, i.e., a zero degree helix angle, provide a surprisingly effective internal heat transfer surface. The results were so surprising that additional, more extensive tests were performed to check the validity of the earlier tests. The more extensive tests conflrmed the surprising results.
Since a heat exchanger tube having internal helical fins is more difficult and more costly to manufacture than a straight finned tube, it is an object of the invention to provide a heat exchanger tube with substantially straight internal fins that provides a greater heat transfer coefficient than that of a similar tube with helical fins.
It has now been discovered that it is possible to provide a heat exchanger tube with substantially straight internal fins having an optimum fin height.
It has also been found that it is possible to provide a tube with substantially straight internal fins that are circumferentially spaced apart at optimum intervals to avoid excessively thin, fragile fins and to avoid excess space between the fins.
In addition, it is possible to provide an effective heat exchanger tube for conveying refrigerant of varying quality.
- 1305~3~S
Furthermore, it is possible to provide a refrigerant flow rate that takes full advantage of substantially straight internal fins.
It is also possible to provide an internally finned S heat exchanger tube that provides heat transfer when conveying a fluorinated hydrocarbon refrigerant.
The invention will be more clearly understood with reference to the attached drawings and the description of the preferred embodiment which follows hereinbelow.
Summa~y of the Invention . ' , .
According to one aspect of the present invention there is provided a heat exchanger comprising a heat exchanger tube conveying a réfrigerant said tube having a plurality of longitudinally running internal fins disposed along an inner surface of said tube at a helix angle of less than 4 degrees with respect to a longitudinal axis of said tube, said fins having a fin height of .007 inches to .030 inches and being circumferentially spaced around said inner surface at intervals of .010 to .040 inches.
Brief ~escriotion of the Drawinos Figure 1 illustrates a half-section of a preferred embodiment of the invention.
Figure 2 illustrate a half-section of a heat exchanger tube having a helix angle that is greater than zero.
Figure 3 shows how a heat transfer multiplier varies as 30 a function of fin height and helix angle.
'i~
Figure 4 shows how a heat transfer multiplier varies as a function of helix angle for a given fin height.
Figure 5 shows how a heat transfer multiplier varies as a function of fin height for a given helix angle.
Figure 6 shows how a heat transfer multiplier varies as a function of mass flow rate for a preferred embodiment of the invention.
Descriptio of the Preferred Embodiment Heat exchanger tube 10, shown in Figure 1, is one embodiment of the subject invention. Tube 10 includes generally straight longitudinal running internal fins 12, i.e., its helix angle ~ is zero degrees with respect to a longitudinal axis 14 of the tube. Fins 12 have a height "h" of .0155 inches and are distributed circumferentially around the inner surface of tube 10 at .017 inch intervals 16. Interval 16 is defined as the distance between a center point of one fin to the center point of an adjacent fin.
Under equivalent test conditions, tube 10 proves to be superior to a variety of other internally finned tubes having various fin heights and helix angles, such as tube 18 shown in Figure 2. The variety of other tubes that have been tested and compared to tube 10 are generally suitable in refrigeration systems. This means that the tubes have a nominal outside diameter of approximately one inch or less and the internal surface of the tubes have a high concentration (intervals of .010 to .040 inches) of relatively minute fins (fin height below .035 inches) to enhance heat transfer while minimizing flow resistance, or pressure drop through the tube. Minimizing pressure drop is especially important when conveying refrigerant, because the refrigerant's temperature and vapor/liquid quality (ratio of vapor to liquid mass) changes significantly with pressure which, in turn, affects the rate of heat transfer across the tube. The fins are closely distributed around the tube's inner circumference at .013 to .033 inch intervals 16 to maximize fin surface area while avoiding the use of excessively thin, fragile fins.
Each tube that was tested had a nominal outer diameter of 3/8 inches and was tested by conveying a refrigerant through the interior of the tube. The specific refrigerant that was used in the tests was "FREON", which is a trademark for a fluorinated hydrocarbon. More specifically, the refrigerant was "FREON R-22"
which is a trademark for difluoromonochloromethane. The temperature of the external surface of the tube was controlled to provide a constant heat flux of 5,000 Btu/hr-ft2. The inlet and outlet temperature of the refrigerant was controllably changed to test each tube as i~ conveyed refrigerant of different vapor/liquid qualities. The quality was varied at five incremental values ranging from .15 to .85, and the results that are shown in Figures 3 through 6 are based on an average quality of .6 with the tubes functioning as an evaporator. Figures 3, 4 and 5 are based on a refrigerant mass flow rate of 200 lbs/hr.
The tests provided the heat transfer coefficient (Btu/hr-ft2-F) of various tubes having internal fins. The coefficients were compared to those of smooth tubes having no internal fins. From the comparison, a dimensionless improvement factor, referred to hereinbelow as a heat transfer multiplier or simply a multiplier, was determined by dividing the heat transfer coefficient of the finned tube by the coefficient a comparable smooth tube.
9~
The results of the tests are summarized in Figure 3.
Eight different internally finned tubes are represented by points "A-H" which have been plotted according to fin height h and helix angle e. As indicated at point B, each point A-H is accompanied by its heat transfer multiplier 20 as determined by actual tests.
Below each multiplier 20, in parentheses, is a calculated - multiplier "Z" based on an empirically derived equation 22 having a 96% coorelation with the measured multipliers 20. Equation 22 defines multiplier Z as a function of fin height h and helix angle e, with h being expressed in mils (1 mil = .001 inches) and e expressed in degrees with respect to a longitudinal axis of the tube.
Regions 24 and 26 of Figure 3 represent combinations of fin height h and helix angle e that provide a multiplier greater than two based on equation 22. In other words, the heat transfer rate of a smooth tube could be expected to double if it were modified to include internal fins having a combination height h and helix angle e that lies within regions 24 or 26. It is worth noting that U.S. Patents 4,545,428 and 4,118,944 have pointed out the importance of regions 28 and 30 respectively which generally coincides with region 26; however, the importance of the relatively narrow region 24 has not been appreciated. Region 24 identifies a specific set of tubes that represent the preferred embodiment of the invention.
Figure 4 illustrates how varying the helix angle affects heat transfer for given a fin height of approximately .008 inches (8 mils). Tubes represented by points C, D, G and H
have a fin height of .008, .0075, .0085, and .008 inches respectively. A V-shape curve 32 represents multiplier Z as a function of helix angle e based on equation 22 for a constant fin height of .008 inches. Curve 32 illustrates how the multiplier increases as the helix angle increases or decreases from a low point 34 of 12 degrees.
i;~O;~39~;
The effect of fin height for a given helix angle is illustrated in Figure 5. Curve 36 represents multiplier Z as a function of fin height based on equation 22 with a constant helix angle of 6. Internally finned tubes represented by points B, C, D and E have a helix angle of 7, 7, 5 and 6 respectively.
Point 38 represents a comparable smooth tube having no internal - fins. Figure 5 shows that for a helix angle of 6, optimum heat transfer is obtained at a fin height of at least .007 inches.
However, its best to limit the height to less than .030 inches to facilitate fin forming. This generally limits the fin height to no more than a nominal thickness 40 (Figure 2) of readily available tubing having a nominal pre-finned wall thickness ranging from .012 to .033 inches. Moreover, when using 3/ô inch O.D. tubing having a nominal wall thickness of .027 inches, an internal fin having a fin height of .020 inches, for example, will only project 7~ across the internal diameter of the tube to provide minimal flow restriction.
The tests also show that the heat transfer multiplier Z
is highest at lower flow rates. This is illustrated by curve 42, shown in Figure 6, which is based on actual data points 44 and 46 and empirically derived points 48. Curve 42 clearly shows that a mass flow rate below 400 lbs/hr is the optimum flow rate for tube 10 which represents one embodiment of the invention. Line 43 represents a smooth tube which, by definition, has a multiplier equal to one. Comparing curve 42 to line 43 shows that tube 10 is superior to a comparable smooth tube at flow rates below 650 lbs/hr. For a tube 10 having a nominal 3/8 inch ouside diameter and a .321 inch inside diameter, 650 lbs/hr provides a mass flow per cross-sectional area (mass flux) of 8,032 lbs/hr-in2.
Tube 10 was also tested in a condensing mode. The tests were similar to the tests performed in the evaporating mode, except the refrigerant being conveyed through the tube was cooled instead of heated. The tests showed that multiplier 20 of tube 10 only decreased from 2.12 in the evaporating mode to 2.02 in the condensing mode. The small 4.7% decrease indicates that tube 10 is suitable to function as both an evaporator and a condenser.
Although the invention is described wih respect to a preferred embodiment, modifications thereto will be apparent to those skilled in the art. Therefore, the scope of the invention is to be determined by reference to the claims which follow.
I claim:
Claims (12)
1. A heat exchanger comprising a heat exchanger tube conveying a refrigerant said tube having a plurality of longitudinally running internal fins disposed along an inner surface of said tube at a helix angle of less than 4 degrees with respect to a longitudinal axis of said tube, said fins having a fin height of .007 inches to .030 inches and being circumferentially spaced around said inner surface at intervals of .010 to .040 inches.
2. The heat exchanger as recited in claim 1, wherein said longitudinally running fins are generally straight running fins, whereby said helix angle is substantially zero degrees.
3. The heat exchanger as recited in claim 1, wherein said fins are circumferentially spaced around said inner surface at intervals of 0.13 to .033 inches.
4. The heat exchanger as recited in claim 1, wherein said fin height is .010 to .020 inches.
5. The heat exchanger as recited in claim 1, wherein said heat exchanger functions as an evaporator and is connected to a refrigeration system having a second heat exchanger, said second heat exchanger being defined by claim 1 and functioning as a condenser.
6. The heat exchanger as recited in claim 1, wherein said refrigerant is a fluorinated hydrocarbon.
7. The heat exchanger as recited in claim 6, wherein said refrigerant is difluoromonochloromethane.
8. The heat exchanger as recited in claim 1, wherein said refrigerant has a vapor/liquid quality that varies as said refrigerant is conveyed through said tube.
9. The heat exchanger as recited in claim 8, wherein said vapor/liquid quality changes between two values within said tube with one value being greater than .6 and the other value being less than .6.
10. A heat exchanger comprising a heat exchanger tube conveying a fluorinated hydrocarbon refrigerant whose vapor/liquid quality changes between two values within said tube with one value being greater than .6 and the other value being less than .6, said tube having a plurality of longitudinally running internal fins disposed along an inner surface of said tube at a helix angle substantially equal to zero degrees with respect to a longitudinal axis of said tube, said fins being circumferentially spaced around said inner surface at intervals of .013 to .033 inches and having a fin height of .010 to .020 inches.
11. The heat exchanger as recited in claim 10, wherein said heat exchanger functions as an evaporator and is connected to a refrigeration system having a second heat exchanger, said second heat exchanger being defined by claim 6 and functioning as a condenser.
12. The heat exchanger as recited in claim 10, wherein said refrigerant is difluoromonochloromethane.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| US12655087A | 1987-11-30 | 1987-11-30 | |
| US126,550 | 1987-11-30 |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1302395C true CA1302395C (en) | 1992-06-02 |
Family
ID=22425434
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000564999A Expired - Lifetime CA1302395C (en) | 1987-11-30 | 1988-04-25 | Heat exchanger tube having minute internal fins |
Country Status (5)
| Country | Link |
|---|---|
| JP (1) | JPH01150797A (en) |
| CA (1) | CA1302395C (en) |
| DE (1) | DE3815095A1 (en) |
| FR (1) | FR2623893B1 (en) |
| GB (1) | GB2212899B (en) |
Families Citing this family (7)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US4938282A (en) * | 1988-09-15 | 1990-07-03 | Zohler Steven R | High performance heat transfer tube for heat exchanger |
| AT398629B (en) * | 1990-10-29 | 1995-01-25 | Vaillant Gmbh | WATER HEATER |
| US5070937A (en) * | 1991-02-21 | 1991-12-10 | American Standard Inc. | Internally enhanced heat transfer tube |
| MX9305803A (en) * | 1992-10-02 | 1994-06-30 | Carrier Corp | HEAT TRANSFER TUBE WITH INTERNAL RIBS. |
| US6164370A (en) * | 1993-07-16 | 2000-12-26 | Olin Corporation | Enhanced heat exchange tube |
| JP2009024899A (en) * | 2007-07-17 | 2009-02-05 | Showa Denko Kk | Evaporator |
| KR102048356B1 (en) * | 2013-03-08 | 2019-11-25 | 엘지전자 주식회사 | Refrigerant pipe, and fin type heat exchanger and air conditioner comprising the same |
Family Cites Families (8)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| IL36327A (en) * | 1970-03-10 | 1973-11-28 | Electrolux Ab | Absorption refrigerating apparatus |
| NL7213941A (en) * | 1972-10-14 | 1974-04-16 | ||
| US4044797A (en) * | 1974-11-25 | 1977-08-30 | Hitachi, Ltd. | Heat transfer pipe |
| JPS5474549A (en) * | 1977-11-25 | 1979-06-14 | Toshiba Corp | Heat conducting tube |
| JPS5726394A (en) * | 1980-07-22 | 1982-02-12 | Hitachi Cable Ltd | Heat conduction pipe with grooves in internal surface |
| JPS60142195A (en) * | 1983-12-28 | 1985-07-27 | Hitachi Cable Ltd | Heat transfer tube equipped with groove on internal surface thereof |
| JPH0769117B2 (en) * | 1985-10-23 | 1995-07-26 | 古河電気工業株式会社 | Small diameter heat transfer tube and its manufacturing method |
| JPS62142995A (en) * | 1985-12-17 | 1987-06-26 | Hitachi Cable Ltd | Heat transfer pipe with inner surface spiral groove |
-
1988
- 1988-04-14 GB GB8808863A patent/GB2212899B/en not_active Expired - Lifetime
- 1988-04-25 CA CA000564999A patent/CA1302395C/en not_active Expired - Lifetime
- 1988-05-04 DE DE19883815095 patent/DE3815095A1/en active Granted
- 1988-05-06 FR FR8806135A patent/FR2623893B1/en not_active Expired - Lifetime
- 1988-06-24 JP JP15507288A patent/JPH01150797A/en active Pending
Also Published As
| Publication number | Publication date |
|---|---|
| GB8808863D0 (en) | 1988-05-18 |
| DE3815095A1 (en) | 1989-06-08 |
| DE3815095C2 (en) | 1990-12-06 |
| GB2212899A (en) | 1989-08-02 |
| FR2623893B1 (en) | 1991-01-25 |
| JPH01150797A (en) | 1989-06-13 |
| FR2623893A1 (en) | 1989-06-02 |
| GB2212899B (en) | 1991-11-20 |
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| JPH0979779A (en) | Heat transfer tube with internal surface groove and heat exchanger |
Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| MKLA | Lapsed |