Abstract:
A heat exchange tube having a flat shape includes a plurality of fluid paths having a circular cross section and extending in a longitudinal direction of the tube. Each fluid path is parallel to each other fluid path. The tube is dimensioned such that a distance between two adjacent fluid paths is defined as Wt, and a circumferential thickness between a surface of the tube and an outermost fluid path is defined as Ht. The distance Wt and the circumferential thickness Ht have a relationship as 0.42≦Ht/Wt≦0.98.

Description:
CROSS REFERENCE TO RELATED APPLICATION 
     This application is based on Japanese Patent Application No. 2003-146661 filed on May 23, 2003, the disclosure of which is incorporated herein by reference. 
     FIELD OF THE INVENTION 
     The present invention relates to a heat exchange tube having multiple fluid paths. The heat exchange tube is suitably used for a heat exchanger in a vapor-compression refrigerant cycle. 
     BACKGROUND OF THE INVENTION 
     A heat exchanger is used for a vapor-compression refrigerant cycle. Specifically, the heat exchanger is used for an air conditioner in an automotive vehicle. In the air conditioner, the heat exchanger works as a condenser. As shown in  FIGS. 8 and 9 , a multi-flow type heat exchanger  10  is used in the air conditioner. The heat exchanger  10  includes a pair of headers  3 , multiple heat exchange tubes  1 , a fin  4  and a side plate  5 . The headers  3  are disposed along with a vertical direction of the heat exchanger  10 . The heat exchange tubes  1  are disposed in parallel between the headers  3 . Both ends of each heat exchange tube  1  are connected to the headers  3 , respectively. The fin  4  is disposed between the heat exchange tubes  1 . The fin  4  is further disposed outside of the outermost heat exchange tube  1 . The side plate  5  is disposed outside of the outermost fin  4 . 
     A separation member  6  is disposed in the header  3  so that the heat exchange tubes  1  are divided into multiple parts P 1 -P 3 . Refrigerant is introduced into the heat exchanger  10  from an inlet  7  of the header  3  disposed upper side of the header  3 . Then, the refrigerant flows through the parts P 1 -P 3 , respectively. While the refrigerant flows through the parts P 1 -P 3 , heat is exchanged between the refrigerant in the heat exchange tubes  1  and the outside air outside of the heat exchanger  10  so that the refrigerant is condensed and liquefied. Then, the liquefied refrigerant flows out of the heat exchanger  10  from an outlet  8  of the header  3  disposed under the header  3 . The heat exchange tube  1  of the heat exchanger  10  is made of, for example, aluminum. The heat exchange tube  1  is formed by an extrusion method to be flattened. The heat exchange tube  1  includes multiple fluid paths. Each fluid path extends in a longitudinal direction and disposed in parallel in a latitudinal direction, as shown in  FIG. 9 . 
     In general, the refrigerant in the air conditioner is, for example, hydro chloro fluoro carbon (i.e., HCFC), or hydro fluoro carbon (i.e., HFC). It is already decided to prohibit using the HCFC refrigerant by year 2020. This is because the HCFC is one of ozone-layer-destroying materials. Further, the HFC refrigerant is one of greenhouse gases. Therefore, the HFC is also strictly limited from discharging to the atmosphere. Thus, alternative materials of chloro fluoro carbon such as the HCFC refrigerant or the HFC refrigerant is required to develop. Specifically, it is required to develop a new technique using the alternative materials. 
     Recently, carbon dioxide (i.e., CO 2 ) is considered as one of alternative materials. Specifically, the CO 2  refrigerant is used in the vapor-compression refrigerant cycle. The CO 2  gas is one of natural gasses in nature. Therefore, the CO 2  gas does not affects on the global environment substantially compared with the chrolo fluoro carbon. 
     However, when the CO 2  refrigerant is used as the refrigerant in the vapor-compression refrigerant cycle, the CO 2  refrigerant has comparatively high pressure in regular use. This is because the refrigerant cycle becomes a super critical refrigerant cycle because of specific thermodynamic properties of the CO 2  gas. Therefore, for example, the pressure of the CO 2  refrigerant in regular use on a high-pressure side of the refrigerant cycle becomes higher than 10 Mpa. Here, the pressure of the chloro fluoro carbon refrigerant has comparatively low pressure in regular use. The pressure of the chloro fluoro carbon refrigerant is, for example, 3 MPa or 4 MPa. Thus, in a case where the CO 2  refrigerant is used as the refrigerant in the refrigerant cycle, it is required to secure the high mechanical strength of the heat exchange tube. Specifically, the heat exchange tube is required to have the withstand pressure three times or more higher than the pressure in regular use on the high-pressure side. That is, the withstand pressure of the heat exchange tube is required to be about 30 MPa or 40 MPa. 
     A heat exchange tube having high withstand pressure is, for example, disclosed in Japanese Patent No. 3313086 (i.e., Japanese Patent Application Publication No. 2000-356488). A fluid path of the heat exchange tube has a rectangular cross section with a rounding corner. Further, thickness of a sidewall of the heat exchange tube becomes thicker. 
     However, it is preferred that the fluid path has a perfect circular cross section in view of the withstand pressure of the heat exchange tube. Further, it is difficult to define the withstand pressure on the basis of only a ratio between the thickness of the heat exchange tube and the width of the fluid path. This is because the heat exchange tube can be made of one of various materials having high mechanical strength. Each material has a different mechanical strength. Therefore, it is difficult to estimate the withstand pressure of the heat exchange tube. 
     SUMMARY OF THE INVENTION 
     In view of the above-described problem, it is an object of the present invention to provide a heat exchange tube with multiple fluid paths having a perfect circular cross section and having high withstand pressure. 
     A heat exchange tube having a flat shape includes a plurality of fluid paths having a perfect circular cross section and extending in a longitudinal direction of the tube. Each fluid path is parallel together. The tube has a certain dimensions in such a manner that a distance between two adjacent fluid paths is defined as Wt, and a circumferential thickness between a surface of the tube and an outmost fluid path is defined as Ht. The distance Wt and the circumferential thickness Ht have a relationship as 0.42≦Ht/Wt≦0.98. 
     In the above heat exchange tube, the fluid paths have a perfect circular cross section, and the tube has sufficient high withstand pressure. Further, the weight of the tube becomes light. 
     Preferably, the fluid paths are aligned in a line along with a latitudinal direction of the tube. More preferably, the tube includes a circumferential surface having a concavity and a convexity corresponding to the fluid path. 
     Preferably, the fluid paths are aligned in multiple lines along with a latitudinal direction of the tube, and two adjacent fluid paths disposed in two adjacent lines, respectively, are disposed alternately. More preferably, the tube includes an circumferential surface having a concavity and a convexity corresponding to the fluid path. 
     Preferably, the tube is used for a high-pressure side heat exchanger in a vapor-compression refrigerant cycle with CO 2  refrigerant. The fluid path has a diameter defined as Dp, and the tube is made of material having a tensile strength defined as S. The relationship among the distance Wt, the tensile strength S and the diameter Dp is defined as (0.73−0.0036×S)×Dp≦Wt≦(1.69−0.0084×S)×Dp . More preferably, the tensile strength S is in a range between 50 N/mm 2  and 130 N/mm 2 , and wherein the tube is made of aluminum based material. Furthermore preferably, the diameter Dp is in a range between 0.4 mm and 2.0 mm. 
     Preferably, the tube is used for a low-pressure side heat exchanger in a vapor-compression refrigerant cycle with CO 2  refrigerant. The fluid path has a diameter defined as Dp, and the tube is made of material having a tensile strength defined as S. The relationship among the distance Wt, the tensile strength S and the diameter Dp is defined as (0.34−0.0024×S)×Dp+0.06≦Wt≦(0.80−0.0056×S)×Dp+0.14. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The above and other objects, features and advantages of the present invention will become more apparent from the following detailed description made with reference to the accompanying drawings. In the drawings: 
         FIG. 1A  is a schematic perspective view showing a heat exchange tube, and  FIG. 1B  is a partially enlarged cross sectional view showing the heat exchange tube according to a first embodiment of the present invention; 
         FIG. 2  is a graph showing a place where the maximum stress is generated in the heat exchange tube, according to the first embodiment; 
         FIG. 3  is a part of a cross sectional view showing a simulation model of the heat exchange tube for simulating the stress in the heat exchange tube, according to the first embodiment; 
         FIG. 4  is a graph showing a place where the maximum stress is generated in the heat exchange tube disposed on the high-pressure side in a CO 2  refrigerant cycle, according to the first embodiment; 
         FIG. 5  is a graph showing a place where the maximum stress is generated in the heat exchange tube disposed on the low-pressure side in the CO 2  refrigerant cycle, according to the first embodiment; 
         FIG. 6A  is a schematic perspective view showing a heat exchange tube, and  FIG. 6B  is a partially enlarged cross sectional view showing the heat exchange tube according to a second embodiment of the present invention; 
         FIG. 7  is a schematic perspective view showing a heat exchange tube according to a third embodiment of the present invention; 
         FIG. 8  is a plan view showing a multi-flow type heat exchanger according to a prior art; and 
         FIG. 9  is an exploded perspective view showing a heat exchange tube and a header in the heat exchanger according to the prior art. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     First Embodiment 
     A heat exchange tube  1  having multiple fluid paths  2  according to a first embodiment of the present invention is shown in  FIGS. 1A and 1B . The heat exchange tube  1  is suitably used for a heat exchanger in a vapor-compression refrigerant cycle. Specifically, the heat exchange tube  1  is used in a heat exchanger in a vapor-compression refrigerant cycle having comparatively high-pressure refrigerant such as carbon dioxide. The heat exchange tube  1  is used in the heat exchanger such as a multi-flow type heat exchanger or a parallel flow type heat exchanger. The fluid paths  2  of the heat exchange tube  1  flow the refrigerant having high temperature, extend in a longitudinal direction of the tube  1 , have perfect circular cross sections, respectively, and are parallel each other in a latitudinal direction of the tube  1 . The fluid paths are aligned in a line in the tube  1 . 
     The heat exchange tube  1  is made of aluminum having long length and formed by an extrusion method. The heat exchange tube  1  is formed to be flattened, and has the fluid path  2  having a perfect circular cross section. The fluid path  2  extends in the longitudinal direction of the heat exchange tube  1 . Multiple fluid paths  2  are disposed in parallel in the latitudinal direction of the tube  1 . As shown in  FIG. 1B , the width of a separation portion between the fluid paths  2  (i.e., the distance between the fluid paths  2 ) is represented as Wt millimeters. The thickness of the tube  1  is represented as Ht millimeters. The thickness of the tube  1  is disposed outer circumference of the tube  1 , i.e., the thickness is disposed between the fluid path  2  and the circumference of the tube  1 . The diameter of the fluid path  2  is represented as Dp millimeters. The total thickness (i.e., the height) of the tube  1  is represented as H millimeters. The tensile strength of the material composing the tube  1  is S N/mm 2 . 
     The distance Wt is defined as follows. In case of the heat exchanger disposed on the high-pressure side, the optimum distance Wt of the tube  1  disposed on the high-pressure side is defined as:
 
 Wt =(1.21−0.006 ×S )× Dp.  
 
     In case of the heat exchanger disposed on the low-pressure side, the optimum distance Wt of the tube  1  disposed on the low-pressure side is defined as:
 
 Wt =(0.57−0.004 ×S )× Dp+ 0.1.
 
     The optimum relationship between the thickness Ht of the tube  1  and the distance Wt is such that:
 
 Ht:Wt= 0.7:1.0 (i.e.,  Ht/Wt= 0.7).
 
     Here, the total thickness H of the tube  1  is defined as:
 
 H=Dp+ 2 ×Ht.  
 
     The above optimum distances and the optimum relationship are obtained as follows. The stress in the tube  1  having different thicknesses Ht and distances Wt is numerically analyzed. As a result, the thickness Ht and the distance Wt have the relationship shown in  FIG. 2 . In  FIG. 2 , a region A represents the tube  1  having a portion disposed between the fluid paths  2 , the portion in which the maximum stress is generated. That is, the maximum stress is generated in the portion of the tube  1  shown as Wt in  FIG. 1B  (i.e., the portion of the tube  1  is a partition portion). A region B represents the tube  1  having another portion disposed between the fluid path  2  and the circumference of the tube  1 , the other portion in which the maximum stress is generated. That is, the maximum stress is generated in the other portion of the tube  1  shown as Ht in  FIG. 1B  (i.e., the other portion of the tube  1  is a circumferential portion). Thus,  FIG. 2  shows the portion, in which the maximum stress is generated. The stress is generated by inner pressure of the fluid in the tube  1 . 
     In the region A in  FIG. 2 , even if the thickness Ht of the circumferential portion becomes thicker, the maximum stress is generated in the partition portion. Therefore, a crack or a break may be generated from the partition portion. On the other hand, in the region B in  FIG. 2 , even if the distance Wt, i.e., thickness of the partition portion becomes thicker, the maximum stress is generated in the circumferential portion. Therefore, a crack or a break may be generated from the circumferential portion. 
     In view of the above relationship between the thickness Ht and the distance Wt, the tube  1  is designed to have the maximum withstand pressure effectively. Specifically, when the ration of Ht/Wt is set to be an optimum value so that the stress generated in the partition portion is almost equal to the stress generated in the circumferential portion, the tube  1  has the maximum withstand pressure. On the basis of the result shown in  FIG. 2 , the optimum value of the ratio of Ht/Wt is defined as:
 
 Ht:Wt= 0.7:1.0 (i.e.,  Ht/Wt= 0.7).
 
     This optimum value is independent from the diameter Dp of the fluid path  2  and the tensile strength S of the material composing the tube  1 . This is confirmed by the analysis of the stress in the tube  1  having different thicknesses Ht and distances Wt. The distance Wt between the fluid paths  2  and the thickness Ht of the tube  1  are determined with holding the optimum value of the ration of Ht/Wt, so that the tube  1  has a sufficient withstand pressure and becomes light weight. 
     The result of the analysis of the stress is described in detail as follows. In the stress analysis, a quarter part of the tube  1  as a simulating model is assumed, as shown in  FIG. 3 . The parameters of the analysis are the tensile strength S, the diameter Dp, the distance Wt, the thickness Ht, and the inner pressure P.  FIG. 4  shows the result of the stress analysis.  FIG. 4  is similar to  FIG. 2 . In  FIG. 4 , the tube  1  is applied with the inner pressure of 40 MPa. In  FIG. 4 , for example, a solid line a 7  represents the relationship between the thickness Ht and the distance Wt in the tube  1  having the diameter Dp of 2.0 mm and the tensile strength S of 130 N/min, when the inner pressure of 40 MPa is applied to the tube  1 . Specifically, when the thickness Ht and the distance Wt are disposed on a part of the solid line a 7  disposed upside of an optimum ratio line L, the crack or the break is generated from the partition portion disposed between the fluid paths  2 . That is, even if the thickness Ht of the circumferential portion becomes thicker, the crack or the break is generated from the partition portion. On the other hand, when the thickness Ht and the distance Wt are disposed on another part of the solid line a 7  disposed downside of the optimum ratio line L, the crack or the break is generated from the circumferential portion disposed between the fluid path  2  and the circumference of the tube  1 . That is, even if the distance Wt between the fluid paths  2  becomes larger, the crack or the break is generated from the circumferential portion. 
     Specifically, when the distance Wt is equal to or larger than 0.9 mm in a case where the thickness Ht is about 0.63 mm, the crack is generated from the circumferential portion. When the thickness Ht is equal to or larger than 0.63 mm in a case where the distance Wt is about 0. 9 mm, the crack is generated from the partition portion. 
     Therefore, the solid line a 7  represents a limitation line of the withstand pressure. That is, when the tube  1  has the thickness Ht and the distance Wt, which are disposed on the right upper side from the solid line a 7 , the tube  1  can bear the inner pressure of 40 MPa. 
     Thus, the intersection between the part and the other part of the solid line a 7  is obtained. The intersection represents that the thickness Ht is 0.63 mm, and the distance Wt is 0.9 mm. When the tube  1  has the thickness Ht of 0.63 mm and the distance Wt of 0.9 mm, the crack is generated from the partition portion or the circumferential portion, i.e., the withstand pressure of the partition portion is substantially equal to that of the circumferential portion. Each intersection of lines a 1 -a 9  is connected together so that the optimum ratio line L is obtained. Here, the optimum ratio line represents the optimum ratio of Ht/Wt=0.7. As a result, even when the diameter Dp and/or the tensile strength S are changed, the withstand pressure of the partition portion is substantially equal to that of the circumferential portion in a case where the optimum ratio of Ht/Wt is 0.7. 
     Here, in  FIG. 4 , the dotted lines a 1 -a 3  represent the tube  1  having the diameter Dp of 0.4 mm. The dashed lines a 4 -a 6  represent the tube  1  having the diameter Dp of 1.0 mm. The solid lines a 7 -a 9  represent the tube  1  having the diameter Dp of 2.0 mm. Also, in  FIG. 4 , the open circle represents the tube  1  having the tensile strength S of 50 N/mm 2 . The closed square represents the tube  1  having the tensile strength S of 80 N mm 2 . The closed triangle represents the tube  1  having the tensile strength S of 130 N/mM 2 . 
       FIG. 5  shows another result of the stress analysis. In  FIG. 5 , the tube  1  is applied with the inner pressure of 30 MPa. Even in this case, the withstand pressure of the partition portion is substantially equal to that of the circumferential portion in a case where the optimum ratio of Ht/Wt is 0.7. 
     Here, if the thickness Ht becomes larger than the optimum ratio of Ht/Wt=0.7, the weight of the tube  1  becomes larger although the withstand pressure of the tube  1  is not changed. Therefore, the weight saving of the tube  1  is prevented. On the other hand, if the distance Wt becomes larger than the optimum ratio of Ht/Wt=0.7, the weight of the tube  1  becomes larger although the withstand pressure of the tube  1  is not changed. Therefore, the weight saving of the tube  1  is prevented. 
     Next, characteristics of the present invention are described as follows. The actual relationship between the distance Wt and the thickness Ht is set to be:
 
0.42 ≦Ht/Wt≦ 0.98.
 
     In this case, the actual ratio of Ht/Wt is within almost ±40% (i.e., in a range between +40% and −40%) of the optimum ratio of Ht/Wt=0.7. Therefore, the tube  1  becomes light weight and has sufficient high withstand pressure. 
     Preferably, the actual relationship between the distance Wt and the thickness Ht is set to be 0.56≦Ht/Wt≦0.84. In this case, the actual ratio of Ht/Wt is within almost ±20% (i.e., in a range between +20% and −20%) of the optimum ratio of Ht/Wt=0.7. Therefore, the weight of the tube  1  becomes much lighter and the tube  1  has sufficient high withstand pressure. 
     More preferably, the actual relationship between the distance Wt and the thickness Ht is set to be 0.63≦Ht/Wt≦0.77. In this case, the actual ratio of Ht/Wt is within almost +10% (i.e., in a range between ±10% and −10%) of the optimum ratio of Ht/Wt=0.7. 
     The optimum distance Wt of the tube  1  disposed on the high-pressure side heat exchanger defined as Wt=(1.21−0.006×S)×Dp is obtained as follows. When the tube  1  has the thickness Ht and the distance Wt having the optimum ratio of Ht/Wt=0.7, the breaking strength of the tube  1  is determined by both of the diameter Dp and the distance Wt or both of the thickness Ht and the tensile strength S. It is required to have the breaking strength of 40 MPa for the tube  1  disposed in the high-pressure side heat exchanger in the CO 2  refrigerant cycle. In view of the stress analysis shown in  FIG. 4 , the optimum distance Wt is obtained as:
 
 Wt =(1.21−0.006 ×S )× Dp.  
 
     Here, in  FIG. 4 , for example, when the tensile strength S is 50 N/mm 2  and the diameter Dp is 0.4 mm, the minimum distance Wt is 0.364 mm, which is the intersection of the dotted line a 3  in  FIG. 4 . The thickness Ht is obtained by the above formula and the relationship of the optimum ratio of Ht/Wt=0.7. 
     The actual relationship between the distance Wt, the diameter Dp and the tensile strength S in the tube  1  disposed on the high pressure side of the CO 2  refrigerant cycle is set to be:
 
(0.73−0.0036 ×S )× Dp≦Wt ≦(1.69−0.0084 ×S )× Dp.  
 
     In this case, the actual distance Wt is within almost ±40% (i.e., in a range between +40% and −40%) of the optimum distance Wt defined as Wt=(1.21−0.006×S)×Dp. Therefore, the tube  1  becomes light weight and has sufficient high withstand pressure. Specifically, the tube  1  has the sufficient withstand pressure on the high-pressure side of the CO 2  refrigerant cycle. 
     Preferably, the actual relationship between the distance Wt, the diameter Dp and the tensile strength S is set to be (0.97−0.0048×S)×Dp≦Wt≦(1.45−0.0072×S)×Dp. In this case, the actual distance Wt is within almost ±20% (i.e., in a range between ±20% and −20%) of the optimum distance Wt of Wt=(1.21−0.006×S)×Dp . Therefore, the weight of the tube  1  becomes much lighter and the tube  1  has sufficient high withstand pressure. 
     More preferably, the actual relationship between the distance Wt, the diameter Dp and the tensile strength S is set to be (1.09−0.0054×S)×Dp≦Wt≦(1.33−0.0066×S)×Dp. In this case, the actual distance Wt is within almost ±10% (i.e., in a range between +10% and −10%) of the optimum distance Wt of Wt=(1.21−0.006×S)×Dp. 
     The optimum distance Wt of the tube  1  disposed on the low-pressure side heat exchanger defined as Wt=(0.57−0.004×S)×Dp+0.1 is obtained as follows. When the tube  1  has the thickness Ht and the distance Wt having the optimum ratio of Ht/Wt=0.7, the breaking strength of the tube  1  is determined by both of the diameter Dp and the distance Wt or both of the thickness Ht and the tensile strength S. It is required to have the breaking strength of 30 MPa for the tube  1  disposed in the low-pressure side heat exchanger in the CO 2  refrigerant cycle. In view of the stress analysis shown in  FIG. 5 , the optimum distance Wt is obtained as:
 
 Wt =(0.57−0.004 ×S )× Dp+ 0.1.
 
     Here, in  FIG. 5 , for example, when the tensile strength S is 50 N/mm 2  and the diameter Dp is 0.4 mm, the minimum distance Wt is 0.248 mm, which is the intersection of the dotted line b 3  in  FIG. 5 . The thickness Ht is obtained by the above formula and the relationship of the optimum ratio of Ht/Wt=0.7. 
     The actual relationship between the distance Wt, the diameter Dp and the tensile strength S in the tube  1  disposed on the low-pressure side of the CO 2  refrigerant cycle is set to be:
 
(0.34−0.0024 ×S )× Dp+ 0.06 ≦Wt ≦(0.80−0.0056 ×S )× Dp+ 0.14.
 
     In this case, the actual distance Wt is within almost ±40% (i.e., in a range between +40% and −40%) of the optimum distance Wt defined as Wt=(0.57−0.004×S)×Dp+0.1. Therefore, the tube  1  becomes light weight and has sufficient high withstand pressure. Specifically, the tube  1  has the sufficient withstand pressure on the low-pressure side of the CO 2  refrigerant cycle. 
     Preferably, the actual relationship between the distance Wt, the diameter Dp and the tensile strength S is set to be (0.46−0.0032×S)×Dp+0.08 ≦Wt≦ (0.68−0.0048×S)×Dp+0.12. In this case, the actual distance Wt is within almost ±20% (i.e., in a range between +20% and −20%) of the optimum distance Wt of Wt=(0.57−0.004×S)×Dp+0.1. Therefore, the weight of the tube  1  becomes much lighter and the tube  1  has sufficient high withstand pressure. 
     More preferably, the actual relationship between the distance Wt, the diameter Dp and the tensile strength S is set to be (0.51−0.0036×S)×Dp+0.09≦Wt≦(0.63−0.0044×S)×Dp+0.11. In this case, the actual distance Wt is within almost ±10% (i.e., in a range between +10% and −10%) of the optimum distance Wt of Wt=(0.57−0.004×S)×Dp+0.1. 
     Here, when the tube  1  is actually designed, it is required to add additional thickness of the tube  1  for compensating a manufacturing tolerance and/or for increasing the withstand pressure so that the tube  1  has sufficient withstand pressure even if the tube  1  would be corroded. The additional thickness of the tube  1  is added on the calculated thickness having the minimum withstand pressure. In general, the additional thickness of the tube  1  is in a range between +0.05 mm and +0.25 mm. Specifically, the amended thickness Ht′ and the amended distance Wt′ are defined as:
 
 Ht+ 0.05 ≦Ht′≦Ht+ 0.25, and
 
 Wt+ 0.05 ≦Wt′≦Wt+ 0.25.
 
     Here, the optimum ratio of ratio of Ht/Wt is 0.7. Therefore, summarizing the above relations of the amended distance Wt′ and the amended thickness Ht′, the following relationship is obtained as:
 
0.7×( Wt′− 0.25)+0.05 ≦Ht′≦ 0.7×( Wt′− 0.05)+0.25.
 
     Therefore, the amended ratio of Ht′/Wt′ is defined as:
 
0.7−0.125 /Wt′≦Ht/Wt′≦ 0.7+0.215 /Wt′.  
 
     For example, when the distance Wt′ is 1 mm, the ratio of Ht′/Wt′ is 0.575≦Ht′/Wt′≦0.915. 
     The tube  1  is made of aluminum-based material having the tensile strength S in a range between 50 N/mm 2  and 130 N/mm 2 . The diameter Dp of the fluid path  2  is set in a range between 0.4 mm and 2.0 mm. When the tube  1  has the above tensile strength S and the fluid path  2 , the tube  1  has a sufficient withstand strength of the pressure in the CO 2  refrigerant cycle. 
     In the first embodiment, the distance Wt, the thickness Ht, the diameter Dp, the tensile strength S and the total thickness H are determined into certain values, or when the cross section of the tube  1  is determined to have a certain cross section, the tube  1  becomes light weight and has sufficient high withstand pressure by utilizing the above relationship. 
     Thus, the heat exchange tube  1  with multiple fluid paths  2  having a perfect circular cross section has high withstand pressure. Further, the weight of the tube  1  becomes light. 
     Second Embodiment 
     Another heat exchange tube  11  according to a second embodiment of the present invention is shown in  FIGS. 6A and 6B . The tube  11  has multiple fluid paths  2  aligned in a thickness direction (i.e., a height direction) of the tube  11 . The neighboring two lines of the fluid paths  2  adjacent in the thickness direction are disposed alternately in the latitudinal direction of the tube  11 . Thus, the formability of the tube  11  is improved. Further, when the withstand pressure of the tube  11  is constant, the cross section of the fluid path  2  can become larger although the total cross section of the tube  11  becomes minimum. Thus, the tube  11  has minimum dimensions, light weight, high performance and low manufacturing cost. 
     Third Embodiment 
     Further another heat exchange tube  21  according to a third embodiment of the present invention is shown in  FIG. 7 . The circumference of the tube  21  is formed to have a concavity and a convexity in accordance with the fluid path  2 . Thus, the weight of the tube  21  is much reduced without decreasing the withstand pressure. That is, the material composing the tube  21  is much reduced. 
     Such changes and modifications are to be understood as being within the scope of the present invention as defined by the appended claims.