Heat exchanger

Abstract: A heat exchanger for use in a refrigeration or air conditioning system for cooling the refrigerant comprises a pair of spaced headers (10,12), and at least one heat exchange tube (20) extending between the headers defining a plurality of hydraulically parallel refrigerant flow paths, each of said flow paths having a hydraulic diameter of up to about 0.07 inches, the hydraulic diameter being defined as the cross-sectional area of the flow path multiplied by four and divided by the wetted perimeter of the flow path.

Description: This invention relates to a heat exchanger for exchanging heat between the ambient and a refrigerant that may be in a liquid or vapour phase, comprising: a pair of spaced generally parallel headers, one of said headers having a refrigerant inlet, and one of said headers having a refrigerant outlet; and a heat exchanger tube extending between said headers and in fluid communication with each of said headers, said tube having a generally flat cross-section and defining a plurality of hydraulically parallel refrigerant flow paths between said headers, each of said refrigerant flow paths having a hydraulic diameter up to 1.778mm (0.07 inches). The invention is particularly applicable to a condenser for condensing a refrigerant using ambient air as a cooling medium. Many condensers employed in air conditioning or refrigeration systems at the present time utilize one or more serpentine conduits on the vapour side. Such condensers are shown, for example, in GB-A-2133525 and JP-U-5913877. The present invention is characterised in that said headers each have a series of openings with the openings in the series on one header being aligned with and facing the openings in the series on the other header; a tube row defined by a plurality of straight said tubes of generally flat cross section extends in parallel between said headers, the opposed ends of said tubes being disposed in corresponding aligned ones of said openings and in fluid communication with the interiors of said headers, at least some of said tubes being in hydraulic parallel to each other; webs within said tubes extend between and are joined to opposed side walls of the tubes at spaced intervals to (a) define a plurality of non-circular said flow paths within each tube, (b) absorb forces resulting from internal pressure within said heat exchanger and tending to expand said tubes, and (c) conduct heat between fluid in said flow paths and both said opposed side walls of said tubes; said webs and/or said flat side walls defined to least one concave zone at the intersection of converging surface segments in each of said flow paths extending along the length thereof; and serpentine fins incapable of supporting said tubes against substantial internal pressure extend between facing ones of said opposed side walls of adjacent tubes. The term "hydraulic diameter" as used herein means the cross sectional area of a flow path multiplied by four and divided by the wetted perimeter of the flow path. The heat exchanger according to the present invention has a relatively low frontal area on the air side that is blocked by tubes allowing an increase in the air side heat exchange surface area without increasing air side pressure drop and without increasing vapour and/or condensate side pressure drop. The invention will become apparent from the following specification taken in connection with the accompanying drawings. Fig 1 is an exploded, perspective view of an embodiment of a condenser made according to the invention; Fig 2 is a fragmentary, enlarged, cross-sectional view of a condenser tube that may be employed in the invention; Fig 3 is a graph of the predicted performance of condensers with the same face area, some made in a prior art design and others made according to the invention, plotting heat transfer against cavity (hydraulic) diameter; Fig 4 is a graph comparing the present invention with the prior art construction showing air flow through each versus (a) the rate of heat transfer, (b) the refrigerant flow rate, and (c) the refrigerant pressure drop; Fig 5 is a further graph comparing the prior art construction with a condenser made according to the invention on the basis of air velocity versus the heat transfer per pound of material employed in making up the core of each; and FIGURE 6 is a further graph comparing the prior art construction with an embodiments of the present invention by plotting air velocity versus pressure drop across the air side of the condenser. An exemplary embodiment of a condenser made according to the invention is illustrated in Figure 1 and is seen to include opposed, speced, generally parallel headers 10 and 12. Preferably, the headers 10 and 12 are made up from generally cylindrical tubing. On their facing sides, they are provided with a series of generally parallels slots or openings 14 for receipt of corresponding ends 16 and 18 of condenser tubes 20. Preferably, between the slots 14, in the area shown at 22, each of the headers 10 and 12 is provided with a somewhat spherical dome to improve resistance to pressure as explained more fully in US-A-4615385 the details of which are herein incorporated by reference. The header 10 has one end closed by a cap 24 brazed or welded thereto. Brazed or welded to the opposite end is a fitting 26 to which a tube 28 may be connected. The lower end of the header 12 is closed by a welded or brazed cap 30 similar to the cap 24 while its upper end is provided with a welded or brazed in place fitting 32. Depending upon the orientation of the condenser, one of the fittings 26 and 32 serves as a vapour inlet while the other serves as a condensate outlet. For the orientation shown in Figure 1, the fitting 26 will serve as a condensate outlet. A plurality of the tubes 20 extend between the headers 10 and 12 and are in fluid communication therewith. The tubes 20 are geometrically parallel to each other and hydraulically in parallel as well. Disposed between adjacent ones of the tubes 20 are serpentine fins 34 although plate fins could be used if desired. Upper and lower channels 36 and 38 extend between and are bonded by any suitable means to the headers 10 and 12 to provide rigidity to the system. As can be seen in Figure 1, each of the tubes 20 is a flattened tube and within its interior includes an undulating spacer 40. In cross-section, the spacer 40 appears as shown in Figure 2 and it will be seen that alternating crests are in contact along their entire length with the interior wall 42 or the tube 20 and bonded thereto by fillets 44 of solder or braze metal. As a consequence, a plurality of substantially discrete hydraulically parallel fluid flow paths 46,48,50,52,54,56,58 and 60 are provided within each of the tubes 20. That is to say, there is virtually no fluid communication from one of such flow paths to the adjacent flow paths on each side. This effectively means that each of the walls separating adjacent fluid flow paths 46,48,50,52,54,56,58 and 60 are bonded to both of sides of the flattened tube 20 along their entire length. As a consequence, there is no gap that would be filled by fluid with a lesser thermal conductivity. As a result, heat transfer from the fluid via the walls separating the various fluid flow paths identified previously to the exterior of the tube is maximized. In addition, it is believed that discrete flow paths of the size mentioned take advantage of desirable effects of heat transfer caused by surface tension phenomena. A second advantage resides in the fact the condensers such as that of the present invention are employed on the outlet side of a compressor and therefore are subjected to extremely high pressure. Conventionally, this high pressure will be applied to the interior of the tubes 20. Where so-called "plate" fins are utilized in lieu of the serpentine fins 34 illustrated in the drawings, the same tend to confine the tubes 20 and support them against the internal pressure employed in a condenser application. Conversely, serpentine fins such as those shown at 34 are incapable of supporting the tubes 20 against substantial internal pressure. According to the described embodiment of the invention, however, the desired support in a serpentine fin heat exchanger is accomplished by the fact that the spacer 40 and specifically the crests thereof are bonded along their entire lengths to the interior wall 42 of each tube 20. This bond results in various parts of the spacer 40 being placed in tension when the tube 20 is pressurized to absorb the force resulting from internal pressure within the tube 20 tending to expand the tube 20. A highly preferred means by which the tubes 20 with accompanying inserts 40 may be formed is disclosed in US-A-4688311 the details of which are also herein incorporated by reference. According to the invention, each of the flow paths 48,50,52,54,56 and 58 and to the extent possible depending upon the shape of the insert 40, the flow paths 46 and 60 as well, have a hydraulic diameter in the range of about 0.381 to 1.778mm (0.015 to 0.070 inches). Given current assembly techniques known in the art, a hydraulic diameter of approximately 0.889mm (0.035 inches) optimizes ultimate heat transfer efficiency and ease of construction. Hydraulic diameter is as conventionally defined, namely, the cross-sectional area of each of the flow paths multiplied by four and in turn divided by the wetted perimeter of the corresponding flow path. The values of hydraulic diameter given are for condensers in R-12 systems. Somewhat different values might be expected in systems using a different refrigerant. Within that range, it is desirable to make the tube dimension across the direction of air flow through the core as small as possible. This in turn will provide more frontal area in which fins, such as the fins 34, may be disposed in the core without adversely increasing air side pressure drop to obtain a better rate of heat transfer. In some instances, by minimizing tube width, one or more additional rows of the tubes can be included. In this connection, the preferred embodiment contemplates that tubes with separate spacers such as illustrated in Figure 2 be employed as opposed to extruded tubes having passages of the requisite hydraulic diameter. Current extrusion techniques that are economically feasible at the present for large scale manufacture of condensers generally result in a tube wall thickness that is greater than that required to support a given pressure using a tube and spacer as disclosed herein. As a consequence, the overall tube width of such extruded tubes is somewhat greater for a given hydraulic diameter than a tube and spacer combination, which is undesirable for the reasons stated immediately preceding. Nonetheless, the invention contemplates the use of extruded tubes having passages with a hydraulic diameter within the stated range. It is also desirable that the ratio of the outside tube periphery to the wetted periphery within the tube be made as small as possible so long as the flow path does not become sufficiently small that the refrigerant cannot readily pass therethrough. This will lessen the resistance to heat transfer on the vapour and/or conduit side. A number of advantages of the invention will be apparent from the data illustrated in Figures 3-6 inclusive and from the following discussion. Figure 3 for example, on the right-hand side, plots the heat transfer rate against the cavity or hydraulic diameter at air flows varying from 12.74 to 90.61m 3 (450 to 3200 Standard Cubic Feet) per minute for production condenser cores made by the applicant. Heat transfer rate is plotted in kW (thousands of BTU per hour) and the hydraulic diameter is plotted in mm (inches). To the left of this data are computer generated curves based on a heat transfer model for a core made according to the present invention, the model constructed using empirically obtained data. Various points on the curves have been confirmed by actual tests. The curves designated "A" represent heat transfer at the stated air flows for a core such as shown in Figure 1 having a frontal area of 0.186m 2 (two square feet) utilizing tubes approximately 0.61m (24 inches) long and having a 0.381mm (0.015 inch) tube wall thickness, a 13.51mm (0.532 inch) tube major dimension, 43.3°C (110°F) inlet air, 82.2°C (180°F) inlet temperature and 1.619 MPa (235 psig) pressure for R-12 and assuming 1.1 degree C (2 degree F) of subcooling of the exiting refrigerant after condensation. The core was provided with 18 fins per 25.4mm (inch) between tubes and the fins were 15.88mm (0.625 inches) by 13.72mm (0.540 inches by 0.152mm (0.006 inches). The curves designated "B" show the same relationship for an otherwise identical core but wherein the length of the flow path in each tube was doubled i.e., the number of tubes was halved and tube length was doubled. As can be appreciated from Figure 3, heat transfer is advantageously and substantially increased in the range of hydraulic diameters of about 0.381 to 1.778mm (0.015 to 0.070 inches) through the use of the invention with some variance depending upon air flow. Turning now to Figure 4, actual test data for a core made according to the invention and having the dimensions stated in Table 1 below is compared against actual test data for a condenser core designated by the applicant as "1E2803". The data for the conventional core is likewise listed in Table 1 below. In Figure 4: heat transfer rate is plotted in kW (thousands of BTU per hour); air flow rate is plotted in m 3 (Standard Cubic Feet) per minute; refrigerant flow is plotted in kg (pounds) per hour; and refrigerant pressure drop is plotted in kPa (PSI). Both the core made according to the invention and the conventional core have the same design point which is, as shown in Figure 4, a heat transfer rate of 7.62kW (26,000 BTU per hour) at an air flow of 50.97m 3 (1800 Standard Cubic Feet) per minute. The actual observed equivalence of the two cores occurred at 8.21kW (28,000 BTU per hour) and 56.63m 3 (2,000 standard cubic feet) per minute; and those parameters may be utilized for comparative purposes. Viewing first the curves "D" and "E" for the prior art condenser and the subject invention respectively it will be appreciated that refrigerant flow for either is comparable over a wide range of air flow values. For this test, and those illustrated elsewhere in Figures 4-6, R-12 was applied to the condenser inlet at 1.619MPa (235 psigj at 82.2°C (180°F). The exiting refrigerant was subcooled 1.1 degrees C (2 degrees F). Inlet air temperature to the condenser was 43.3°C (110°F). The greater refrigerant side pressure drop across a conventional core than that across a core made according to the invention suggests a greater expenditure of energy by the compressor in the conventional system than in the one made according to the subject invention as well. Curves "F" and "G", again for the prior art condenser and an embodiment of the condenser of the subject invention, respectively, show comparable heat transfer rates over the same range of air flows. Curves "H" and "J" respectively for the conventional condenser and the condenser of an embodiment of the subject invention illustrate a considerable difference in the pressure drop of the refrigerant across the condenser. This demonstrates one advantage of the invention. Because of the lesser pressure drop across the condenser when made according to the invention, the average temperature of the refrigerant, whether in vapour form or in the form of condensate will be higher than with the conventional condenser. As a consequence, for the same inlet air temperature, a greater temperature differential will exist which, according to Fourier's law, will enhance the rate of heat transfer. There will also be a lesser air side pressure drop in a core made according to an embodiment of the invention than with the conventional core. This is due to two factors, namely, the lesser depth of the core and the greater free flow area not blocked by tubes; and such in turn will save on the fan energy required to direct the desired air flow rate through the core. Yet, as shown by the curves "F" and "G" the heat transfer rate remains essentially the same. It has also been determined that a core made according to an embodiment of the invention, when compared with the conventional core, holds less refrigerant. Thus, the core of embodiment of the invention reduces the system requirement for refrigerant. Similarly, there is lesser space required for installation of the inventive core because of its lesser depth. As can be seen from the table, and in consideration with the data shown in Figure 4, it will be appreciated that a core made according to the invention can be made of considerably lesser weight than a conventional core. Thus, Figure 5 compares, at various air velocities, the heat transfer rate per unit mass of core of the conventional condenser (curve "K") versus heat transfer per unit mass of core of a condenser made according to the invention (curve "L" ). In Figure 5 heat transfer rate per unit mass is plotted in W kg -1 (BTU per pound) and air flow is plotted in m 3 (Standard Cubic Feet) per minute. Thus Figure 5 demonstrates a considerable weight savings in a system may be obtained without sacrificing heat transferability by using the core of the present invention. CONDENSER CORE PHYSICAL PROPERTIES FOR FIGS. 3-6 INCLUSIVE CORE PROPERTIES CURRENT PRODUCTION 1E2803 PRESENT INVENTION Depth mm (in.) 24.97 (.938) 13.72 (.540) Heights mm (in.) 311.81 (12.276) 304.8 (12.00) Length mm (in.) 612.90 (24.13) 599.19 (23.259) Face Area m 2 (ft. 2 ) 0.191 (2.057) 0.18 (1.938) Weight kg (lbs.) 2.577 (5.682) 0.933 (2.057) Ratio outside surface / inside surface 4.478 5.391 FIN PROPERTIES Fins per 25.4mm 12 18 Fins Rows 13 21 Fin Thickness mm (in.) 0.203 (.008) 0.102(.004) Fin Height mm (in.) 19.06 (.7502) 12.75 (.5018) Free Flow Area m 2 (ft. 2 ) 0.134 (1.444) 0.144 (1.554) Surface Area m 2 (ft. 2 ) 3.45 (37.110) 3.102 (33.389) Hydraulic Diameter mm (in.) 3.312 (.1304) 2.311 (.0910) Fin Weight kg (lbs.) 0.981 (2.163) 0.450 (.993) TUBE PROPERTIES No. Circuits 2 20 Tube Rows 14 20 Tube Thickness mm (in.) 4.75 (.187) 1.91 (0.75) Tube Wall mm (in.) 0.686 (.027) 0.381 (.015) Tube Length mm (ft.) 385.27 (15.168) 51.99 (2.047) Free Flow Area mm 2 (in. 2 ) 100.39 (.1556) 206.45 (.3200) Hydraulic Diameter mm (in.) 2.0 (.07871) 0.767 (.0302) Outside Tube Surface m 2 (ft. 2 ) 0.412 (4.431) 0.325 (3.494) Inside Tube Surface m 2 (ft. 2 ) 0.862 (9.276) 0.636 (6.842) Tube Weight kg (lbs.) 1.596 (3.519) 0.483 (1.064) Figure 6, in curve "M" thereon, illustrates the air side pressure drop, plotted in Pa (inches of water), for a conventional core and for a core according to the invention for various air flows plotted in m 3 (Standard Cubic Feet) per minute. Curve "N" illustrates the air side pressure drop for the core of the present invention. It will be appreciated that the air side pressure drop, and thus fan energy, is reduced when a core made according to the invention is utilized.

Claims: A heat exchanger for exchanging heat between the ambient and a refrigerant that may be in a liquid or vapour phase, comprising: a pair of spaced generally parallel headers (10,12), one of said headers having a refrigerant inlet (26 or 32), and one of said headers having a refrigerant outlet (32 or 26); and a heat exchanger tube (20) extending between said headers (10,12) and in fluid communication with each of said headers, said tube having a generally flat cross-section and defining a plurality of hydraulically parallel refrigerant flow paths (46,48,50,52,54,58,60) between said headers, each of said refrigerant flow paths (46,48,50,52,54,58,60) having a hydraulic diameter up to 1.778mm (0.07 inches); characterised in that: said headers (10,12) each have a series of openings (14) with the openings in the series on one header being aligned with and facing the openings in the series of the other header; said heat exchanger tube comprising a tube row defined by a plurality of straight tubes (20) of generally flat cross section and extending in parallel between said headers, the opposed ends of said tubes (20) being disposed in corresponding aligned ones of said openings (14) and in fluid communication with the interiors of said headers (10,12), at least some of said tubes (20) being in hydraulic parallel to each other; webs (40) within said tubes extend between and are joined to opposed side walls (42) of the tubes at spaced intervals to (a) define a plurality of non-circular flow paths (46-60) within each tube (20), (b) absorb forces resulting from internal pressure within said heat exchanger and tending to expand said tubes (20), and (c) conduct heat between fluid in said flow paths and both said opposed side walls of said tubes; said webs and/or said flat side walls define at least one concave zone at the intersection of converging surface segments in each of said flow paths extending along the length thereof; and serpentine fins (34) incapable of supporting said tubes (20) against substantial internal pressure extend between facing ones of said opposed side walls of adjacent tubes. The heat exchanger of claim 1 characterised in that said outlet (32 or 26) is a condensate outlet and said heat exchanger is a condenser. A heat exchanger according to claim 1 or claim 2 characterised in that between said openings (14), each of said headers (10,12) is provided with a part-spherical dome. The heat exchanger of any preceding claim characterised in that said webs are defined by an undulating insert (40) bonded to said opposed side walls (42). The heat exchanger of any preceding claim characterised in that there are a plurality of said concave zones for at least some of said flow paths. A heat exchanger according to any preceding claim characterised in that the web is joined to said flat side walls by fillets (44) of solder or braze metal. A heat exchanger according to any preceding claim characterised in that said headers (10,12) are defined by generally cylindrical tubes. A heat exchanger according to any preceding claim characterised in that the non-circular flow paths (46 - 58) are discrete.

Inventors: DUDLEY JACK C, GUNTLY LEON ARNOLD, DUDLEY, JACK C., GUNTLY, LEON ARNOLD
Applicants: MODINE MFG CO, MODINE MANUFACTURING COMPANY
Application year: 1986
Main classification: F28D1
IPC codes: F28D1, F25B39, F28F3, F28F9