Title of Invention	MULTIFLOW TYPE CONDENSER FOR AUTOMOBILE AIR CONDITIONER
Abstract	A multiflow type condenser for an automobile air conditioner comprising: a pair of header pipes disposed in parallel with each other and arranged to have an inlet and an outlet pipes; a pluratlity of flat tubes each connected to said header pipes at opposite ends thereof, each of said flat tubes having a plurality of inside fluid paths, a hydraulic diameter of each of said inside fluid paths being in the range of about 1 to 1.7 mm; a plurality of corrugated fins each disposed between adjacent flat tubes; at least a pair of baffle disposed in said header pipes one by one; each of said baffles having a projection inserted into a slit provided with each header pipes and dividing each header pipes into a plurality of chambers; at least one by-pass passageway formed in baffle to route a vapor-abundant phase of said refrigerant from an upper chambe'r to a lower chamber within the same header pipes by providing a communication path between the adjacent chambers; a ratio of a hydraulic diameter of said by-pass passageway over said hydraulic diameter of each of said inside fluid paths being in the range of about 0.28 to 2.25; and an area of a pass on the inlet side is about 30% to 65% of an overall area of all of said passes.

Title of Invention

MULTIFLOW TYPE CONDENSER FOR AUTOMOBILE AIR CONDITIONER

Abstract

A multiflow type condenser for an automobile air conditioner comprising: a pair of header pipes disposed in parallel with each other and arranged to have an inlet and an outlet pipes; a pluratlity of flat tubes each connected to said header pipes at opposite ends thereof, each of said flat tubes having a plurality of inside fluid paths, a hydraulic diameter of each of said inside fluid paths being in the range of about 1 to 1.7 mm; a plurality of corrugated fins each disposed between adjacent flat tubes; at least a pair of baffle disposed in said header pipes one by one; each of said baffles having a projection inserted into a slit provided with each header pipes and dividing each header pipes into a plurality of chambers; at least one by-pass passageway formed in baffle to route a vapor-abundant phase of said refrigerant from an upper chambe'r to a lower chamber within the same header pipes by providing a communication path between the adjacent chambers; a ratio of a hydraulic diameter of said by-pass passageway over said hydraulic diameter of each of said inside fluid paths being in the range of about 0.28 to 2.25; and an area of a pass on the inlet side is about 30% to 65% of an overall area of all of said passes.

Full Text	This invention relates to a multiflow type condenser for use in an air conditioning system, and more particularly to a condenser for automobiles in which high efficiency of heat transfer is acieved by pormitting a liquid refrigarant changed in phase during condensing process to be by-passed between chambers formed in the headers of the condenser Recent, tendency t o condensers for automobiles, which receive a gaseous refrigerant, condense the refrigerant through heat exchange with an air, and then discharge to an evaporator via an expansion means, shows compact designs with high performance or heat, exchange according to the demand of small size and lightweight for can-related parts. As typcal one, a parallel flow type condenser includes a plurality of f1at tubes with a plurality of corrugated fins, each corrugated fins being intervened between adjacent flat tribes, and a pair of headers to which each flat tube i s connected at both ends thereof. For aid in understanding, referring to FIG . 13, the para11el f1ow type condenser 60 includes first and second headers 61 and 67, a plurality of flat tubes 63, and a plurality of corrugated fins 64 disposed between adjacent flat tubes 63 . Both ends of each flat tube are connected between the first and second headers 61 and 62, and at least one baffle 65 is provided within each header 63, 62 so that there are furnished multiple passes each defined by f1at tubes 63. The ref r i geranl flows through the condensor in a zigzag pattern. The condenser with the above construction is smaller in size, more lightweight, and yet of high efficiency in heat transfer than the conventional serpentine type condensers. Therefore, the paral1el flow type condenser is widely employed in an automobile a air conditioning system. In general, the refrigerant is introduced into a condenser in a vapor phase, and as the rofrigerant f1ows from an inlet toward an outlet the refrigerant is completely changed into a liquid phase in the area on the outlet, side after experiencing a gas/1iquid two-phase state. Accordingly, the refrigerant exits the condenser i n 1iquid phase to an external element of a refrigerant circuit. Namely, a vapor-abundant phase of the refrigerant flows through an upper area of the condenser, whi1e a 1iquid-abundant phase condensed from the vapor phase gradually increases approaching to an lower area of the condenser, and therefore, it appeals that two-phase of the refrigerant flows through the condenser as a whole. During phase change of the refrigerant, thin liquid film formed on an inside wall of each flat tube positioned in the area through which the vapor-adundant phase flows acts as a thermal resistance to binder heat transfer between the refrigerant and the air, and furthermore, due to rapid flow rate of the vapor phase as compared with the liquid phase, the liquid film serves as a flow resistance to f1ow of the refrigerant through the condenser so that a pressure drop, i.e. pressure loss takes place between the inlet and the outlet, which necessarily increases syst em energy requirements . Commonly, it is important in designing a condenser to have an increased area for heat transfer and yet a lower pressure drop on the refrigerant side in order to enhance the performance of t he condenser. As methods of increasing the heat transfer area, i.e. an effective cross-sectional area of the flat tubes for the refriqerant passage, there would be considered two alternatives; one is to decrease a hydraulic diameter of each of inside flow paths which are formed within each flat tube to allow the refriqerant to be passed therethrough, while the other is to increase the number of passes so as to make the length of the overall fluid paths for the refrigerant: passage to be longer, each pass being composed of a plurality of flat tubes. As one of decreasing the hydraulic diameter of inside flow paths, U.S. Patent No. 4,998, 580 discloses a tube having a plurality of fluid flow paths formed by a undulating spacer within the tube, Each of the fluid flow pahts has a very small hydraulic diameter. llowover, the hydraulic diameters of the fluid flow paths are so small that an excessive pressure drop is provoked due to the corresponding increase of refrigerant passage resistance. In a condenses to which the tubes each having such a small size of fluid flow paths are applied, the overall length of fluid paths for the refregerant passage becomes shorter than a condenser with relatively large hydraulic diameters or more passes, since a mall number of passes for the refrigerant passage must be maintained in i order to prevent the exvessive pressure drop in the above mentioned construction of condenser. Accordingly, in U.S. Patent No. 4 , 988,580, if the number of the refrigerant passes increases, for example, over three, too much pressure drop on the refrigerant side occurs and results in increment of system energy requirements. As the method of increasing the overall fluid paths for the refrigerant passage, as shwon in FIG. 13, a plurality of baffles or partitions are provided in the headers, by provision of which the refrigerant introduced into the condenser flows across the condenser in a zigzag fashion, and as a consequence, the effect of increasing the effective cross-sectional area of tubes is obtained. It seems that this type in designing the condenser for use in automobile air conditioning system is utilized more. In this construction of condenser, considering the phase change of the refrigerant from vapor into Liquid occurring during passage of the refrigerant through the condenser, an effective area or the number of tubes in the uppermost pass on the inlet side is relatively larger find effective areas of passes are progressively reduced toward the lowermost pass on the outlet side because of large with the liquid refrigerant. By these constructions, most heat exchange takes place in the uppermost pass on the inlet side and, in addition, the flow resistance of refrigerant across the condenser- is reduced as well . however, when the excessively small hydraulic diameter of thubes on the too long fluid paths is selected to enhance the heat transfer efficiency of condensers, the heat transfer effiency increases, whi1e load exerted on a compressor rises according to the increase of pressure drop duo to large flow resistance of the refrigerant between the inlet and the outlet of the condenser. Accordingly, to prevent the excessive pressure drop from being took place and to obtain the desired heat transfer efficiency, it is required that the number of U-turns in flow of the refrigerant be minimized for the condenser tubes with a small size of hydraulic diameters and the number of U-turns be at least two for the condenser with tubes of relatively large hydraulic diameters. In the meantime, for a condenser in which the length of fluid pahts of the refrigernat is established long by allowing the refrigerant to flow in a zigzag fashion because of provision of at; least one baffle in the headers, prior arts are known that includes by-pass passage ways formed at the center of the baffles to make a pressure drop according to increse of the fluid path length to be minimized and to permit a 1iquid refrigerant condensed passing through passes to be by-passed to an outlet side of the condensor. As one example, U.S. Patent No. 4,243,094 discloses a condenser including a pair of headers, a plurality of tubes with fins surrounded each of the tubes, and baffles having a bore. Bores are of a size which allows the condensed liquid through each pass to flow therethrough by capi11ary action into a adjacent lower chamber in the same header without passing through a subsequent pass. 094 Patent describes that centrally disposed bores are so smal1 that they act as capillary tubes and effectively prevent gaseous f1uid from passing therethrough. Therefore, the bores insure that only fluid in a 1iquid state will be passed therethrouqh . However, since '094 Patent does not mention expressly the number of passes for the refrigerant passage, sizes of the hydraulic diameter of tubes and the bores (by pass passageways/, and relation therebetween, it seems to be hard to apply '094 Patent to actual dosign of condenser i n that t he desired heat transfer efficiency could be obtained when how many passes for the refregerant passage are selected, that the size of by-pass passageways must bo defined to what extent, and that how the by-pass passageways should be established in view of the1 number of passes for the refrigerant passage and the hydraulic diameter of tubes. Furthermore, it is difficult to form bores in the baffles and to dispose the baffles within the headers, considering that the bores should have a small diameter and a long length to accomplish capillary action in fluid f1ow. Another prior art concerning by-pass of the condensed liquid refrigerant, Japanese Unexamined Utility Model No, 63-173688 (applioation No. 62-064734) discloses, as shown in FIGS.14 and 1ba and 15b, a condenser including a pair of headers 70 having tubes 78 each connected to the headers at both ends thereof, and baffle means 73 having an upper member 74, a meshed member 77 and a lower member 75, Baffle means 73 divides an internal space of each header 70 into upper and 1ower chambers 71 and 72, respectively. Each upper and lower member 74 and 7b is provided with a hole 76, and liquid refregerant is by-passed from the upper chamber 71 into the lower chamber 72 through the holes 76 and the moshed member 77, However, the condenser with the above construction does not disclose the relation between the heat transfer efficiency and the pressure drop, the number of passes for the refregerant passage, the si ze of by-pass passageways, and the relation therebetween, except simple description about by-passing of the liquid refrigerant through the by-pass passageways formed in baffle means. SUMMARY OF THE INVENTION The present invention is directed to overcoming one or more of the above problems, has its object to provide multiflow type condenser wherein the condenser echances a heat transfer efficiency and minimizes a pressure drop on the refrigerant side as i well, by differentiating an effective area of each pass for an refrigerant passage in consideration of a phase of the refregerant flowing through the passes . Another object of the present invention is to provide a condensor which effectively by-passes a 1iquid refrigerant by optimizing a size of by-pass passageways according to a hydraulic diameter of tube. Still another object of the present invention is to provide a condenser with a by-pass passageway to be easily formed. According to the present invention, there is provided a multiflow type condenser for an automobile air conditions comprising. a pair of header pipes disposed in parallel with each other and arranged to have an inlet and on outlet, said header pipes being ol1iptical in cross-section; a plurallily of flat tubes each connected to said header pipes at opposite ends thereof, each of said that tubes having a plurality of inside fluid paths, a hydraulic diameter of each of said inside fluid paths being in the range of about 1 to 1.7 mm; a plurality of corrugated fins each disposed between adjacent at least a path of baffle disposed in said header pipes one by one; each of said baffles having a projection inserted in to a slit provided with header pipes and dividing each header pipes into a plurality of chambers so that a refregerant flows through a plurality of passes each defined by a plurality of tubes in zigzag fashion between said inlet and said out let, an outer peripheral sunface of each baffle contacting with an Inner peripheral surface of said respective header pipes. at least one by-pass passsgeway formed around a position at which the chambers in each header pipe are divided by the baffle therein to route a vapor-abundant phose of siad refrigerant from an upper chamber to a lower within the same header pipes by providing a communication path between the adjacent chambers; a ratio of a hydraulic diameter of said by-pass passageway over said hydraulic diameter of each of said inside fluid paths being in the range of about 0.205 to 2.25, and an area of a pass on the inlet side defined by the chamber on the inlet side into which said refregerant is introduced through said inlet and formed in one of said header pipes, the opposed chamber formed the other os said header pipes, and a plurality of tubes extending between the chambers is 30% to 65% of an overal1 area of all of said passes. BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a front view of a condenser according to the present i nvention . FIG. 2 is a partially exploded perspective view showing i the joining relation between header pipes and baffles, and between header pipes and tubes. FIG. 3 is a sectional view taken along a line 11-13 accordi ng to one embodiment of t he present invention. FIG. 4 is a sectional view showing a by-pass passageway accordi ng to another embodiment of the present invention, FIG. 5 is a sectional view showing a by-pass passageway ■i according to st i 11 another embodiment, of the present invention FIG. 6a and 6b show examples of forming a by-pass - i passageway in out line. FIG. 7 shows a refrigerant circuit of an automobile air conditioning system. FIG, 8 is a p-h diagram of the refrigerant circuit of FIG. 7. FIG. 9 is a graph showing the a relationship between a heat transfer efficiency and a pressure drop according to variations of the size of by-pass passageway versus a hydraulic diameter of a tube . FIG. 10 is a graph showing a relationship between heat, transfer or efficiency and a pressure drop according to variations of the ratio of the number of tubes constituting a pass on an inlet side with respect to the overal1 tubes. FIG. 11 is a graph showing a relationship between a heat transfor efficiency and a pressure drop with respect to variations of hydraulic diameter of tube , FIG. 12 is a graph showing a relationship between a heat transfer efficiency with respect to variations of the nember of passes. FIG 13 is a front view of a conventional condenser. FIG 14 is an enlarged sectional view of elements around a baffle means of another conventional condenser, FIG, 15a and 15b are a perspective view and a exploded view, respectively, of the baffle means of FIG. 14. DESCRIPTION OF THE PREFERRED EMBODIMENTS Referring now to FIG. 1, there is shown a condenser 10 which comprises a plurality of flat tubes 11 disposed in parallel relation, and a plurality of corrugated fins 12, each fin 12 being intervened between adjacent flat tubes 11.. Each of the flat tubes, 11 is connected to a first header pipe 13 at its one end, and to a second header pipe 14 at the other end thereof. The condenser also has a pair of side plate 20 and 21 disposed at the positions thereof. Both ends of each of the header pipes 13 and 14 are closed by blind caps 17 and 18. An inlet pipe 15 is connected to the first header pipe 15 adjacent its upper end and outlet pipe 16 is also connected to the first header pipe 15 adjacent its lower end. Though both inlet and outlet pipes 15 and 30 are shown as connected t o the first header pipe 13, t he pipes 15 and 16 may be connect ed to the f irst and second header pipes 13 and 14 t respectivel y, according to changes in the number of passes for t he refrigerant passage. Both the first, and second header pipes 13 and 14 certain t herei n baffles 19 adapted to def ine a plurality of passes for the refrigerant passage, each pass being defined by a plurality of that tubes 11. In this embodiment of FIG. 1, there are defined passes P1 , P2, P3 and P4, and t he number of passes is changed according to the number of baffles 19. In the multiflex type condenser with t he above constitution, the refrigerant f lows through the passes in a zigzag fashion, until the refregerant is drawn off through the outlet pipe 16 after introduct ion into the condense 10. In an examplary embodiment of the condense of FIG. 1 , by the baffles, three chambers 13a, 13b and 13c are defined in the first header pipe 13, find two chambers 14a and 14b in the he second header pipe 14 , Tinning now to FIGS. 2 and 3, as shown therein, each flat tube 11 includes a plurality of inside fluid paths 11a each defined by inside walls. Each of the holder pipes 13 and 14 is made of a header 22 and a tank 23, both components form together an elliptical cross-section. Preferably, the shpae of cross-section of each tank 23 is semi-circular so as to reduce f1ow resistance of the rofrigerant in the header pipes. Otherwise, the header pipes 13 and 14 mav have a circular cross-section without consisting of two components. The header pipes; 13 and 14 wi th a circular cross- section can be manufactured by seaming or by extrusion using such as a clad aluminium plato . Headers 22 are provided with a plurality of slots 24 through which flat tubes 11 are inserted and brazed. Baffles 19 are positioned within the header pipes 13 and 14, and the outer circumferences of the baffles 19 follows the inner circumferences of the header pipes 13 and 14 so that the outer circumferential surfaces of the baff1es 19 contact the inner circumferenti al surfaces of the header pipes 13 and 14 when the header pipes 13 and 1 4 and the baff1es 19 are joined together. Otherwise, grooves (not shown) are formed on the inner surfaces of header pipes 13, 14 at which the baffles 19 are positioned, and the size of each baffle 19 is defined such that the outer circumforential surface of baffle is fitted into the grooves. Each baffle 39 is provided with a projectio 26 outwardly extended therefrom, and the projection 26 is inserted into a slit 27 formed in the tank 23 of each header pipe 13, 14 . When the baffle 19 is fi11ed into the slit 27, ( he projection 26 extends outward each header pipe 33, 34 to allow the outwardly oxtended portion of the projection to be pressed on the external surface of each header pipe 13, 34 and to cover the slit 27 by caulking or other methods. By doing this, it is possible to prevent the baffles 19 from both displaced during movement of the condenser for brazing and no loakage of refrigerant is 1ikely to occur . Each baffle 19 is provided with at least one by-pass means. One embodiment of a by- pass passageway according to the invention is shown in FIG. 3. Referring to FIG. 3 together with FIG.2, at least one cut-out portion 25 is formed in the outer peripheral portion of the baffle 39 by press working at the same time of making the baffle 19. A by-pass passageway 25a is provided when t lie baffle 19 is combined with the respective headers 13,14 so that a liquid refrigerant changed from a vapor phase is passed therethrough. Namely, the by-pass passageway 25a provides a communication path between adjacent chambers among the chambers 13a, 13b, 13c, 14a and 14b each defined by the header pipes 13 and 14 and the baffles 19 so as to directly rout some of the liquid refregerant condensed throuqh passes from chamber to chamber. The by passageway 25a may formed at caentral portion of the baffle 19, but preferebly, is formed in the outer peripheral portion of the baffle because of case in machining. When the by-pass passegeway 26a, i.e current portion 25 is formed at the central portion of the baffle 19, there also problems in that the by-pass passageway should be machined after firstly forming the baffle 19 and that the machining tool can not enduse long because the tools should be thin when the by-pass passageway formed is smaller than a give nice. However, forming of the by pass passageway 25a in the outer peripheral portion of the baffle 19 makes its formation easy because not only formation of the baffle 19 and the cut-out portion 25 can be made in a lump, but also it is advantageous to move the position of by pass passageway in view of the refregerant flow charecteristics, Turning now to FIG 4, a further embodiment of the by pass passageway is shown. In this embodiment , a by-pass passageway 20 is formed on an inside surface of each header pipe 13.14, The by-pass passageway 28 can be formed along the longitudinal axis of each header pipe 13, 14 by extension of roll forming or only on the positio at which the baffle 19 is disposed by press working. FIG 5 shows still another embodiment of the by-pass passageway and FIG 6a and 6b show metods of maching the by-pass passageway in cutline. As shown, an embodiment is illustrated to supplement defects in machining according when the by--pass passageway is formed in a central portion of the baffle 19, and to effectively by-pass the liquid refregerant. In this embodiment, a by-pass passageway 29 is made by lancing, burring or scratching. namely, a portion in which the by-pass passageway 29 is formed is net cut off them the baffle 19, and the portion has a tolded portion has which quids the liquid refregerant at the time of by- passing. Referring now to FIG.7 , a refregerant circuit is includes a compressor 36, a condenses 31, a expansion mechanism 30 and an evaporator 39. In the refregerant circuit 35, the refregerant is compressed in the compressor 36 to high pressure of about the 20 kg/cm2 and sent to the condenser 37 . High pressure from the compressor 36 is applied to an inlet 1 of t he condenser 37, the refregerant is changedfrom vapes to liquid f lowing through the passes of the condenser 37 (4 passes as shown in FIG. 1), and then, exits from the condensor 37 through an outjel 0. The pressure and temperature of the liquid refregerant drop to about 2- 5kg/cm passing through the expansion mechanism and the refregerant is introduced in to the evaporation 30 in which heat exchange takes place between the refregerant and air. Thereafter, the refregerant travels into the compressor 36 and circulates t he refregerant circuit. FIG 8 is a diagram showing an ideal cycle and an cycle of the refregerant circum of FIG. 7. As shown in FIG, there occures no pressure drop dpr on the refregerant side flewing through the condensor 37 in the ideal refregerant cycle 10, while in the actual cycle AC, a certain range of pressure drop dPr takes place because refrigerant is subject to a flow resistance at the time the refregerant travels throuqh the passes for the refregerant passage. Namely, when measurement is made to the actual refregerant cycle AC, i.e. between the inlet 1 and outlet 0 of the condensor 37, a certain range of pressure drop occurs irrespective of presence of the by-pass passageways. In addition, a pressure drop also occurs on the air side passinq through the corrugated fins 12 (FIG 1) . Excessive drops both on the refr iqerant and all sides increase the operation of compressor, and in turn, the system energy requiroments . As the desiqn of condenser for use in a car ai , conditioner changes from the serpentine type to the parol1el flow type or the multiflow type, a relatively large single tube used for onhausinq the heat transfer efficiency in the serpentina type condensor has been replaced by a plurality of flat tubes. Both ends of each flat tube are connected to spaced and parallel headers SO a to define a plurality of passes for the refrigerant passge. The refregerant enters the condensor throuqh the inlet formed in one header and flows in parallel through each flat tube. Accordinqly, to accomplish a required performance in the parallel flow type condensor , on the hand, the hydralic diameter of flat tubes is restricted within a given range which is smaller than than the normal hydraulic diameter of flat tubes, on the other hand, the condenser in devided by baffle means so as to define a plurality of passes . As described above, restriction of the hydraulic diameter of each of flat on inside fluid paths formed i n flat tubes below a given value increases the beat transfer efficiency and also the passage resistance of refregerant passing through each flat tube of inside fluid path., Accordingly, an excess i ve pressure drop is followed , which in turn, leads to increment of system energy required in the Overa1I refriqerant circuit, and as a consequence, one of few passage most be maint ained. On the other hand, when the diameter of flat tubes is in the normal range, i.e. from about, 1 mm to aobut 1.2 mm, The pressure drop decreases because the passages resistance of refrigerant passing through each flat t ube or inside fluid path in comparison with the flat tubes having below 1 mm hydraulic diameter. Therefore, passes can be defined over the flat tubes with relatively small hydraulic diameters and results in increment of the length of flow paths and the heat transfer efficiency. For refrence, hydruulic diarneter D, is defined as follows:: in which Δ is the cross-sectional area of the tube (each of the inside fluid paths when they are formed within each tube) and P is the welled parameter of the corresponding tube, i.e inside fluid Considering the above mentioned condens. per the condensor with a by-pass passageway, it was discovered that the heat transfer and pressure drop relatioship was found to be improved as described in further detail later when the designs of condensors are made to restrict the hydraulic diameter of that tubes within in certain prescribed limits for minimizing the presure drop of refregerant by reduecinq the passage resisteance of refregerant flowing the tubes, t o by-pass the liquid refregerant from chamber chamber by providinq optim sized by-pass passageways with respect to the hydraulic diameters of flat tubes for preventinq deterjoration of the heat transfer performance according to decrement of the passage resistance of refregerant, and to optimize the effective areas of passes of passes in view of f low charecterstics between the vapour and the liquid, To design the above condenserr the hydraulic diameter was choosen between 1 and 1.7 mm because if the hydraulic diameter of flat tubes is below 1 mm, excessive pressure drop occures and thus, the length of luid paths must be short, and because if the hydralic diameter of flat tubes is beyond 1.7 mm, the length of accordingly the must be long to meet the condensor performance and accordingly, the condenser becomes large. The test was performed for the condensor having the by-pass passageways of the hydraulic diameter of about 1 mm f ormed in t he baffles against the conventional condensor without by-pass passageways. In testing it was found that the condensor with the condenser with the by-pass passageways has the lessur pressure drop and heat transfer efficiency as compared with the condensor without the by-pass passageways . Therefore, another test was performed to ascertain the relatioship between the hydraulic chamber of each by-pass passageway and the hydraulic diameter of each flat tube. In the test, the hydraulic diameter of each flat tube (each inside fluid path when formed in the flat. tube) ranged from 1 to 1.7 mm, and the hydraulic diameter of each by-pass passageway was chosen in the range of twice 1.7 mm to half 0.1 mm corresponding to about 0.5 to 3.4 mm). The result of test is reproduced in FIG. 9 Referring now to FIG. 9, as can be seen, the desired performance of the condensor is not obtained if the ratio of the hydraulic diameter of the by-pass passways over the hydraulic diameter of the tube, D/D is beyond or below a certain prescribed limits. with the condensor having the by-pass passageways, it is seen that the heat transfer efficiency diminishes, on one hand, the press drop is improved , on the other hand . From the results of tests, it was discovered that when the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the tube, D/D is excessively small (below 0.28 as can be seen in the FIG 9), maching of the by-pass passageways and expectation to the effects of by-pass of the liquid refregerant are difficult, while if D/D is excessively large beyond 2.25 in 16.9), to accomplish the principal object of the principal object of providing the by-pass passageways is hard because the possibility increases of by-passing not only the liquid refregerant but also the gasonus refregerant. In addition thouqh the hydraulic diameter of the by-pass passageway over the hydralic diameter of the tube is preferably defined as reverse preportional relationship therebetween, when the hydraulic diameter of the tube is small (below about 1 mm) or large (beyond bout. 1.7 mm), the hydraulic diameter must be chosen in view of the effective areas of the passes with respect to the tubes having the middle, range of hydraulic diameter ranging from 1 mm to 1.7 mm. Analogus results were found with various shapes of the by-pass passageways wherein the tests were executed to the condensers having the by-pass passageways arranged according to these provided with the cut-out portions 25 from the baffles 19 as shown in FIGS. 2 and 3, to those formed in the inside surfaces of the header pipes 13 and 14 in FIG, 4, and to those formed by searching in FIGS. 5 nd 6. This demonstrates that that the position and shape of the by-pass passageway does not affect the condonser performance, Moreover, considering that the liquid refrigerant (padually increases approaching to the lowermost pass, the number and size of the by-pass passageways providing the communication path for the liquid refrigerant between the upper and middle chambers 13a and 13b of the first header pipe 13 must be preferably larger tahn those between the middle and lower chambers 13b and 13c of the first header pipe 13. However, it was ascertained from the tests that there are not much effects in performance between the condensor provided with progressively increased size of by-pass passageways approaching to the lowermost pass and the condenser with the same size of the by-pass passageways. In FIG. 9, curves A and B show that by-passing the condensed liquid refrigerant is focused on improvement of the pressure drop rather than the heat transfer efficiency, and thus, the pressure drop of the condenser with the by-pass passageways improves to some extent but the heat transfer efficiency thereof depreciates. FIG. 9 further shows that the heat transfer efficiency can be improved with respect to the condenser having the by-pass passageways by optimizing the ratio of the hydrulic diameter of the by-pass passageway over the hydraulic diameter of the tube. Accordingly, in testing the condenser in which the effective area of each pass was also considered in addition to the relation between the tubes and the by-pass passageways and in which the effecitve area per pass was changed considering the degree of phase change and the flow rate of refrigerant in each pass, the heat transfer and pressure drop relationship was found to be substantially improved as compared to the conventional designs as described hereinafter. Turning to FIG. 10-12, there are shown the test results with the condenser of the invention and the conventional condenser as changing the hydraulic diameters of the tube and by-pass passageway, and the number of passes, FIG. 10 shows trends between the heat transfer efficiency and pressure drop relation in combination with the effective areas of passes. In this case, the condenser had four passes and the ratio D/D was 0.95. Referring now to FIG. 10, it is seen from curves C and F sersus D and F for the condenser of the present invention and the conventional condenser, respectively, that when the number of tubes constituting a pass on the refrigerant inlet side (the first uppermost pass) over the number of overall tubes constituting all passes of the condenser is within 40%, with both the present and conventional condensers the heat transfer efficiency diminishes while the pressure drop increases. However, when the ratio of the number of tubes of the pass on the inlet side over the number of overall tubes ranges from 40% to 65%, the condenser of the present invention illustrates improvements of performance in both the heat transfer effieiency and pressure drop over the conventional condenser which also has the by-pass passageways . Moreover, in testing the condensors with there and five passes, respectively, it was found for the three pass condenser to operate in optimum performance when the ratio of the number of tubes, which means area, of the pass on the inlet side over the number of overall tubes, which maans the entire area of the condenser, ranges from 55% to 65%, and for the five pass condenser when the ratio ranges from 30% to 45%, Therefore, is is confirmed that the degree of phase change in the pass on the inlet side significant1y aff affecs the heat transfer performance, and the desired heat transfer performance occurs when the relation between the flow rate of liquid refregerant to be by-passed and the passes through which the vapor to be condensed flows without by-pass is selected in optimum. Namely, because the vapor introduced into the condenser through the in1et pipe has a relatively larqe volume and thus, the 1argest volume of the vapor is condensed through the pass on the inlet side, when not by-passing the condensed liquid the flow resistance both the pressure drop and flow resistance occur due to the flow ratio difference between the vapor and the liquid. However, when by-passing the liquid refrigerant of relatively large volume, the vapor flows smoothly through the tubes and through even the passes near the lowermost pass without 1arge difference in flow rate as compared with the flow rate in the pass on the inlet side. My designing the condender with the above conditions, the number of passes can be increased to a certain extent even with the small hydraulic diameter because both the pressure drop and heat transfer of efficiency improve, on one hand, while when utilizing the large hydraulic diameter of tube, the number of passes can also be increased, which means increment of the length of fluid pahts, without the disadvantageous pressure drop, on the other hand. From these facts, it is construed that with the smae sized condensers, superior performance is obtained in the condenser according to the present invention over the conventional designs irrespective of presence of the by-pass passageways, and in turn, more compact condonsers are provided according to the invention when desgn is made to obtain the same performance. Turning to FIG. 11 , there is shown the relation between the heat transfer efficiency and the pressure drop with vairations of the hydraulic diameter of the tube in the range of 1 to 1.7 mm. Prior art ] is a conventional condenser without the by-pass passageways while prior art 11 is the condenser with the by-pass passageways formed in prior art 1. In FIG 11 , the number of passes and the ratios of effective areas of the passes on the inlet side over the number of overall tubes were four and about 30%-40%, respectively, for both prior arts 1 and 11. As can be soon, the pressure drop of prior art 11 with the by-pass passageways is lesser than that of prior art I without the by-pass passageways, while the heat transfer efficiency of prior art 11 deprociates relative to that of prior art 1, and accordingly, the performance of the condenser with the by-pass passageways is inferior to that of the condenser without the by-pass passageways. However, with the condenser accordinq to the present invention to which the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the flat tube and the raito of the area of the pass on the inlet side over the area of overall passes, 30-60% are applied, when the hydraulic diameter of the tube increases, the heat transfer efficiency of the present invention is superior to those of prior arts 1 and 11 whi1e the pressure drop of the present invention is superior to prior art 1, on one hand, inferior to prior art 11, on the other hand. The reason the pressure drop of the present invention is a little higher than that of prior art 11 is construed as more vapor toqhether with the liquid is by-passed in the condenser of prior art 11 than in the condenser of the present invention because the area of the pass on the inlet side of prior art 11 is smaller than that of the present invention. Namely, FIG. 11 shows that the amounts of the vapor condensed through the pass on the inlet side and the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the tube is related with each other irrespective of the hydraulic diameters used in the normal tubes, and the condenser performance is best when the area of the pass on the inlet side is chosen in the range shown in FIG. 10. Accordingly, the desired heat transfer efficiency and pressure drop is accordingly when optimizing the raito of the hydraulic diameter of the by-passageway over the hydraulic diameter of the tube and selecting the number of of tubes constition the pass on the inlet side in a qiven prescribed range according to the number of passes. formed in the condenser. referring now to FIG 12, there are shown the results of tests with variations of the number of passes under the conditions as discribed in relation to FIG 11 in which prior art 1 is a common condenser without the by-pass passageways and prior art 13 with the by-pass passageways but a small area of the pass on the inlet side than that of the present invention. FIG- 12 shows that too many passes accompanies restrictions because increase of the number of passes enhances the heat transfer efficiency but raises the pressure drop. Namely, the heat transfer efficiency increases with rapid rising of the pressure in prior art 1, while in prior art 11, 1 he pressure drop is slow with the inferior heat transfer efficiency to prior art 1, and thus, the same trends are identified as in FIG. 11. On the other hand, with the condenser of the present invention, the heat, transfer efficiency increases but the pressure drop increases slowly, and accordingly, increment of the number of passes to a given extent under the smae conditions accompanies lessor restrictions . In consideration of the data as shown in FIGS. 9-12 the performance of condenser in aspects of the heat transfer efficiency and pressure drop is improved by designing in view of three conditions : first, the hydraulic diameter of each tube used in the multiflow type condensor; second, the hydraulic diameter of the by-pass passageways over the hydraulic diameter of the bube ; and finally, the ratio of the number of the tubes constituting the pass on the inlet side over the number of overall tubes constituting the overall passes of the condenser, i.e. the area of the pass ( P1 in FIG.1 ) on the inlet side . Namely, when the hydraulic diameter of each tube ranges from 1 to 1.7 mm, the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the tube, D/D is between 0.28 and 2.25, and the ratio of the pass on the inlet side over the area of the overall pases ranges from 30% to 65%, the performance of parallel flow type condenser is improved as compared to condensers not to fulfi11 the above three conditions, irrespective of presence of the by-pass passageways. For example, optimized performance was observed with the condenser having four passes, and each tube of the ratio of D/D ranging from 0.45 to 1.85, with the ratio of the area of the pass on the inlet side over the area of the overall passes ranging 40% to 55%. The present invention has been described in an illustrative manner. Many modifications and variations of the present invention are possible in light of the above theachings. Therefore, the spi1it and scope of the invention are to be limited ] , A multi f1ow type condensor f or an automobile air condit i oner comprising : a pai r of header pipes disposed in parallel with each other and arranged to have an inlet and an outlet , said header pipes being elliptical in cross section; a plurality of flat tubes vibes each connected to said header pipes at opposite ends thereof , each of said flat tubes having a plurality of inside fluid paths, a hydraulic diameter of each of said inside fluid paths being in the range of about 1 to 1.7 mm; a plurality of corrugated fins each disposed between adjacent flat tubes; at least a pan of baffle disposed in said header pipes one by one; each of said bafles havinq a projection inserted into a slit provided with each header pipes and dividing each header pipes into a plurality of chambers so that a refrigerant flows through a plurality of passes each defined by a plurality of tubes in zigzag fashion between said inlet and said out let , an out er peripheral surface of each baff1e contacting with an inner peripheral surface of said respective header pipes; at least one by-pass passageway formed around a posi tion at which the chambers in each header pipe are di vidad by t.he baffle therein to route a vapor- abundant phase of said refrigerant, from an upper chamber to alowor ehamber within t he same header pipes by providing a communication path between the adjacent chambers ; passageway ever sai d hydraulic diameter of each of sai d in the fluid paths being in the range of about 0.78 to 2.25 and an area of a pass on the inlet side defined by the chamber on the i nlet side i nt o which sai d ref ri gerant is i ntroduced through said inlet and formed in one of said header pipes, the opposed chamber formed the other of said header pipes, and a plurality of tubes extending between the chambers is about 30% to 65% of an ovorall area of all of said passes. /. The condenser of claim 1 , who? ei n 5;ai d passes arc* three and eot i d area oi said pass on the inlet side is about. bbH- to ^'?V d .sa i (J ovo a.) .1 ar oa of sai d passes , sa i d aioa of said pass on the inlet side is about 40* to vo 4 . The condenser of claim ;» , whoj ci n said passes are five ,inb said area said pass on t he inlet side is about 30% to 40% of said overal1 area of said passes. 5. The condenser of claim 1, wherein said by-pass passageway is formed in cent ra l port ions of said baffles,respectively, by sfi ^il chi tiq, (■">. The condenser of o3 aim 1 , wherei n sai d by-pass pass/jfjew,!;/ j s f oj mod in outer peri pher al port 2 ons of said 1M i 1 11 ■:;, respect i vc.l y. 7. The condenser of c] aim ] , wherei n said by-pass passageway i s formed :i n sai d baf fles, respect .i ve.l y, such that a numboi of j >y pass passaqoways pi ogressi ve .1 y .incrofjsos from .said inlet 1 -MI id out] et , It , The condenser of c] ai rn 1 , whoj oi n sai d by-pass passageway .i s f ormod i n sai d baf f ] es, re spool i ve] y, such that said r ai i o ni sa i d hydr au.l :i o d.i arnet or of sai d by-pass passageway ovo t sa i d hydrau.l if: diameter of each of said .inside fluid paths progressively ineroasfs wilhin said ratio from said inlet to said outlet . 9, Tho condenser of claim 1 , wherein said by-pass passn sa id header pipes . 10, The condenser of claim 9, wherein said by-pass passageways one formed such that said ratio of said hydraulic diameter of said by- pass passageway over said hydraulic diameter of each of said inside fluid paths progressive1y increases within said ratio from said inlet to said outlet. 11 . The condenser of claim 1 , wherein said project i on extends outside each header pipes and is pressed around an outer surface of each of said header pipes by caulking means. i 12. A multiflow type condenser for an automobile air conditioner substantially as hereinbefore described with reference to the accompanying drawings. Dated this 25th day of September 1998

Full Text

This invention relates to a multiflow type condenser for use in an air conditioning system, and more particularly to a condenser for automobiles in which high efficiency of heat transfer
is acieved by pormitting a liquid refrigarant changed in phase
during condensing process to be by-passed between chambers formed
in the headers of the condenser

Recent, tendency t o condensers for automobiles, which receive a gaseous refrigerant, condense the refrigerant through heat exchange with an air, and then discharge to an evaporator via an expansion means, shows compact designs with high performance or heat, exchange according to the demand of small size and lightweight for can-related parts. As typcal one, a parallel flow type condenser includes a plurality of f1at tubes with a plurality of corrugated fins, each corrugated fins being intervened between adjacent flat tribes, and a pair of headers to which each flat tube i s connected at both ends thereof.

For aid in understanding, referring to FIG . 13, the para11el f1ow type condenser 60 includes first and second headers 61 and 67, a plurality of flat tubes 63, and a plurality of corrugated fins 64 disposed between adjacent flat tubes 63 . Both ends of each flat tube are connected between the first and second headers 61 and 62, and at least one baffle 65 is provided within each header 63, 62 so that there are furnished multiple passes each defined by f1at tubes 63. The ref r i geranl flows through the condensor in a zigzag pattern. The condenser with the above construction is smaller in size, more lightweight, and yet of high efficiency in heat transfer than the conventional serpentine type condensers. Therefore, the paral1el flow type condenser is widely employed in an automobile a air conditioning system.
In general, the refrigerant is introduced into a condenser in a vapor phase, and as the rofrigerant f1ows from an inlet toward an outlet the refrigerant is completely changed into a liquid phase in the area on the outlet, side after experiencing a gas/1iquid two-phase state. Accordingly, the refrigerant exits the condenser i n 1iquid phase to an external element of a refrigerant circuit. Namely, a vapor-abundant phase of the refrigerant flows through an upper area of the condenser, whi1e a 1iquid-abundant phase condensed from the vapor phase gradually increases approaching to an lower area of the condenser, and therefore, it appeals that two-phase of the refrigerant flows through the

condenser as a whole. During phase change of the refrigerant, thin liquid film formed on an inside wall of each flat tube positioned in the area through which the vapor-adundant phase flows acts as a thermal resistance to binder heat transfer between the refrigerant and the air, and furthermore, due to rapid flow rate of the vapor phase as compared with the liquid phase, the liquid film serves as a flow resistance to f1ow of the refrigerant through the condenser so that a pressure drop, i.e. pressure loss takes place between the inlet and the outlet, which necessarily increases syst em energy requirements .
Commonly, it is important in designing a condenser to
have an increased area for heat transfer and yet a lower pressure drop on the refrigerant side in order to enhance the performance of t he condenser. As methods of increasing the heat transfer area, i.e. an effective cross-sectional area of the flat tubes for the refriqerant passage, there would be considered two alternatives; one is to decrease a hydraulic diameter of each of inside flow paths which are formed within each flat tube to allow the refriqerant to be passed therethrough, while the other is to increase the number of passes so as to make the length of the overall fluid paths for the refrigerant: passage to be longer, each pass being composed of a plurality of flat tubes.
As one of decreasing the hydraulic diameter of inside flow paths, U.S. Patent No. 4,998, 580 discloses a tube having a

plurality of fluid flow paths formed by a undulating spacer within the tube, Each of the fluid flow pahts has a very small hydraulic diameter. llowover, the hydraulic diameters of the fluid flow paths are so small that an excessive pressure drop is provoked due to the corresponding increase of refrigerant passage resistance. In a condenses to which the tubes each having such a small size of fluid flow paths are applied, the overall length of fluid paths for the refregerant passage becomes shorter than a condenser with relatively large hydraulic diameters or more passes, since a mall
number of passes for the refrigerant passage must be maintained in
i order to prevent the exvessive pressure drop in the above mentioned
construction of condenser. Accordingly, in U.S. Patent No.
4 , 988,580, if the number of the refrigerant passes increases, for
example, over three, too much pressure drop on the refrigerant side
occurs and results in increment of system energy requirements.
As the method of increasing the overall fluid paths for
the refrigerant passage, as shwon in FIG. 13, a plurality of
baffles or partitions are provided in the headers, by provision of
which the refrigerant introduced into the condenser flows across
the condenser in a zigzag fashion, and as a consequence, the effect
of increasing the effective cross-sectional area of tubes is
obtained. It seems that this type in designing the condenser for
use in automobile air conditioning system is utilized more. In this
construction of condenser, considering the phase change of the

refrigerant from vapor into Liquid occurring during passage of the refrigerant through the condenser, an effective area or the number of tubes in the uppermost pass on the inlet side is relatively larger find effective areas of passes are progressively reduced

toward the lowermost pass on the outlet side because of large
with the liquid refrigerant. By these constructions, most heat exchange takes place in the uppermost pass on the inlet side and, in addition, the flow resistance of refrigerant across the condenser- is reduced as well .
however, when the excessively small hydraulic diameter of thubes on the too long fluid paths is selected to enhance the heat transfer efficiency of condensers, the heat transfer effiency increases, whi1e load exerted on a compressor rises according to the increase of pressure drop duo to large flow resistance of the refrigerant between the inlet and the outlet of the condenser. Accordingly, to prevent the excessive pressure drop from being took place and to obtain the desired heat transfer efficiency, it is required that the number of U-turns in flow of the refrigerant be minimized for the condenser tubes with a small size of hydraulic diameters and the number of U-turns be at least two for the
condenser with tubes of relatively large hydraulic diameters.
In the meantime, for a condenser in which the length of fluid pahts of the refrigernat is established long by allowing the

refrigerant to flow in a zigzag fashion because of provision of at; least one baffle in the headers, prior arts are known that includes by-pass passage ways formed at the center of the baffles to make a pressure drop according to increse of the fluid path length to be minimized and to permit a 1iquid refrigerant condensed passing through passes to be by-passed to an outlet side of the condensor. As one example, U.S. Patent No. 4,243,094 discloses a condenser including a pair of headers, a plurality of tubes with fins surrounded each of the tubes, and baffles having a bore. Bores are of a size which allows the condensed liquid through each pass to flow therethrough by capi11ary action into a adjacent lower chamber in the same header without passing through a subsequent pass. 094 Patent describes that centrally disposed bores are so smal1 that they act as capillary tubes and effectively prevent gaseous f1uid from passing therethrough. Therefore, the bores insure that only fluid in a 1iquid state will be passed therethrouqh .
However, since '094 Patent does not mention expressly the number of passes for the refrigerant passage, sizes of the hydraulic diameter of tubes and the bores (by pass passageways/, and relation therebetween, it seems to be hard to apply '094 Patent to actual dosign of condenser i n that t he desired heat transfer efficiency could be obtained when how many passes for the refregerant passage are selected, that the size of by-pass

passageways must bo defined to what extent, and that how the by-pass passageways should be established in view of the1 number of passes for the refrigerant passage and the hydraulic diameter of tubes. Furthermore, it is difficult to form bores in the baffles and to dispose the baffles within the headers, considering that the bores should have a small diameter and a long length to accomplish capillary action in fluid f1ow.
Another prior art concerning by-pass of the condensed liquid refrigerant, Japanese Unexamined Utility Model No, 63-173688 (applioation No. 62-064734) discloses, as shown in FIGS.14 and 1ba and 15b, a condenser including a pair of headers 70 having tubes 78 each connected to the headers at both ends thereof, and baffle means 73 having an upper member 74, a meshed member 77 and a lower member 75, Baffle means 73 divides an internal space of each header 70 into upper and 1ower chambers 71 and 72, respectively. Each upper and lower member 74 and 7b is provided with a hole 76, and liquid refregerant is by-passed from the upper chamber 71 into the lower chamber 72 through the holes 76 and the moshed member 77, However, the condenser with the above construction does not disclose the relation between the heat transfer efficiency and the pressure drop, the number of passes for the refregerant passage, the si ze of by-pass passageways, and the relation therebetween, except simple description about by-passing of the liquid refrigerant through the by-pass passageways formed in

baffle means.
SUMMARY OF THE INVENTION
The present invention is directed to overcoming one or more of the above problems, has its object to provide multiflow type condenser wherein the condenser echances a heat transfer
efficiency and minimizes a pressure drop on the refrigerant side as
i
well, by differentiating an effective area of each pass for an refrigerant passage in consideration of a phase of the refregerant flowing through the passes .
Another object of the present invention is to provide a condensor which effectively by-passes a 1iquid refrigerant by optimizing a size of by-pass passageways according to a hydraulic diameter of tube.
Still another object of the present invention is to provide a condenser with a by-pass passageway to be easily formed.
According to the present invention, there is provided a multiflow type condenser for an automobile air conditions comprising.
a pair of header pipes disposed in parallel with each other and arranged to have an inlet and on outlet, said header pipes being ol1iptical in cross-section;
a plurallily of flat tubes each connected to said header pipes

at opposite ends thereof, each of said that tubes having a plurality of inside fluid paths, a hydraulic diameter of each of said inside fluid paths being in the range of about 1 to 1.7 mm;
a plurality of corrugated fins each disposed between adjacent
at least a path of baffle disposed in said header pipes one by
one;
each of said baffles having a projection inserted in to a slit provided with header pipes and dividing each header pipes into a plurality of chambers so that a refregerant flows through a plurality of passes each defined by a plurality of tubes in zigzag fashion between said inlet and said out let, an outer peripheral sunface of each baffle contacting with an Inner peripheral surface of said respective header pipes.
at least one by-pass passsgeway formed around a position at which the chambers in each header pipe are divided by the baffle therein to route a vapor-abundant phose of siad refrigerant from an upper chamber to a lower within the same header pipes by providing a communication path between the adjacent chambers;
a ratio of a hydraulic diameter of said by-pass passageway over said hydraulic diameter of each of said inside fluid paths being in the range of about 0.205 to 2.25, and
an area of a pass on the inlet side defined by the chamber on the inlet side into which said refregerant is introduced through

said inlet and formed in one of said header pipes, the opposed chamber formed the other os said header pipes, and a plurality of tubes extending between the chambers is 30% to 65% of an overal1 area of all of said passes.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a front view of a condenser according to the present i nvention .
FIG. 2 is a partially exploded perspective view showing
i
the joining relation between header pipes and baffles, and between header pipes and tubes.
FIG. 3 is a sectional view taken along a line 11-13 accordi ng to one embodiment of t he present invention.
FIG. 4 is a sectional view showing a by-pass passageway accordi ng to another embodiment of the present invention,
FIG. 5 is a sectional view showing a by-pass passageway
■i
according to st i 11 another embodiment, of the present invention
FIG. 6a and 6b show examples of forming a by-pass
- i
passageway in out line.
FIG. 7 shows a refrigerant circuit of an automobile air conditioning system.
FIG, 8 is a p-h diagram of the refrigerant circuit of
FIG. 7.

FIG. 9 is a graph showing the a relationship between a heat transfer efficiency and a pressure drop according to variations of the size of by-pass passageway versus a hydraulic diameter of a tube .
FIG. 10 is a graph showing a relationship between heat, transfer or efficiency and a pressure drop according to variations of the ratio of the number of tubes constituting a pass on an inlet side with respect to the overal1 tubes.
FIG. 11 is a graph showing a relationship between a heat transfor efficiency and a pressure drop with respect to variations of hydraulic diameter of tube ,
FIG. 12 is a graph showing a relationship between a heat transfer efficiency with respect to variations of the nember of passes.
FIG 13 is a front view of a conventional condenser.
FIG 14 is an enlarged sectional view of elements around a baffle means of another conventional condenser,
FIG, 15a and 15b are a perspective view and a exploded view, respectively, of the baffle means of FIG. 14.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
Referring now to FIG. 1, there is shown a condenser 10 which comprises a plurality of flat tubes 11 disposed in parallel

relation, and a plurality of corrugated fins 12, each fin 12 being intervened between adjacent flat tubes 11.. Each of the flat tubes, 11 is connected to a first header pipe 13 at its one end, and to a second header pipe 14 at the other end thereof. The condenser also has a pair of side plate 20 and 21 disposed at the positions thereof. Both ends of each of the header pipes 13 and 14 are closed by blind caps 17 and 18. An inlet pipe 15 is connected to the first header pipe 15 adjacent its upper end and outlet pipe 16 is also connected to the first header pipe 15 adjacent its lower end. Though both inlet and outlet pipes 15 and 30 are shown as connected t o the first header pipe 13, t he pipes 15 and 16 may be connect ed to the f irst and second header pipes 13 and 14 t respectivel y, according to changes in the number of passes for t he refrigerant passage.
Both the first, and second header pipes 13 and 14 certain t herei n baffles 19 adapted to def ine a plurality of passes for the refrigerant passage, each pass being defined by a plurality of that tubes 11. In this embodiment of FIG. 1, there are defined passes P1 , P2, P3 and P4, and t he number of passes is changed according to the number of baffles 19. In the multiflex type condenser with t he above constitution, the refrigerant f lows through the passes in a zigzag fashion, until the refregerant is drawn off through the outlet pipe 16 after introduct ion into the condense 10. In an examplary embodiment of the condense of FIG.

1 , by the baffles, three chambers 13a, 13b and 13c are defined in the first header pipe 13, find two chambers 14a and 14b in the he second header pipe 14 ,
Tinning now to FIGS. 2 and 3, as shown therein, each flat tube 11 includes a plurality of inside fluid paths 11a each defined by inside walls. Each of the holder pipes 13 and 14 is made of a header 22 and a tank 23, both components form together an elliptical cross-section. Preferably, the shpae of cross-section of each tank 23 is semi-circular so as to reduce f1ow resistance of the rofrigerant in the header pipes. Otherwise, the header pipes 13
and 14 mav have a circular cross-section without consisting of two components. The header pipes; 13 and 14 wi th a circular cross-
section can be manufactured by seaming or by extrusion using such
as a clad aluminium plato .
Headers 22 are provided with a plurality of slots 24
through which flat tubes 11 are inserted and brazed. Baffles 19 are
positioned within the header pipes 13 and 14, and the outer
circumferences of the baffles 19 follows the inner circumferences
of the header pipes 13 and 14 so that the outer circumferential
surfaces of the baff1es 19 contact the inner circumferenti al
surfaces of the header pipes 13 and 14 when the header pipes 13 and
1 4 and the baff1es 19 are joined together. Otherwise, grooves (not
shown) are formed on the inner surfaces of header pipes 13, 14 at
which the baffles 19 are positioned, and the size of each baffle 19

is defined such that the outer circumforential surface of baffle is fitted into the grooves. Each baffle 39 is provided with a projectio 26 outwardly extended therefrom, and the projection 26 is inserted into a slit 27 formed in the tank 23 of each header pipe 13, 14 . When the baffle 19 is fi11ed into the slit 27, ( he projection 26 extends outward each header pipe 33, 34 to allow the outwardly oxtended portion of the projection to be pressed on the external surface of each header pipe 13, 34 and to cover the slit 27 by caulking or other methods. By doing this, it is possible to prevent the baffles 19 from both displaced during movement of the condenser for brazing and no loakage of refrigerant is 1ikely to occur .
Each baffle 19 is provided with at least one by-pass means. One embodiment of a by- pass passageway according to the invention is shown in FIG. 3. Referring to FIG. 3 together with FIG.2, at least one cut-out portion 25 is formed in the outer peripheral portion of the baffle 39 by press working at the same time of making the baffle 19. A by-pass passageway 25a is provided when t lie baffle 19 is combined with the respective headers 13,14 so that a liquid refrigerant changed from a vapor phase is passed therethrough. Namely, the by-pass passageway 25a provides a communication path between adjacent chambers among the chambers 13a, 13b, 13c, 14a and 14b each defined by the header pipes 13 and 14 and the baffles 19 so as to directly rout some of the liquid

refregerant condensed throuqh passes from chamber to chamber. The
by passageway 25a may formed at caentral portion of the baffle
19, but preferebly, is formed in the outer peripheral portion of
the baffle because of case in machining. When the by-pass
passegeway 26a, i.e current portion 25 is formed at the central
portion of the baffle 19, there also problems in that the by-pass
passageway should be machined after firstly forming the baffle 19
and that the machining tool can not enduse long because the tools
should be thin when the by-pass passageway formed is smaller than
a give nice. However, forming of the by pass passageway 25a in the
outer peripheral portion of the baffle 19 makes its formation easy
because not only formation of the baffle 19 and the cut-out portion
25 can be made in a lump, but also it is advantageous to move the
position of by pass passageway in view of the refregerant flow
charecteristics,
Turning now to FIG 4, a further embodiment of the by pass passageway is shown. In this embodiment , a by-pass passageway 20 is formed on an inside surface of each header pipe 13.14, The by-pass passageway 28 can be formed along the longitudinal axis of each header pipe 13, 14 by extension of roll forming or only on the positio at which the baffle 19 is disposed by press working.
FIG 5 shows still another embodiment of the by-pass passageway and FIG 6a and 6b show metods of maching the by-pass passageway in cutline. As shown, an embodiment is illustrated to

supplement defects in machining according when the by--pass
passageway is formed in a central portion of the baffle 19, and to effectively by-pass the liquid refregerant. In this embodiment, a by-pass passageway 29 is made by lancing, burring or scratching. namely, a portion in which the by-pass passageway 29 is formed is net cut off them the baffle 19, and the portion has a tolded portion has which quids the liquid refregerant at the time of by-

passing.
Referring now to FIG.7 , a refregerant circuit is includes a compressor 36, a condenses 31, a expansion mechanism 30 and an evaporator 39. In the refregerant circuit 35, the refregerant is compressed in the compressor 36 to high pressure of about the 20 kg/cm2 and sent to the condenser 37 . High pressure from the compressor 36 is applied to an inlet 1 of t he condenser 37, the refregerant is changedfrom vapes to liquid f lowing through the passes of the condenser 37 (4 passes as shown in FIG. 1), and then, exits from the condensor 37 through an outjel 0. The pressure and temperature of the liquid refregerant drop to about 2- 5kg/cm passing through the expansion mechanism and the refregerant is introduced in to the evaporation 30 in which heat exchange takes place between the refregerant and air. Thereafter, the refregerant travels into the compressor 36 and circulates t he refregerant circuit.
FIG 8 is a diagram showing an ideal cycle and an

cycle of the refregerant circum of FIG. 7. As shown in FIG,
there occures no pressure drop dpr on the refregerant side
flewing through the condensor 37 in the ideal refregerant cycle 10,
while in the actual cycle AC, a certain range of pressure drop dPr
takes place because refrigerant is subject to a flow resistance
at the time the refregerant travels throuqh the passes for the
refregerant passage. Namely, when measurement is made to the actual
refregerant cycle AC, i.e. between the inlet 1 and outlet 0 of the
condensor 37, a certain range of pressure drop occurs irrespective
of presence of the by-pass passageways. In addition, a pressure
drop also occurs on the air side passinq through the corrugated
fins 12 (FIG 1) . Excessive drops both on the refr iqerant
and all sides increase the operation of compressor, and in turn, the
system energy requiroments .
As the desiqn of condenser for use in a car ai ,
conditioner changes from the serpentine type to the parol1el flow
type or the multiflow type, a relatively large single tube used for
onhausinq the heat transfer efficiency in the serpentina type
condensor has been replaced by a plurality of flat tubes. Both ends
of each flat tube are connected to spaced and parallel headers SO
a to define a plurality of passes for the refrigerant passge. The
refregerant enters the condensor throuqh the inlet formed in one
header and flows in parallel through each flat tube. Accordinqly,
to accomplish a required performance in the parallel flow type

condensor , on the hand, the hydralic diameter of flat tubes is restricted within a given range which is smaller than than the normal hydraulic diameter of flat tubes, on the other hand, the condenser in devided by baffle means so as to define a plurality of passes . As described above, restriction of the hydraulic diameter of each of flat on inside fluid paths formed i n flat tubes below a given value increases the beat transfer efficiency and also the passage resistance of refregerant passing through each flat tube of inside fluid path., Accordingly, an excess i ve pressure drop is followed , which in turn, leads to increment of system energy required in the Overa1I refriqerant circuit, and as a consequence, one of few passage most be maint ained. On the other hand, when the diameter of flat tubes is in the normal range, i.e. from about, 1 mm to aobut 1.2 mm, The pressure drop decreases because the passages resistance of refrigerant passing through each flat t ube or inside fluid path in comparison with the flat tubes having below 1 mm hydraulic diameter. Therefore, passes can be defined over the flat tubes with relatively small hydraulic diameters and results in increment of the length of flow paths and the heat transfer efficiency.
For refrence, hydruulic diarneter D, is defined as
follows::
in which Δ is the cross-sectional area of the tube (each of the

inside fluid paths when they are formed within each tube) and P is
the welled parameter of the corresponding tube, i.e inside fluid
Considering the above mentioned condens. per the condensor with a by-pass passageway, it was discovered that the heat transfer and pressure drop relatioship was found to be improved as described in further detail later when the designs of condensors are made to restrict the hydraulic diameter of that tubes within in certain prescribed limits for minimizing the presure drop of refregerant by reduecinq the passage resisteance of refregerant flowing the tubes, t o by-pass the liquid refregerant from chamber chamber by providinq optim sized by-pass passageways with respect to the hydraulic diameters of flat tubes for preventinq deterjoration of the heat transfer performance according to decrement of the passage resistance of refregerant, and to optimize the effective areas of passes of passes in view of f low charecterstics between the vapour and the liquid,
To design the above condenserr the hydraulic diameter was choosen between 1 and 1.7 mm because if the hydraulic diameter of flat tubes is below 1 mm, excessive pressure drop occures and thus, the length of luid paths must be short, and because if the hydralic diameter of flat tubes is beyond 1.7 mm, the length of accordingly the must be long to meet the condensor performance and accordingly, the condenser becomes large. The test was performed

for the condensor having the by-pass passageways of the hydraulic diameter of about 1 mm f ormed in t he baffles against the conventional condensor without by-pass passageways. In testing it was found that the condensor with the condenser with the by-pass passageways has the lessur pressure drop and heat transfer efficiency as compared with the condensor without the by-pass passageways . Therefore, another test was performed to ascertain the relatioship between the hydraulic chamber of each by-pass passageway and the hydraulic diameter of each flat tube. In the test, the hydraulic diameter of each flat tube (each inside fluid path when formed in the flat. tube) ranged from 1 to 1.7 mm, and the hydraulic diameter of each by-pass passageway was chosen in the range of twice 1.7 mm to half 0.1 mm corresponding to about 0.5 to 3.4 mm). The result of test is reproduced in FIG. 9
Referring now to FIG. 9, as can be seen, the desired performance of the condensor is not obtained if the ratio of the hydraulic diameter of the by-pass passways over the hydraulic diameter of the tube, D/D is beyond or below a certain prescribed limits. with the condensor having the by-pass passageways, it is seen that the heat transfer efficiency diminishes, on one hand, the press drop is improved , on the other hand .
From the results of tests, it was discovered that when the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the tube, D/D is excessively small

(below 0.28 as can be seen in the FIG 9), maching of the by-pass passageways and expectation to the effects of by-pass of the liquid refregerant are difficult, while if D/D is excessively large
beyond 2.25 in 16.9), to accomplish the principal object of the principal object of providing the by-pass passageways is hard because the possibility increases of by-passing not only the liquid refregerant but also the gasonus refregerant. In addition thouqh the hydraulic diameter of the by-pass passageway over the hydralic diameter of the tube is preferably defined as reverse preportional relationship therebetween, when the hydraulic diameter of the tube is small (below about 1 mm) or large (beyond bout. 1.7 mm), the hydraulic diameter must be chosen in view of the effective areas of the passes with respect to the tubes having the middle, range of hydraulic diameter ranging from 1 mm to 1.7 mm.
Analogus results were found with various shapes of the by-pass passageways wherein the tests were executed to the condensers having the by-pass passageways arranged according to these provided with the cut-out portions 25 from the baffles 19 as shown in FIGS. 2 and 3, to those formed in the inside surfaces of the header pipes 13 and 14 in FIG, 4, and to those formed by searching in FIGS. 5 nd 6. This demonstrates that that the position and shape of the by-pass passageway does not affect the condonser performance, Moreover, considering that the liquid refrigerant

(padually increases approaching to the lowermost pass, the number and size of the by-pass passageways providing the communication path for the liquid refrigerant between the upper and middle chambers 13a and 13b of the first header pipe 13 must be preferably larger tahn those between the middle and lower chambers 13b and 13c of the first header pipe 13. However, it was ascertained from the tests that there are not much effects in performance between the condensor provided with progressively increased size of by-pass passageways approaching to the lowermost pass and the condenser with the same size of the by-pass passageways. In FIG. 9, curves A and B show that by-passing the condensed liquid refrigerant is focused on improvement of the pressure drop rather than the heat transfer efficiency, and thus, the pressure drop of the condenser with the by-pass passageways improves to some extent but the heat transfer efficiency thereof depreciates. FIG. 9 further shows that the heat transfer efficiency can be improved with respect to the condenser having the by-pass passageways by optimizing the ratio of the hydrulic diameter of the by-pass passageway over the hydraulic diameter of the tube.
Accordingly, in testing the condenser in which the effective area of each pass was also considered in addition to the relation between the tubes and the by-pass passageways and in which the effecitve area per pass was changed considering the degree of phase change and the flow rate of refrigerant in each pass, the

heat transfer and pressure drop relationship was found to be substantially improved as compared to the conventional designs as described hereinafter.
Turning to FIG. 10-12, there are shown the test results with the condenser of the invention and the conventional condenser as changing the hydraulic diameters of the tube and by-pass passageway, and the number of passes,
FIG. 10 shows trends between the heat transfer efficiency and pressure drop relation in combination with the effective areas of passes. In this case, the condenser had four passes and the
ratio D/D was 0.95.
Referring now to FIG. 10, it is seen from curves C and F sersus D and F for the condenser of the present invention and the conventional condenser, respectively, that when the number of tubes constituting a pass on the refrigerant inlet side (the first uppermost pass) over the number of overall tubes constituting all passes of the condenser is within 40%, with both the present and conventional condensers the heat transfer efficiency diminishes while the pressure drop increases. However, when the ratio of the number of tubes of the pass on the inlet side over the number of overall tubes ranges from 40% to 65%, the condenser of the present invention illustrates improvements of performance in both the heat transfer effieiency and pressure drop over the conventional condenser which also has the by-pass passageways . Moreover, in

testing the condensors with there and five passes, respectively, it was found for the three pass condenser to operate in optimum performance when the ratio of the number of tubes, which means area, of the pass on the inlet side over the number of overall tubes, which maans the entire area of the condenser, ranges from 55% to 65%, and for the five pass condenser when the ratio ranges from 30% to 45%, Therefore, is is confirmed that the degree of phase change in the pass on the inlet side significant1y aff affecs the heat transfer performance, and the desired heat transfer performance occurs when the relation between the flow rate of liquid refregerant to be by-passed and the passes through which the vapor to be condensed flows without by-pass is selected in optimum. Namely, because the vapor introduced into the condenser through the in1et pipe has a relatively larqe volume and thus, the 1argest volume of the vapor is condensed through the pass on the inlet side, when not by-passing the condensed liquid the flow resistance both the pressure drop and flow resistance occur due to the flow ratio difference between the vapor and the liquid. However, when by-passing the liquid refrigerant of relatively large volume, the vapor flows smoothly through the tubes and through even the passes near the lowermost pass without 1arge difference in flow rate as compared with the flow rate in the pass on the inlet side.
My designing the condender with the above conditions, the number of passes can be increased to a certain extent even with the

small hydraulic diameter because both the pressure drop and heat transfer of efficiency improve, on one hand, while when utilizing the large hydraulic diameter of tube, the number of passes can also be increased, which means increment of the length of fluid pahts, without the disadvantageous pressure drop, on the other hand. From these facts, it is construed that with the smae sized condensers, superior performance is obtained in the condenser according to the present invention over the conventional designs irrespective of presence of the by-pass passageways, and in turn, more compact condonsers are provided according to the invention when desgn is made to obtain the same performance.
Turning to FIG. 11 , there is shown the relation between the heat transfer efficiency and the pressure drop with vairations of the hydraulic diameter of the tube in the range of 1 to 1.7 mm. Prior art ] is a conventional condenser without the by-pass passageways while prior art 11 is the condenser with the by-pass passageways formed in prior art 1.
In FIG 11 , the number of passes and the ratios of effective areas of the passes on the inlet side over the number of overall tubes were four and about 30%-40%, respectively, for both prior arts 1 and 11. As can be soon, the pressure drop of prior art 11 with the by-pass passageways is lesser than that of prior art I without the by-pass passageways, while the heat transfer efficiency of prior art 11 deprociates relative to that of prior art 1, and

accordingly, the performance of the condenser with the by-pass passageways is inferior to that of the condenser without the by-pass passageways. However, with the condenser accordinq to the present invention to which the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the flat tube and the raito of the area of the pass on the inlet side over the area of overall passes, 30-60% are applied, when the hydraulic diameter of the tube increases, the heat transfer efficiency of the present invention is superior to those of prior arts 1 and 11 whi1e the pressure drop of the present invention is superior to prior art 1, on one hand, inferior to prior art 11, on the other hand. The reason the pressure drop of the present invention is a little higher than that of prior art 11 is construed as more vapor toqhether with the liquid is by-passed in the condenser of prior art 11 than in the condenser of the present invention because the area of the pass on the inlet side of prior art 11 is smaller than that of the present invention. Namely, FIG. 11 shows that the amounts of the vapor condensed through the pass on the inlet side and the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the tube is related with each other irrespective of the hydraulic diameters used in the normal tubes, and the condenser performance is best when the area of the pass on the inlet side is chosen in the range shown in FIG. 10. Accordingly, the desired heat transfer efficiency and pressure drop

is accordingly when optimizing the raito of the hydraulic diameter of
the by-passageway over the hydraulic diameter of the tube and
selecting the number of of tubes constition the pass on the inlet
side in a qiven prescribed range according to the number of passes. formed in the condenser.
referring now to FIG 12, there are shown the results of tests with variations of the number of passes under the conditions as discribed in relation to FIG 11 in which prior art 1 is a common condenser without the by-pass passageways and prior art 13 with the by-pass passageways but a small area of the pass on the inlet side than that of the present invention. FIG- 12 shows that too many passes accompanies restrictions because increase of the number of passes enhances the heat transfer efficiency but raises the pressure drop. Namely, the heat transfer efficiency increases with rapid rising of the pressure in prior art 1, while in prior art 11, 1 he pressure drop is slow with the inferior heat transfer efficiency to prior art 1, and thus, the same trends are identified as in FIG. 11. On the other hand, with the condenser of the present invention, the heat, transfer efficiency increases but the pressure drop increases slowly, and accordingly, increment of the number of passes to a given extent under the smae conditions accompanies lessor restrictions .
In consideration of the data as shown in FIGS. 9-12 the performance of condenser in aspects of the heat transfer

efficiency and pressure drop is improved by designing in view of three conditions : first, the hydraulic diameter of each tube used in the multiflow type condensor; second, the hydraulic diameter of the by-pass passageways over the hydraulic diameter of the bube ; and finally, the ratio of the number of the tubes constituting the pass on the inlet side over the number of overall tubes constituting the overall passes of the condenser, i.e. the area of the pass ( P1 in FIG.1 ) on the inlet side .
Namely, when the hydraulic diameter of each tube ranges from 1 to 1.7 mm, the ratio of the hydraulic diameter of the by-pass passageway over the hydraulic diameter of the tube, D/D is between 0.28 and 2.25, and the ratio of the pass on the inlet side over the area of the overall pases ranges from 30% to 65%, the performance of parallel flow type condenser is improved as compared to condensers not to fulfi11 the above three conditions, irrespective of presence of the by-pass passageways. For example, optimized performance was observed with the condenser having four passes, and each tube of the ratio of D/D ranging from 0.45 to 1.85, with the ratio of the area of the pass on the inlet side over the area of the overall passes ranging 40% to 55%.
The present invention has been described in an illustrative manner. Many modifications and variations of the present invention are possible in light of the above theachings. Therefore, the spi1it and scope of the invention are to be limited

] , A multi f1ow type condensor f or an automobile air condit i oner
comprising :
a pai r of header pipes disposed in parallel with each other and arranged to have an inlet and an outlet , said header pipes being elliptical in cross section;
a plurality of flat tubes vibes each connected to said header pipes at opposite ends thereof , each of said flat tubes having a plurality of inside fluid paths, a hydraulic diameter of each of said inside fluid paths being in the range of about 1 to 1.7 mm;
a plurality of corrugated fins each disposed between adjacent flat tubes;
at least a pan of baffle disposed in said header pipes one by one;
each of said bafles havinq a projection inserted into a slit provided with each header pipes and dividing each header pipes into a plurality of chambers so that a refrigerant flows through a plurality of passes each defined by a plurality of tubes in zigzag fashion between said inlet and said out let , an out er peripheral surface of each baff1e contacting with an inner peripheral surface of said respective header pipes;

at least one by-pass passageway formed around a posi tion at which the chambers in each header pipe are di vidad by t.he baffle therein to route a vapor- abundant phase of said refrigerant, from an upper chamber to alowor ehamber within t he same header pipes by providing a communication path between the adjacent chambers ;
passageway ever sai d hydraulic diameter of each of sai d in the fluid paths being in the range of about 0.78 to 2.25 and
an area of a pass on the inlet side defined by the chamber on the i nlet side i nt o which sai d ref ri gerant is i ntroduced
through said inlet and formed in one of said header pipes, the
opposed chamber formed the other of said header pipes, and a plurality of tubes extending between the chambers is about 30% to 65% of an ovorall area of all of said passes.
/. The condenser of claim 1 , who? ei n 5;ai d passes arc* three and eot i d area oi said pass on the inlet side is about. bbH- to ^'?V d .sa i (J ovo a.) .1 ar oa of sai d passes ,
sa i d aioa of said pass on the inlet side is about 40* to vo 4 . The condenser of claim ;» , whoj ci n said passes are five ,inb

said area said pass on t he inlet side is about 30% to 40% of said overal1 area of said passes.
5. The condenser of claim 1, wherein said by-pass passageway is formed in cent ra l port ions of said baffles,respectively, by
sfi ^il chi tiq,
(■">. The condenser of o3 aim 1 , wherei n sai d by-pass pass/jfjew,!;/ j s f oj mod in outer peri pher al port 2 ons of said 1M i 1 11 ■:;, respect i vc.l y.
7. The condenser of c] aim ] , wherei n said by-pass passageway i s formed :i n sai d baf fles, respect .i ve.l y, such that a numboi of j >y pass passaqoways pi ogressi ve .1 y .incrofjsos from .said inlet 1 -MI id out] et ,
It , The condenser of c] ai rn 1 , whoj oi n sai d by-pass passageway .i s f ormod i n sai d baf f ] es, re spool i ve] y, such that said r ai i o ni sa i d hydr au.l :i o d.i arnet or of sai d by-pass passageway ovo t sa i d hydrau.l if: diameter of each of said .inside fluid paths progressively ineroasfs wilhin said ratio from said inlet to said outlet .
9, Tho condenser of claim 1 , wherein said by-pass passn
sa id header pipes .
10, The condenser of claim 9, wherein said by-pass passageways one formed such that said ratio of said hydraulic diameter of said by- pass passageway over said hydraulic diameter of each of said inside fluid paths progressive1y increases within said ratio from
said inlet to said outlet.
11 . The condenser of claim 1 , wherein said project i on extends outside each header pipes and is pressed around an outer surface of each of said header pipes by caulking means.
i 12. A multiflow type condenser for an automobile air
conditioner substantially as hereinbefore described with
reference to the accompanying drawings.
Dated this 25th day of September 1998

Documents:

2170-mas-1998-abstract.pdf

2170-mas-1998-claims filed.pdf

2170-mas-1998-claims granted.pdf

2170-mas-1998-correspondnece-others.pdf

2170-mas-1998-correspondnece-po.pdf

2170-mas-1998-description(complete)filed.pdf

2170-mas-1998-description(complete)granted.pdf

2170-mas-1998-form 1.pdf

2170-mas-1998-form 19.pdf

2170-mas-1998-form 26.pdf

2170-mas-1998-form 3.pdf

2170-mas-1998-form 4.pdf

2170-mas-1998-form 6.pdf

2170-mas-1998-other documents.pdf

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Patent Number

212447

Indian Patent Application Number

2170/MAS/1998

PG Journal Number

07/2008

Publication Date

15-Feb-2008

Grant Date

03-Dec-2007

Date of Filing

25-Sep-1998

Name of Patentee

HALLA CLIMATE CONTROL CORP

Applicant Address

1689-1, SHINIL-DONG, TAODOK-GU, TAEJON,

Inventors:

#	Inventor's Name	Inventor's Address
1	YONG GWI AHN	1689-1, SHINIL-DONG, TAEDOK-GU, TAEJON,
2	SANG YUL LEE	1689-1, SHINIL-DONG, TAEDOK-GU, TAEJON,
3	SEUNG HWAN KIM	1689-1, SHINIL-DONG, TAEDOK-GU, TAEJON,
4	SANG OK LEE	1689-1, SHINIL-DONG, TAEDOK-GU, TAEJON,
5	KWANG HEON OH	1689-1, SHINIL-DONG, TAEDOK-GU, TAEJON,
6	YONG HO KIM	1689-1, SHINIL-DONG, TAEDOK-GU, TAEJON,

PCT International Classification Number

F 28 B 1/06

PCT International Application Number

N/A

PCT International Filing date

PCT Conventions:

#	PCT Application Number	Date of Convention	Priority Country
1	97-49276	1997-09-26	Republic of Korea
2	98-38816	1998-09-19	Republic of Korea