US20120097380A1

US20120097380A1 - Heat exchanger

Info

Publication number: US20120097380A1
Application number: US13/318,040
Authority: US
Inventors: Andreas Richard Hilgert; Peter Tobias Klug; Thomas Zenon Zakrwski; Leonid Walter
Original assignee: Eaton Fluid Power GmbH
Current assignee: Eaton Fluid Power GmbH
Priority date: 2009-04-30
Filing date: 2010-04-30
Publication date: 2012-04-26
Also published as: EP2425193B1; CN102460054A; PL2425193T3; GB0909221D0; EP2425193A2; WO2010124871A2; HUE034718T2; ES2643324T3; CN102460054B; WO2010124871A3; SI2425193T1

Abstract

A heat exchanger comprising an outer conduit and an inner conduit, arranged inside of and along the longitudinal axis of the outer conduit, the inner conduit and the outer conduit being arranged to form a fluid flow channel between the inner surface of the outer conduit and the outer surface of the inner conduit, the fluid flow channel having a cross sectional form, in a plane substantially perpendicular to the longitudinal axis of the outer conduit, that is elongate being substantially greater in the circumferential direction of the outer conduit than in the radial direction of the outer conduit.

Description

TECHNICAL FIELD

The invention relates generally to the technical field of heat exchangers and particularly, but not exclusively, to internal heat exchangers and more particularly those used for use air-conditioning systems for automotive applications.

BACKGROUND

Air-conditioning systems of motor vehicles, for example, are frequently equipped with a so-called internal heat exchanger. Such heat exchangers may be used to increase the operating efficiency of the system by pre-heating the refrigerant supplied to the suction side of a compressor of the air-conditioning system and at the same time cooling the refrigerant (liquid side) being conveyed to an expansion device. One example of an inner heat exchanger is disclosed in DE10 2006 017 816 B4. This document discloses a single piece extruded aluminium heat exchanger element. In this one extruded profile channels are formed for conveying both liquid side and suction side refrigerant. Whilst, extruded heat exchanger elements of this type offer high levels of heat exchange between the suction and the liquid sides, they suffer from certain drawbacks: they require machining and/or cleaning before they can be used; welding or brazing must be used in order to connect the suction line to the profile; and, the geometry of the heat exchanger is fixed by the extrusion tool, meaning that new tools must be developed for a new applications requiring different extrusion profiles.
In order to achieve a desired heat transfer between the suction side and the liquid side the heat exchanger must have a given heat exchange area. Sometimes, space is at a premium, for example in automotive applications. In such cases it is desirable to be able to use heat exchangers of reduced outer dimensions. This often means that it is required to form or bend the heat exchanger as a U-shaped pipe or into other shapes so that it may be installed in a given space. This in turn requires that the heat exchanger pipe be designed in a sufficiently bendable manner so that it may be deformed without collapsing its fluid conveyance channels. Moreover, it may also mean that the outer diameter of the heat exchanger is limited or constrained.
In view of such design requirement it would therefore be desirable to provide a heat exchanger that overcomes at least some of the above mentioned problems.

SUMMARY

According to the present invention there is provided a heat exchanger and a method of manufacturing of a heat exchanger as defined in the appended claims.

BRIEF DESCRIPTION OF THE DRAWINGS

The above and other aspects, features and advantages of the invention will be apparent from the following detailed description of illustrative embodiments which is to be read in connection with the accompanying drawings, in which:

FIG. 1 is a schematic diagram of an air conditioning system for an automotive application comprising an internal heat exchanger;

FIG. 2 shows a schematic illustration of the internal heat exchanger shown in FIG. 1 in a U-shaped configuration;

FIG. 3 a shows a perspective view of an internal heat exchanger according to a first embodiment of the invention in its assembled state but prior to being bent into a U-shaped configuration;

FIG. 3 b shows a photograph of the exterior of an internal heat exchanger according to the first embodiment;

FIG. 3 c shows a photograph of the exterior of a section of the deformed portion of the inner tube of the internal heat exchanger according to the first embodiment;

FIG. 3 d shows a schematic illustration of the exterior of a section of the deformed portion of the inner tube of the internal heat exchanger according to the first embodiment, more clearly showing its helical structure;

FIG. 4 is an image schematically illustrating part of the inner tube of the internal heat exchanger according to a first embodiment, which illustrates one exemplary method of creating a helical structure in a portion of the inner tube;

FIG. 5 a to FIG. 5 c show cross sectional views of the internal heat exchanger according to the first embodiment, illustrating exemplary alternative profiles for the internal heat exchanger inner tube;

FIG. 6 is a schematic illustration of the flow of refrigerant in the internal heat exchanger of the first embodiment;

FIG. 7 shows part of an image of FIG. 4, showing how parameters of the inner tube may be varied to achieve different performance characteristics of the internal heat exchanger of the first embodiment;

FIG. 8 a shows a perspective view of an internal heat exchanger according to a second embodiment in its assembled state but prior to being bent into its final U-shaped configuration;

FIG. 8 b shows a cross sectional view of the internal heat exchanger according to the second embodiment;

FIG. 9 shows an alternative design of the outer tube of the internal heat exchanger according to the second embodiment, with spiral routed refrigerant fluid flow channels.

DETAILED DESCRIPTION

Referring now to the drawings, several embodiments of the present invention are shown in detail. The drawings are not necessarily to scale and certain features may be exaggerated to better illustrate and explain the present invention. Further, the embodiments set forth herein are not intended to be exhaustive or otherwise limit or restrict the invention to the precise configurations shown in the drawings and disclosed in the following detailed description.
Referring to FIG. 1 an air conditioning system 1 suitable for use in a motor vehicle is schematically illustrated. The air conditioning system 1 includes a compressor 2, which may be driven, for example, by the engine of the vehicle or by a separate electric motor or the like. The compressor 2 has an inlet 4, connected to a low pressure line 21, via which where the compressor 2 takes in refrigerant, or coolant, at low pressure. The compressor 2 also has an outlet 3, via which pressurized refrigerant is output, into a high pressure line 5. The high pressure line 5 leads to a cooling device 6 where the compressed and thus heated refrigerant is cooled and condensed. Therefore, the cooling device 6 is also referred to as a condenser. In this example, the refrigerant used is R-134a that works at low pressure.
At an outlet 7 of the cooling device, the refrigerant is discharged to another high pressure line 8 that leads to a high-pressure inlet 9 of an internal heat exchanger 11. The internal heat exchanger 11 has a high-pressure outlet 12 that is in turn connected to an expansion valve 15 via a high pressure line 14. The expansion valve 15 relaxes the refrigerant that is introduced into an evaporator 16. The refrigerant evaporates in the evaporator 16 and, as a result, absorbs thermal energy from the environment; in this example, cooling the air supplied to the interior of the motor vehicle. The resultant refrigerant vapor is then transported from the evaporator 16, via a low-pressure line 17, to the low-pressure inlet 18 of the internal heat exchanger 11. This refrigerant vapor flows through the internal heat exchanger 11 in a counter-current direction to the refrigerant that is being fed through the high-pressure inlet 9. In so doing, the refrigerant vapor cools the pressurized refrigerant, thus itself becoming heated. The refrigerant vapor is discharged, having been heated, at the low-pressure outlet 19 of the internal heat exchanger 11. It is then conducted, via a low-pressure line 21, to the inlet 4 of the compressor 2.
The internal heat exchanger 11 allows the temperature of the refrigerant flowing to the compressor 2 to be increased, which in turn increases the temperature of the refrigerant at the outlet 3 of the compressor. Therefore, the cooling device 6 releases a greater amount of thermal energy. At the same time, the internal heat exchanger 11 lowers the temperature of the refrigerant fed to the evaporator 16, thus providing an improved heat transfer between the evaporator 16 and ambient air. In this manner, the internal heat exchanger 11 may be used to increase the efficiency of the air conditioning system.
FIG. 2 shows a further schematic illustration of the internal heat exchanger 11. In this example, it is shown as a U-shaped bent pipe 22. It will be appreciated that the exact shape of the heat exchanger will depend upon its application. However, in certain applications, but not all, bending of the heat exchanger 11 is required. Where it is required, the coaxial tube should be able to be bent sufficiently without causing the fluid flow channels or conduits to collapse or break. The bent pipe 22 has two legs 23, 24, that are bent away from each other at their upper ends.
The high-pressure inlet 9 and the high-pressure outlet 12 are in fluid connection with the remainder of the system 1 at position 26 a. The low-pressure inlet 18 and the low-pressure outlet 19 of the internal heat exchanger 11 are in fluid connection with the remainder of the system 1 at position 26 b. As can be seen from the figure, positions 26 a and 26 b are located at or relatively close to the terminations of at the upper ends of the bent pipe 22.
Referring now to FIGS. 3 a-3 d, the structure of the internal heat exchanger 11 will be described in more detail. FIG. 3 a shows a perspective view of the internal heat exchanger 11 of a first embodiment in its assembled state but prior to being bent into its final U-shaped configuration. As can be seen from the figure, the internal heat exchanger 11 includes an outer tube 30, and inner tube 32, of which end portions 32 a and 32 b are visible from this figure Both the outer tube 30 and inner tube 32 being designed as refrigerant conduits. The inner tube 32 is located inside and runs the entire length of the outer tube 30. The internal and external diameters of the outer tube 30 are 18 mm and 20 mm, respectively. The internal and external diameters of the parts of the inner tube 32 that extend beyond the outer tube 30 and can be seen in the figure are 12 mm and 15 mm, respectively. It will be understood that the dimensions of the outer tube 30 and inner tube 32 are selected for a given application and will therefore change in dependence upon application. The inner diameter of the outer tube 30 may from range 9-19 mm for automotive or car applications, 20-39 mm for bus applications and, 23-50 mm for train applications. In one example having R-134a as the refrigerant, the outer tube is 24 mm outer diameter with a 20 mm inner diameter. The starting material, or base tube, for the inner tube is 18 mm outer diameter with an inner diameter of 15 mm.
Also shown in the figure are the high-pressure inlet 9 and the high-pressure outlet 12 of the internal heat exchanger 11. Each of these is connected to a suitable orifice in the outer tube 30 using a conventional process such as welding or brazing. The weld points are referenced 34 in the figure. In this manner, a fluid connection is formed between the high-pressure inlet 9 and the high-pressure outlet 12 via the outer tube 30. The connection orifices may be machined, or otherwise manufactured using any convenient process. In this manner, the outer tube 30 may be used as a connection sleeve which allows the system costs to be reduced. The extreme end points 36 of the outer tube 30 are joined to inner tube 32 to ensure that the joint is effectively sealed against leakage of the refrigerant. Again a conventional process may be used; for example o-rings, crimping and or welding or brazing. FIG. 3 b shows a photograph of an example of an internal heat exchanger 11 similar to that shown in FIG. 3 a
In the figure the inner tube 32 has end portions 32 a and 32 b that are circular. These respectively form the low-pressure inlet 18 and the low-pressure outlet 19 of the internal heat exchanger 11. In this example, the end portions 32 a and 32 b are unmodified base tube material. Therefore end portions 32 a and 32 b may be configured to be the required lengths to provide the function of low pressure tubes 21 and 17, shown in FIG. 1. This in turn means that no suction side connection tubes are needed; thus obviating the need for costly connection processes, such as welding and eliminating the risk of refrigerant leakage at such connection points.
Between the end portions 32 a and 32 b of the inner tube 32 is a central portion 32 c that has been deformed into a helical shape along its longitudinal axis. A photograph of the exterior of a section of the deformed portion 32 c of the inner tube of the internal heat exchanger 11 according to the first embodiment is shown in FIG. 3 c. The central portion 32 c may be deformed using any convenient deforming procedure. In the present example it is deformed through a repeated clamping process. However, other deforming processes or apparatus, such as a press or hammer, may be used. In this example, the clamping process is implemented using shaped opposing clamping surfaces to achieve the desired exterior profile of the portion 38 b. The marks 38 a left in the outer surface of the deformed portion 32 c by the action of the clamping process may be seen in FIG. 3 c. Furthermore, it can be seen from FIG. 3 c that the deformed portion 32 c has a helical profile. This helical profile can be more clearly seen from the schematic illustration of a section of portion 32 c illustrated in FIG. 3 d.
Referring to FIG. 4 the method of manufacturing elliptical helix of central portion 32 c, according to this example, will now be described. FIG. 4 shows an image schematically illustrating a part of the inner tube 32, including part of central portion 32 c, arranged about its longitudinal axis 42. As can be seen from the figure, the left hand end 32 a of the inner tube 32 is not deformed and is circular is cross section. Adjacent the left hand end 32 a end of the inner tube 32 is portion 44 a that has been deformed to an approximate elliptical shape of predetermined dimensions. These dimensions may be controlled using the parameters of the deforming process; for example the linear extent of the clamping operation and the shape, dimensions and material properties of the clamping surfaces.
In the figure, the major axis 46 a of the elliptical portion 44 a is shown orientated vertically. When the clamp is removed from portion 44 a of the inner tube 32, the inner tube 32 is advanced a fixed predetermined distance along its longitudinal axis 42 to bring the portion 44 b of the tube adjacent the clamping surfaces and the inner tube 32 is rotated by a fixed angle in a given direction about its longitudinal axis; in this example 45 degrees. The clamping operation is then repeated. This process is then repeated along the desired length of central portion 32 c of the inner tube 32, as is illustrated by deformed portions 44 b-44 f. In this manner an approximate helical structure of approximately fixed helical pitch and approximately constant elliptical cross section may be formed. With the exception of its helical form, the central portion 32 c of the inner tube 32 is free or substantially free of projections and is relatively smooth in both its circumferential direction and its longitudinal direction. The inventors have found that this process of manufacture may be largely automated by using a bending machine set to zero bend radius. Thus, the creation of the helical structure of the central portion 32 c of the inner tube 32 may be a relatively rapid and inexpensive process.
Once the inner tube 32 is formed, it is assembled with the outer tube 30, by inserting the inner tube 32 inside the outer tube 30. The fit between the inner tube 32 and the outer tube 30 may be any convenient fit, such as a loose fit or a slight interference fit. Thus, inner tube 32 and the outer tube 30 may be assembled by hand or be automated. The welding or braising, including crimping if this is required, of the extreme end points 36 of the outer tube 30 to inner tube 32 may then be carried out. This may be done in the region where the non-deformed end sections 32 a and 32 b of the inner tube 32 transition into the adjacent deformed portion 32 c.
FIG. 5 c shows a cross sectional view, in the direction of arrows A-A shown in FIG. 3 a, of the internal heat exchanger 11, and illustrates the inner tube 32 and the outer tube 30 once assembled. As can be seen from the figure, the inner tube 32 forms an approximate ellipse, the major axis of which is approximately equal to the internal diameter of the outer tube 30; i.e. 18 mm. It will be understood that the cross sectional profile of the inner tube 32 could be varied either to meet heat exchange requirements or in order to meet manufacturing requirements. For example as an ellipse, as is illustrated in FIG. 5 a could be used. Other examples could include a triangular or quadrilateral shape, such as an approximate square as is illustrated in FIG. 5 b could also be used. Indeed, other cross sectional profiles may be contemplated, which have increased numbers of sides.
In this example shown in FIG. 5 c, the inner tube 32 contacts the inner surface of outer tube 30 at points 56 a and 56 b, thus forming two substantially line contacts between the outer surface of the inner tube 32 and the inner surface of outer tube 30 which run the entire length of the helical structure of the central portion 32 c of the inner tube 32. In this manner, two refrigerant fluid flow channels 52 a and 52 b are formed between the outer surface of the inner tube 32 and the inner surface of outer tube 30. The fluid flow channels 52 a and 52 b carry liquid side refrigerant. In some embodiments a certain degree of fluid connection between the fluid flow channels 52 a and 52 b may be permitted. The extent of this permitted fluid connection may be dependent upon the application. A third refrigerant fluid flow channel 50 lies on the inside of the inner tube 32. The third refrigerant fluid flow channel 50 carries refrigerant supplied to the suction side of the compressor. The three refrigerant fluid flow channels run substantially the entire length of the helical structure of the central portion 32 c of the inner tube 32.
The third refrigerant fluid flow channel 50 has a cross sectional area which is substantially equal to, or is only marginally reduced relative to the cross sectional area of the base circular tube from which it is formed, and from which the remainder of the suction side, low pressure lines of the air conditioning system 1, are made. This means that the pressure drop caused per unit length of the fluid flow channel 50 is substantially the same as, or not significantly increased relative to, that of the base circular tube from which it is formed, such as low pressure line 21. By avoiding significant pressure loss on the suction side of the internal heat exchanger 11, a considerable loss in the efficiency of the air conditioning system 1 may be avoided, especially in systems operating at lower pressures.
In addition, the inventors have surprisingly discovered that the creation of the helical structure of the central portion 32 c of the inner tube 32 does not cause a significant or measurable drop in pressure in the fluid flow channel 50 relative to a correspondingly profiled tube with no helical structure. The surprising lack of pressure drop in the suction side of the internal heat exchanger 11 of the present embodiment may strongly contribute to the efficiency of the air conditioning system 1.
Whilst in applications for which the internal heat exchanger 11 of the present embodiment is designed benefit from no significant drop per unit length in pressure in the fluid flow channel 50 relative to a correspondingly profiled tube with no helical structure, it will be appreciated that in other applications of the invention a greater pressure drop may be permitted. This may be for example, 2%, 5% or 7% increase relative to a correspondingly profiled tube with no helical structure. However, in some embodiments for certain applications, the suction side pressure drop per unit length of the internal heat exchanger 11 may be up to 30% higher than that of the normal suction side line. In other embodiments this figure may be 10% or 20%.
It will be appreciated that in certain known heat exchangers, in which the design causes such a pressure drop, it may not be easy to remedy. One reason for this is that the technical characteristics of the low pressure fluid flow channel of the heat exchanger may not be easily changed to overcome this problem. For example, it may not be possible to change the cross sectional area of the channel due to space constraints or bending constraints. Additionally, this may not be possible due to the fact that manufacturing costs may be unduly increased due to increased operations being required. Furthermore, it may not be possible to change the internal geometry or flow characteristics of the low pressure fluid flow channel since this may adversely affect the heat exchanging characteristics of the device.
As can be see from FIG. 5 c, the area across which heat may be exchanged between fluid flow channel 50 and each of fluid flow channels 52 a and 52 b is large, being approximately equivalent to half of the external area of the inner tube 32. Moreover, due to the cross sectional shape of the refrigerant fluid flow channels 52 a and 52 b the efficiency of heat exchange between the flow channel 50 and each of fluid flow channels 52 a and 52 b is increased. The fluid flow channels 52 a and 52 b are approximately crescent shaped, having a relatively small height or thickness in the radial direction and a relatively high length of contact with the external circumference of the inner tube 32. This length of contact is illustrated, in the case of fluid flow channels 52 a by line 58 in the figure. It will be appreciated that this line of contact provides a convex heat transfer surface (the external surface of the inner tube 32) against which the fluid in fluid flow channels 52 a and 52 b flows; and thereby a large and efficient heat exchange surface over the length of the fluid flow channels 52 a and 52 b.
FIG. 6 illustrates the flow of refrigerant in the internal heat exchanger 11 according to the present embodiment. The refrigerant flowing in refrigerant fluid flow channel 50 is referenced 60 and the refrigerant flowing in refrigerant fluid flow channels 52 a and 52 b is referenced 62 a and 62 b, respectively. As can be seen from the figure, in this example the refrigerant flowing in refrigerant fluid flow channels 52 a and 52 b follows a helical path along the internal heat exchanger 11 and completes three complete cycles around the fluid in fluid flow channel 50.
It will be understood that the heat exchange characteristics required for a different applications will vary. Accordingly, the heat transfer surface of the present embodiment may be varied. Clearly, the exterior dimensions, such as length and diameter, of the internal heat exchanger may be varied where space permits. Where this is not possible or not desired, parameters of the inner tube 32 may be varied as is illustrated in FIG. 7. FIG. 7 illustrates part of the image of FIG. 4 illustrating several deformed portions 44 of the inner tube 32; where:

“a”=width of base form, determined by the height of the base form if the cross sectional area is equal to the base, or starting, material tube;
“b”=height of base form:
“c”=depth of the of base form along the longitudinal axis of the inner tube 32
“d”=distance between two deformations
“e”=angle between symmetry axis of two deformations
“f”=length of straight portion of the base form, which depends upon “a” and “b” and is zero if the form is elliptical.

The heat transfer surface, the flow velocity and therefore the heat transfer may be adjusted by modifying the geometry of the inner tube 32. The parameters “a”, “b” and “f” determine the cross section of the liquid flow channels 52 a and 52 b and therefore the flow velocity and the heat transfer coefficient. The parameters “c” and “e” determine heat exchange, or contact length and therefore the liquid side heat transfer surface. In general: (i) the efficiency of the internal heat exchanger 11 may be increased by decreasing “c” and “e”; i.e. by increasing the number of deformations per unit length of the inner tube 32 and decreasing the slope of the helix; this may be in the range of 20 to 45 degrees for example;

(ii) the efficiency of the internal heat exchanger 11 may be decreased by increasing “c” and “e”; i.e. by decreasing the number of deformations per unit length of the inner tube 32 and increasing the slope of the helix; this may be in the range of 45 to 90 degrees for example.

It will be appreciated that if the internal heat exchanger 11 is to be formed as U-shaped pipe or other shape, the internal heat exchanger 11 should have sufficient bending stability. The bending stability of the internal heat exchanger 11 may be increased by decreasing the value of parameter “f”.
Referring now to FIG. 8, a second embodiment of the internal heat exchanger 11 for use in the air conditioning system 1 will now be described. Structures and functions of the second embodiment that are similar to corresponding structures and functions of the first embodiment will not be described further. The same reference numerals will be used to refer to structures and functions of the second embodiment as were used to refer to corresponding structures and functions of the first embodiment.
FIG. 8 a shows a perspective view of the internal heat exchanger 11 of the second embodiment in its assembled state but prior to being bent into its final U-shaped configuration. In this embodiment, the inner tube 32 is a plain tube of unmodified base tube material. This may be the same base tube material as was described with respect to the first embodiment. The outer tube 30 however is formed or deformed to provide a predetermined cross sectional profile having longitudinal ridges 82, as is described in more detail below. The outer tube 30 may be formed in any convenient manner; for example extrusion or through a rolling or other deformation process. In the case of the rolling process, circular cross section tube material may be used as the starting material. Angularly spaced, longitudinal grooves may be created in the exterior surface of the tube, leaving ridges 82 between the grooves. Other aspects of the internal heat exchanger 11, including the high-pressure inlet 9 and the high-pressure outlet 12 and their attachment to the outer tube may be the same as described with referenced to the first embodiment.
Referring to FIG. 8 b, the form of the internal heat exchanger 11 of the second embodiment will now be described in more detail. FIG. 8 b shows a cross sectional view, in the direction of arrows A-A in FIG. 8 a, of the internal heat exchanger 11. As can be seen from the figure, the inner tube 32 has a circular cross section. The internal volume of the inner tube 32 forms a fluid flow channel corresponding to fluid flow channel 50 of the first embodiment. The outer tube 30 has a substantially circular cross sectional profile in which 8 ridges 82 a-82 h are arranged. The ridges 82 a-82 h run parallel to the longitudinal axis of outer tube 30 and are arranged at regular angular spacing about the its longitudinal axis. The outside diameter of the inner tube 32 is approximately equal to the minimum internal diameter of the formed outer tube 30; i.e. 18 mm. In the figure, the inner tube 32 contacts the inner surface of outer tube 30 at points 80 a to 80 h. Between the points 80 a to 80 h, the ridges 82 a-82 h in the outer tube 30 form spaces between the inner tube 32 and the outer tube 30. The spaces run along the length of the internal heat exchanger 11 and form refrigerant fluid flow channels 84 a-84 h corresponding to refrigerant fluid flow channels 52 a and 52 b described with reference to the first embodiment.
As is the case with the first embodiment, the fluid flow channels 84 a-84 h of the second embodiment are approximately crescent shaped, having a relatively small height in the radial direction and a relatively high length of contact with the external circumference of the inner tube 32. This length is shown in an illustrative manner by line 86 a in the figure, which lies along the outer circumference (on the outer surface) of inner tube 32 between points 80 a and 80 b. It will be appreciated that this line of contact provides a convex heat transfer surface (the external surface of the inner tube 32) against which the fluid in fluid flow channels 84 a-84 h flows; and thereby a large and efficient heat exchange surface over the length of the fluid flow channels 84 a-84 h.
The number, size and geometry of the refrigerant fluid flow channels 84 a-84 h determines the heat transfer characteristics of the internal heat exchanger 11. These factors also influence the bending properties of the internal heat exchanger 11. In general:

(i) as the cross sectional area of the refrigerant fluid flow channels 84 a-84 h decreases and the number of the refrigerant fluid flow channels 84 a-84 h increases the heat transfer coefficient and bending ability of the internal heat exchanger 11 increases;
(ii) as the cross sectional area of the refrigerant fluid flow channels 84 a-84 h increases and the number of the refrigerant fluid flow channels 84 a-84 h decreases, the heat transfer coefficient and bending ability of the internal heat exchanger 11 decreases.

In practice, the number of the refrigerant fluid flow channels 84 a-84 h as well as the size of the channels is limited by the geometry, the physical properties of the starting material and the production process. In certain situations, it may be difficult to reach the desired heat transfer capacity of the internal heat exchanger 11 for a given application. Limitations on the number of and cross sectional area of the fluid flow channels 84 a-84 h may be experienced due to manufacturing constraints. Also, the total length of the internal heat exchanger 11 may be fixed. In this situation, the contact length, and therefore area, can be enlarged by routing the fluid flow channels 84 a-84 h around the central of the internal heat exchanger 11. This may be achieved creating an outer tube 30 with a suitable profile. This may be through extrusion or through a rolling or other deformation process. An internal heat exchanger 11 according to the second embodiment with spiral routed refrigerant fluid flow channels 84 a-84 h, created by a formed or deformed outer tube 30 is illustrated in FIG. 9.
It will be appreciated that internal heat exchanger 11 shown in FIG. 9, like that shown in FIG. 8, benefits from a convex heat transfer surface (the external surface of the inner tube 32) and a large and efficient heat exchange surface over the length of the fluid flow channels 84 a-84 h. It will also be appreciated that the pitch or angle of the spiral may be selected to obtain a given contact length, and thus area, so that the internal heat exchanger 11 has a desired heat exchange performance.
Because of the fact that the heat transfer capacity is a function of heat transfer coefficient and heat transfer surface (contact area), it is possible to adjust the desired capacity over the surface (by keeping the heat transfer coefficient as high as possible). The contact area or contact time can be adjusted over the length of the fluid flow channels 84 a-84 h.
It will be understood that the above described embodiments give rise to certain advantages. The contours of the inner tube 32 can be placed anywhere along, or even along only a part of, the length of the inner tube 32. Moreover, heat transfer may be adjusted by changing the geometry of the interface between the inner tube 32 and the outer tube 30, and this may be done without significantly changing the forming tool, such as a clamp, or process used. This provides considerable flexibility in terms or manufacturing. Heat exchanger applications with different performance criteria may be achieved without having to significantly modify the manufacturing process or tooling. As the inner tube may be made from standard tubing material, it is low cost. No expensive extrusions are required and no suction side connection tubes are needed, which may help to ensure that manufacturing is simplified and reliability of the system is increased. Bending flexibility may be adjusted by altering the geometry of the deformed tube. The outer tube 30 may be used as a connection sleeve which further allows the system costs to be reduced. Despite the fact that that low-pressure channel may be particularly large, reducing the tendency for suction side pressure drop, a relatively small outside diameter may be achieved.
It will also be understood that various changes may be made to the above described embodiments. For example, whilst the internal heat exchangers of the embodiments have been described such that the high and low pressure fluid flows through the heat opposite directions, or “counter current”, these embodiments could also be implemented using a “same direction” implementation. Furthermore, whilst the refrigerant used in the above described embodiments is R-134a, other refrigerants could equally be used. For example, other low pressure refrigerants or refrigerants that work at high pressures, such as carbon dioxide. Moreover, although the above described embodiments have been described in relation to automotive applications, it will be appreciated that the invention may be applied to a wide range of other applications. These may include for example, busses, lorries, trains and non-mobile applications. Additionally, whilst the above described embodiments have been described as utilizing base tube material that is circular in cross section, other cross sections could be used, such as elliptical cross sections.
The preceding description has been presented only to illustrate and describe exemplary embodiments of the methods and systems of the present invention. It is not intended to be exhaustive or to limit the invention to any precise form disclosed. It will be understood by those skilled in the art that various changes may be made and equivalents may be substituted for elements thereof without departing from the scope of the invention. In addition, many modifications may be made to adapt a particular situation or material to the teachings of the invention without departing from the essential scope. Therefore, it is intended that the invention not be limited to the particular embodiment disclosed as the best mode contemplated for carrying out this invention, but that the invention will include all embodiments falling within the scope of the claims. The invention may be practiced otherwise than is specifically explained and illustrated without departing from its spirit or scope. The scope of the invention is limited solely by the following claims.

Claims

1-33. (canceled)

34. A heat exchanger comprising:

an outer conduit having an inner surface and a longitudinal axis; and

an inner conduit having an outer surface, the inner conduit arranged inside of and along the longitudinal axis of the outer conduit;

wherein the inner conduit and the outer conduit are arranged to form a fluid flow channel between the inner surface of the outer conduit and the outer surface of the inner conduit; and the fluid flow channel has a cross sectional form, in a plane substantially perpendicular to the longitudinal axis of the outer conduit, that is elongate being substantially greater in the circumferential direction of the outer conduit than in the radial direction of the outer conduit.