HEAT EXCHANGER

The invention concerns a heat exchanger, in particular, for a motor vehicle, in accordance with the preamble of claim 1.

In the construction of heat exchangers or heat interchangers for motor vehicles, higher requirements are increasingly demanded for exchange performance with a simultaneously restricted installation space. In particular the cooling of combustion gas for the purpose of recirculating it to a combustion engine means that increasingly high heat outputs have to be removed. Also with other heat interchangers, such as oil coolers or charge-air coolers, increasingly higher requirements are being made for transfer or exchange performance. In addition to the high performance, the heat interchangers or heat exchangers must also withstand increasingly higher pressures. In particular with heat exchangers or heat interchangers in which a gaseous fluid, for example, combustion gas or charge air, flows at the site of the heat interchanger to be cooled, the pressures rise steadily due to higher and higher engine loads.

Basically, the performance and the resistance to pressure of a heat interchanger can be increased in that the flow channel cross sections are made smaller. The pressure drop thereby rises considerably both with gas-conducting as well as oil-conducting channels, so that a high output (pump, engine) is required in order to pump the fluids through the coolers.

When using the heat interchanger as a combustion gas cooler on the high pressure side of the engine, there is also the danger that considerable quantities of soot lead to a drop in the heat interchanger performance. Moreover, there is the risk that the cooler will become clogged due to the soot accumulations. The problem of soot accumulation is intensified by a reduction of the channel cross sections.

The goal of the invention is to create a heat exchanger in accordance with the preamble of claim 1, which, with a limited installation space, has a high heat exchanger performance and can be produced at a low cost.

With a heat exchanger, in particular for a motor vehicle, with at least one flow channel with a fluid throughflow, which at least in sections has a curved shape, preferably with several flow channels with a fluid throughflow which at least in sections have a curved shape, the goal is obtained in that the flow channel which at least in sections has a curved shape is provided in an extruded profile. The flow channel has a fluid throughflow for the purpose of a heat exchange. The extruded profile has the advantage that it can be produced at low cost. In investigations carried out within the framework of the invention under consideration, it was determined that the heat exchanger performance of a heat exchanger can be clearly increased by the extruded profiles curved according to the invention, without having to select excessively small flow channel cross sections, which can lead to an enormous rise in pressure drop, or in the case of combustion gas coolers, to clogging due to soot particles.

A preferred embodiment of the heat exchanger is characterized in that several flow channels, which at least in sections have a curved shape, are provided in an extruded profile. At least two flow channels arranged next to one another within an extruded profile are particularly advantageous thereby. A separating interior wall which is integrally connected with the remaining material of the extruded profile is provided between two adjacent flow channels in the extruded profile. In this way, a large contact surface between the fluid and the heat-exchanging material of the extruded profile can be created at low cost and in an operationally reliable matter. In addition, the extruded profile has the advantage that the separation wall or the separation walls, which are also designated as webs, clearly increase the resistance to internal pressure of the extruded profile, so that with such profiles higher pressures can also be employed without damage or deformations appearing in the extruded profile. The interior walls or separation walls or webs in the extruded profile also lead to an increase in surface area, and thus to an increase in the rate of heat release.

It is generally preferred that several extruded profiles be provided to enable an effective heat exchange between the cooling agent and the fluid in the flow channels.

Another preferred embodiment of the heat exchanger is characterized in that the extruded profile has at least one outside wall with a surrounding flow of a medium, in particular, a cooling agent, and at least one inside wall along which flows a fluid, in particular, a combustion gas. At least two flow channels are thereby provided within one extruded profile in a particularly advantageous manner. In this way, a large surface area of contact between the fluid and the heat-exchanging material of the extruded profile can be made available at low cost and in an operationally reliable manner.

Another preferred embodiment of the heat exchanger is characterized in that the outside wall with the surrounding flow of the medium has a cross-sectional shape which is at least partially rounded. In this way, the flexibility of the extruded profile is improved. In accordance with another essential aspect of the invention, an initially linearly extruded profile is provided, in an additional processing step, with the curved shape, in particular, with an undulating profile.

Another preferred embodiment of the heat exchanger is characterized in that several extruded profiles are provided in a particularly integrated manner, which comprises at least one flow channel that has, at least in sections, a curved shape, preferably several flow channels that have, at least in sections, a curved shape. The heat exchanger according to the invention has flow channels which are separate from one another and are preferably shaped in an undulating form. This makes it possible to implement enlarged channel cross sections, and in the case of a combustion gas cooler a clogging problem due to soot does not develop. The high output density of the heat exchanger according to the invention is essentially obtained in that a preferably undulating deflection of a fluid in an undulating extruded profile leads to a vortex formation in the flow. The pressure drop rises only slightly in comparison to a straight extruded profile with the same channel cross-sectional area. The increased heat output can be attributed to a prolongation of the flow path and to turbulence or vortex formation—both on the side of the fluid to be cooled and also on the cooling fluid side.

Another preferred embodiment of the heat exchanger is characterized in that the at least one extruded profile is made of an alloy based on aluminum. Aluminum has quite good corrosion resistance and can be extruded in a low-cost manner in largely arbitrary cross-sectional forms. In particular condensation is formed during the cooling of combustion gas that has a very low pH value and is therefore very corrosive. This is the case with all combustion gas coolers—that is, both with coolers which are used in high-pressure combustion gas recycling and also with coolers which are used in low-pressure combustion gas recycling. With sufficient cooling, aluminum can definitely be used in the construction of combustion gas heat interchangers. It has thereby been shown in particular that extruded profiles without a lengthy heat treatment offer a very good corrosion protection, since the fine grain of the aluminum is not destroyed. This fine grain structure, however, is the prerequisite for corrosion not to produce deep furrows, but rather only a surface material erosion, which in turn guarantees a long service life of the heat exchanger.

Another preferred embodiment of the heat exchanger is characterized in that the at least one flow channel has a corrosion-inhibiting coating. In particular in designing the heat exchanger as a combustion gas heat exchanger, it is possible to prolong the service life of the heat exchanger by means of such coatings.

Another preferred embodiment of the heat exchanger is characterized in that the at least one flow channel has a 180° deflection, in addition to the shape which is curved at least in sections. In this way, a heat exchanger with a U-shaped throughflow is created which is also designated as a U-flow heat exchanger. In a preferred embodiment, two bases are used with a heat exchanger with a U-shaped throughflow. The extruded profiles fit on both sides into these bases. The deflection of the fluid to be cooled preferably takes place in a separate return cap.

Another preferred embodiment of the heat exchanger is characterized in that the curved shape has turbulence-producing bends or undulations. The extruded profile can have different sections with different bends and/or undulations.

Another preferred embodiment of the heat exchanger is characterized in that the curved shape varies transversely to, and/or in the extension direction of the flow channel. The turbulence-producing bends or undulations in the flow channel vary so as to reduced undesired pressure drops.

Another preferred embodiment of the heat exchanger is characterized in that the at least one flow channel has an increasing undulation downstream, in particular, an increasing amplitude and/or a decreasing pitch. For the optimal adaptation of the heat interchanger performance and pressure drop to prespecified requirements, the amplitude and the pitch of the curved shape, in particular, the undulations, are changed in the longitudinal direction of the flow channel. The undulation preferably increases thereby with increasing flow path of the fluid to be cooled, so as to keep as low as possible the pressure drop rise. In this context, increase of the undulation means that either the amplitude increases toward the rear or the division decreases toward the rear. It is also possible to combine the change of the amplitude with the change of pitch.

The variability has advantages with regard to pressure drop, since the preferably hot fluid in the front area of the cooler has the tendency to produce a high pressure drop due to a low fluid density. An additional, artificially produced high turbulence due to a strong undulation in the front area of the flow channel would lead to very high pressure drops there. This strong turbulence is thereby not absolutely necessary in the front cooler area for a high heat output, since the large temperature differential between the fluid to be cooled and the cooling medium is sufficient to confer a high performance even with a low turbulence.

In contrast to the front cooler area a strong turbulence is to be preferred in the rear cooler area since here the increased pressure drop is due less to an increased fluid density, and the temperature differential between the two fluids is too small to transfer the required heat power. Only by means of the high turbulence that is produced by a strong flow channel undulation can the performance be increased sufficiently, even with a low temperature differential between the two fluids. Basically, it is also possible thereby to increase the waviness continuously from the front to the rear.

In addition to the pure sine-like undulation, undulation forms which do not run uniformly are also conceivable. Thus, flow channels which have a sawtooth or a trapezoidal shape, with some straight sections, are conceivable.

Other preferred embodiments of the heat exchanger are characterized in that the ratio between the amplitude and the thickness of the extruded profile is in the range of 0-2, in particular, in the range of 0-0.7, with particular preference, in the range of 0-0.3.

Another preferred embodiment of the heat exchanger is characterized in that the ratio between the pitch and the thickness of the extruded profile is in the range of 3-10.

Another preferred embodiment of the heat exchanger is characterized in that the extruded profile has a thickness in the range of 3-12 mm, preferably in the range of 5-9 mm.

Other preferred embodiments of the heat exchanger are characterized in that the extruded profile has bent or curved areas and/or is reshaped to have undulations. Individual sections of the extruded profile can be reshaped to be undulating. It is, however, also possible for the extruded profile to be reshaped with undulations over its entire length or a great part of its length.

Another preferred embodiment of the heat exchanger is characterized in that the extruded profile is reshaped in a sawtooth or trapezoidal manner. Individual sections of the extruded profile can be reshaped in a sawtooth or trapezoidal manner. However, it is also possible for the extruded profile to be reshaped in a sawtooth or trapezoidal manner over its entire length or a great part of its length. With the trapezoidal undulation, the fluid is initially deflected and vortices are produced. In the following straight stretches, the vortices decay slowly and also increase the heat output in the straight section. Only when the turbulence and the vortices have largely subsided is the turbulence once again stirred up by renewed bends. Basically, moreover, all other conceivable bend forms are also possible. With the sawtooth shape, the rising as well as the descending branch can be steeper than the other branch. The goal of all these variants is to minimize the pressure drop without the heat exchanger performance declining too much.

Also, undulation modifications can be implemented in a heat exchanger with a U-shaped throughflow. Thus, a front flow path before the deflection can have no undulation or a slight one and a rear flow path behind the deflection can have a strong undulation. In this way, a high output with a moderate pressure rise is also obtained for a cooler with a U-shaped throughflow.

Another preferred embodiment of the heat exchanger is characterized in that a bypass flap is upstream or downstream from the heat exchanger. The bypass flap is used to direct the fluid, uncooled, past the area of the cooled flow channels. With a cooler with a straight throughflow, a bypass channel must also be provided, which ideally is thermally insulated via an insulating tube. Preferably, the insulating tube is constructed from stainless steel. With a cooler with a U-shaped throughflow, the bypass function is obtained, for example, in an entry diffuser, in that the bypass flap allows the fluid to flow past the cooler.

Another preferred embodiment of the heat exchanger is characterized in that the fluid is a combustion gas of a combustion engine of a motor vehicle. In particular, the goal of cooling very hot combustion gas can in general be obtained particularly well by a heat exchanger according to the invention, since it has a very high heat exchange performance for a given installation space.

Another preferred embodiment of the heat exchanger is characterized in that the fluid is a charge air of a combustion engine of a motor vehicle. Here also it is possible to obtain clear improvements with the heat exchanger according to the invention.

Another preferred embodiment of the heat exchanger is characterized in that the fluid is a lubricating oil from a lubricating oil circulation of a motor vehicle. Here too it is possible to obtain clear improvements with the heat exchanger according to the invention.

Another preferred embodiment of the heat exchanger is characterized in that the extruded profile is fixed at the end to a base element.

Another preferred embodiment of the heat exchanger is characterized in that both ends of the extruded profile empty into the base element.

Another preferred embodiment of the heat exchanger is characterized in that the extruded profile is fixed at the end to two base elements. The extruded profile extends between the two base elements.

Another preferred embodiment of the heat exchanger is characterized in that the ends of the extruded profile empty into one of the base elements.

With a heat exchanger with a straight throughflow, the flow channels preferably empty, on the entry and exit sides, into a base where they are joined thermally (welded or soldered), joined mechanically (calked or sealed off), or cemented. The bases are connected with a housing of the heat exchanger by welding, soldering, screwing, crimping, or cementing. For a heat exchanger with a straight throughflow, a diffuser is then added to the housing on both sides that is screwed on, welded, soldered, or cemented. For a heat exchanger with a U-shaped throughflow, a diffuser is added only on the entry side, wherein the diffuser contains a separation wall. The entry diffuser or the exit diffuser can each contain a bypass flap so as to direct the fluid, uncooled, past the area of the cooled flow channels.

Another preferred embodiment of the heat exchanger is characterized in that the base element or the base elements are preferably connected with a diffuser in a material-bonding manner. The material-bonding connection can be produced, for example, by welding, soldering, or cementing. The base elements can also be screwed together with the diffuser.

Another preferred embodiment of the heat exchanger is characterized in that the flow channels or extruded profiles are connected, in a material-bonding manner, with the base element or the base elements, for example, by cementing, furnace soldering, flame soldering, induction soldering, or welding.

Another preferred embodiment of the heat exchanger is characterized in that the entire heat exchanger is or will be soldered in a furnace. Basically, in a soldering process the entire cooler, with all the sealing surfaces such as the tube-bottom connection or the bottom-housing connection, are soldered in a soldering furnace (vacuum or Nocolok). In order to maintain the advantages of the fine grain structure for a good corrosion behavior, however, only a local, short-term thermal heating, in particular, in the area of the joint sites, is advantageous. This can be obtained by a local flame soldering, induction soldering, or welding, such as laser welding.

Another preferred embodiment of the heat exchanger is characterized in that the extruded profile is located in a housing with, in particular, a liquid cooling agent throughflow. A particularly effective cooling of the fluid can be obtained in that the flow channels are situated in the housing. However, it is also possible for the housing to be absent and for the fluid to be cooled by means of cooling air.

Another preferred embodiment of the heat exchanger is characterized in that the housing has an inflow and an outflow for the cooling agent.

Another preferred embodiment of the heat exchanger is characterized in that at least one conducting element for guiding the cooling agent is located in the housing. For the further improvement of the flow around the flow channels, baffle plates can be preferably placed on the side with the cooling agent; they can brace the flow channels in vibrations occur and prevent damage to the cooler. Such baffle elements can direct the flow to certain areas and/or produce turbulences in the cooling agent.

Another preferred embodiment of the heat exchanger is characterized in that support agents in the housing are situated to hold the flow channels. The support agents are used to limit the oscillation amplitude of the flow channels and thus to prevent crack formation even with strong vibrations.

Another preferred embodiment of the heat exchanger is characterized in that ribs, baffle plates, or other elements, in particular, support elements, are located between the extruded profiles. The support means can be designed as ribs or turbulence producers and clearly increase the transfer of heat.

Another preferred embodiment of the heat exchanger is characterized in that the elements are soldered in, cemented in, or clamped in between the extruded profiles. To the extent that the support means in a housing are located in the liquid cooling agent, they can also be made of a thermally nondemanding material (such as plastic), so as to lower costs.

Another preferred embodiment of the heat exchanger is characterized in that the housing is essentially made of aluminum. Production costs are reduced in this way.

Another preferred embodiment of the heat exchanger is characterized in that the housing is essentially made of plastic. The production is simplified in this way.

If the cooling medium is not a cooling medium but rather cooling air, then the housing can also be dispensed with. Such a cooler can then be incorporated in the cooling module or another suitable site in the engine compartment, where it is sufficiently supplied with cooling air.

The invention moreover concerns a method for the production of a heat exchanger which will be described first, in which an extruded tube is reshaped in such a way that it has, at least in sections, a curved shape.

Other advantages, features, and details of the invention can be deduced from the following description, in which various embodiments are described in detail with reference to the drawings. The figures show the following:

FIG. 1, a schematic sectional view of a heat exchanger in accordance with a first embodiment, without a bypass channel;

FIG. 2, a schematic sectional view of a heat exchanger similar to that in FIG. 1, with a bypass channel;

FIG. 3A, a schematic sectional view of a heat exchanger in accordance with another embodiment, with a U-shaped throughflow and with a bypass flap;

FIG. 3B, a heat exchanger similar to that in FIG. 3A, without a bypass flap;

FIG. 3C, an extruded profile according to the invention, in cross section;

FIG. 4, a schematic sectional view of a heat exchanger similar to that in FIG. 3B, according to another embodiment;

FIG. 5, a schematic sectional view of a heat exchanger similar to that in FIG. 3B, in accordance with another embodiment;

FIG. 6, a schematic sectional view of a heat exchanger in accordance with another embodiment;

FIGS. 7A and B, two embodiments of undulating flow channels;

FIG. 8, another embodiment of a flow channel with trapezoidal undulations;

FIG. 9, a schematic sectional view of a heat exchanger similar to that in FIG. 1, without a housing;

FIG. 15.2, a representation of the preferred selection of a hydraulic diameter based on measurements and calculations, with a view to an improved heat transfer;

FIG. 17.2, a demonstration of a hydraulic diameter, based on measurements and calculations, in which a stabilization of a pressure drop can be expected at a defined level even with increasing operating time of the flow channel;

FIG. 18.2, a representation of a preferred selection of a hydraulic diameter based on measurements and calculations, with reference to the ratio of the circumference that can be wetted with the first fluid and an outer circumference of the flow channel;

FIG. 19A.2, a modification of a preferred embodiment of a cross section of a flow channel with extruded channel jacket and with the webs extruded with the channel jacket;

FIG. 100A.2, a modification of another embodiment as in FIG. 19A.2, with partial webs;

FIGS. 111A.2 and 111B.2, two modifications of another embodiment as in FIG. 19A.2, with partial webs.

In FIG. 1, a heat exchanger 1 is represented schematically in section. The heat exchanger 1 comprises a housing 2, which emerges from a collecting box 4. The collecting box 4 represents a diffuser and is equipped with an inlet connection 5. Gas is supplied to the collecting box 4 through the inlet connection 5, as is indicated by an arrow 6. A collecting box 8 is located on the opposite side of the housing 2; it also represents a diffuser. The collecting box 8 has a gas outlet connection 9. The exiting gas is indicated by an arrow 10.

Moreover, an inlet connection 14 for a cooling agent is provided on the housing 2. The entering cooling agent is indicated by an arrow 15. Furthermore, the housing 2 is equipped with an outlet connection 16 for the cooling agent. The exiting cooling agent is indicated by an arrow 17. The interfaces between the housing 2 and the collecting box 4, 8 are each defined by a base element 21, 22. Flow channels 24-29 extend between the base elements 21, 22. The flow channels 24-29 are formed in tubes, which are constructed as the extruded profile. In accordance with an essential aspect of the invention, the extruded profiles with the flow channels 24-29 do not have a straight-line, but rather an undulating shape.

FIG. 2 schematically represents in section a heat exchanger 31 similar to the heat exchanger 1 from FIG. 1. To designate the same parts, the same reference symbols are used. In order to avoid repetitions, reference is made to the preceding description of FIG. 1. The differences between the embodiments of FIGS. 1 and 2 are mainly discussed below.

The heat exchanger 31 represented in FIG. 2 comprises a housing 32 with an integrated bypass channel 33. The bypass channel 33 creates a direct connection between the collecting boxes 4 and 8, circumventing the cooled flow channels 24-29. To control the flow, a bypass flap 34 is provided in the collecting box 4. In the position of the bypass flap 34, represented with a solid line, the flow runs through the flow channels 24-29 and not through the bypass channel 33. If the bypass flap 34 is moved into a position 35, indicated with a broken line, then the flow runs only through the bypass channel 33 and not through the flow channels 24-29.

In FIGS. 3A, 3B, and 4, various embodiments of a heat exchanger 41 with a housing 42 are depicted. The housing 42 is equipped with an inlet connection 43 for the cooling agent. The entering cooling agent is indicated by an arrow 43A. The housing 42 is in addition equipped with an outlet connection 44 for the cooling agent. The exiting cooling agent is indicated by an arrow 44A. The housing 42 passes at one side into a collecting box 45 that is equipped with a bypass flap 46 or a mixing valve 46. An arrow 48 indicates that a gas flow is supplied to the collecting box 45. The entering gas flow 48 is cooled by the cooling agent 43A, 44A in the housing 42. The cooled exiting gas flow is indicated by an arrow 49.

A base element 51 is located at the interface between the housing 42 and the collecting box 45. Flow channels 53-55 open into the base element 51; they also proceed from the base element 51. The flow channels 53-55 do not run straight but rather undulate in one section 58 and are deflected by 180° in another section 59.

The mixing valve or bypass flap (46 in FIG. 3A) was omitted in the embodiment depicted in FIG. 3B. Otherwise, the embodiment shown in FIG. 3B is identical with the embodiment depicted in FIG. 3A.

FIG. 3C indicates that the flow channel 53 from FIGS. 3A and 3B is constructed as an extruded profile. The extruded profile 53 comprises four channels 61-64, which are respectively separated from one another by a web 65, 66, 67. The webs 65-67 are also designated as interior walls or separation walls. Gas for the cooling is conducted through the channels 61-64. Channels 61-64 are delimited on the outside by an outer wall 68. The outer wall 68 essentially has a rectangular cross section with rounded-off corners.

The embodiment shown in FIG. 4 indicates that flow channels 71-73 can also be located in the housing 42; they run curved only in one section 75. After a section 76 in which the flow channels 71-73 are deflected by 180°, the flow channels 71-73 run in a straight line in another section 77.

In FIG. 5, a heat exchanger 81 is depicted schematically in section that comprises a housing 82. On one side of the housing 82, a collecting box 84 is provided which is itself subdivided. The entering gas is indicted by an arrow 85. The exiting gas is indicated by an arrow 86. The housing 82 is in addition equipped with an inlet connection 91 for the cooling agent. The entering cooling agent is indicated by an arrow 92. The housing 82 is in addition equipped with an outlet connection 94 for the cooling agent. The exiting cooling agent is indicated by an arrow 95.

A base element 98 is provided at the interface between the collecting box 84 and the housing 82. Another base element 99 is provided on the adjacent end of the housing 82.

At the base element 99, the housing 82 has a deflection section 100. Flow channels 101A, 102A, 103A, 104A, 105A, and 106A extend between the two base elements 98 and 99. The flow channels 101A have an undulating shape. The entering gas 85 from the collecting box 84 arrives at the base element 99 via the flow channels 104A-106A. The gas exiting from the flow channels 104A-106A is deflected jointly in the deflection section 100 of the housing 82 and arrives at the collecting box 84 once again via the flow channels 101A-103A.

FIG. 6 represents in section a heat exchanger 101 that comprises a housing 102. The housing 102 proceeds from a collecting box 104. Gas is supplied to the collecting box 104, as is indicated by an arrow 105. On the opposite side, the housing 102 is delimited by a collecting box 106. The exiting gas is indicated by an arrow 107. The housing 102 is in addition equipped with an inlet connection 109 for the cooling agent. The entering cooling agent is indicated by an arrow 110. The housing 102 is in addition equipped with an outlet connection 111 for the cooling agent. The exiting cooling agent is indicated by an arrow 112.

A base element 114, 115 is in each case provided at the interfaces between the housing 102 and the collecting boxes 104, 106. Flow channels 121 or 126 extend between the two base elements 114 and 115. The flow channels 121-126 are provided in a section 131 with a flatter undulation than in another section 132.

In FIG. 7A, a section of a tube 140 is depicted in a top view. The tube 140 is constructed as an extruded profile and is equipped with a flow channel 141. As is depicted in FIG. 3C, the tube 140 can also be equipped, however, with several flow channels. The tube 140 has an essentially sinusoidal undulating form. The amplitude of the undulating shape is designated by A. The pitch is designated by T. The thickness of tube 140 is designated by d.

FIG. 7B indicates that a tube 145, constructed as an extruded profile with a flow channel 146, can also be undulating in sawtooth form. The amplitude of the sawtooth-like undulation is designated with A. The pitch is designated by T. The thickness of the tube 145 is designated by d.

FIG. 8 depicts a tube 148 with a flow channel 149, which has a trapezoidal undulation. The amplitude of the trapezoidal undulation is designated with A. The pitch of the trapezoid undulation is designated by T. The thickness of the tube 148, constructed as an extruded profile, is designated by d.

FIG. 9 shows a heat exchanger 151 schematically in section. The heat exchanger 151 comprises an inlet element 152 for a fluid. The entering fluid is indicated by an arrow 153. The heat exchanger 151 comprises, moreover, an outlet element 154 for the fluid. The exiting fluid is indicated by an arrow 155. Flow channels 161-166 run between the inlet element 152 and the outlet element 154; they are provided in extruded profiles. Conducting elements 157 for cooling air are situated between the individual extruded profiles. The conducting elements 157 are used simultaneously or alternatively as support elements, and can be cemented, soldered, or wedged with the extruded profiles.

In the embodiment shown in FIG. 9, the extruded profiles with the flow channels 161-166 preferably have a surrounding flow of cooling air. Therefore, in the heat exchanger 151 shown in FIG. 9, it is possible to dispense with a housing. The heat exchanger 151 can be incorporated in a cooling module or on another suitable site in the engine compartment where it is provided with sufficient cooling air.

FIG. 15.2 shows a heat transfer behavior or degree of exchange, and thus the exemplary behavior of a heat transfer performance of a heat interchanger with reference to a calculation based on measurement data, for an example of a heat interchanger designed as a combustion gas cooler. The data are indicated for typical inlet conditions, wherein a combustion gas pressure in the range of 1 bar was selected for simplification. The results, however, are exemplary also for other combustion gas pressures. A curve A shows the behavior of a heat interchanger when not dirtied; a curve B, the behavior of a heat interchanger in the dirtied state. FIG. 15.2 represents the degree of exchange as a function of the hydraulic diameter.

As can be seen with the aid of curve A in FIG. 15.2, the degree of exchange/heat transfer, which is decisive for the heat interchanger performance, increases further with a declining hydraulic diameter for the case that the heat interchanger is not dirtied. The degree of exchange is found in an acceptable range below a hydraulic diameter of 6 mm. As can be seen with the aid of curve B in FIG. 15.2, the degree of exchange declines further below a certain hydraulic diameter in an unacceptable manner, for the case that the heat interchanger is dirtied. Such a lower limit of a hydraulic diameter lies at 1.5 mm. The concept of the invention thus provides for the flow channel to be characterized by a hydraulic diameter which is formed as four times the ratio of the area of the throughflow cross section to a circumference which is wettable by the combustion gas, and which lies in a range between 1.5 mm and 6 mm. Moreover, one can see from the differently shaded areas of FIG. 15.2 that in a preferred manner, the hydraulic diameter should be in a range between 2 mm and 5 mm. As the area of dark shading shows, the comparatively flat upper level of a degree of exchange in a dirtied heat interchanger is in the preferred range of a hydraulic diameter between 2.5 and 3.5 mm or 2.8 mm and 3.8 mm, wherein the latter range is relevant above all for a high-pressure heat interchanger. It has been shown that as a result of an upstream combustion gas purifier before the heat interchanger in the form of the combustion gas cooler, the degree to which a low-temperature heat interchanger is dirtied is less relevant than for a high-pressure heat interchanger in the form of a combustion gas cooler, which is usually exposed to higher particle and fouling loads than a low-temperature heat interchanger. Nevertheless, a pressure drop is relevant for a low-temperature heat interchanger just as it is for a high-pressure temperature heat interchanger.

From the upper curve in FIG. 17.2, one can see that the pressure drop increases further—in this case depicted on the basis of a pressure drop for a flow channel with a limit-value hydraulic diameter of 1.5 mm—with increasing fouling—indicated as operating time in hours. On the other hand, it has been shown that with a selection of a hydraulic diameter of 3.2 mm—also with a selection of a hydraulic diameter in the range between 3.0 mm and 3.4 mm, preferably between 3.1 mm and 3.3 mm—the degree of fouling is obviously stabilized even with increasing operating time, so that the pressure drop is stabilized at an acceptable level.

FIG. 18.2 represents the ratio of the circumference that is wettable by a combustion gas and an outer circumference of the flow channel, as a function of the hydraulic diameter. A preferred ratio is produced from the previously explained, shaded areas of a preferred hydraulic diameter of 2 mm to 5 mm, in particular, 2.8 mm to 3.8 mm. The aforementioned ratio lies in the range between 0.1 and 0.5 in order to obtain improved degrees of exchange and degrees of pressure drop. A comparable tendency can also be determined with the additional constructive designs, described in more detail below, of a cross section in a flow channel with a throughflow. Thus, FIG. 18.2 shows the explained ratio for various web spacings a, (in this case, for two examples a=2 mm and a=5 mm) and for various values of a ratio of a distance between two opposite partial webs to a height of the tube cross section, which, in this case, is designated by k. The ratio k should be in a range below 0.8 mm, preferably in a range between 0.3 mm and 0.7 mm. In this case, the ratio k of a distance e between two opposite partial webs to a height b of the tube cross section increases from 0.25-0.75 in the direction of the arrow. This analysis is valid both for a combustion gas cooler within the framework of a high-pressure design in a combustion gas recycling system, and a combustion gas cooler within the framework of a low-pressure design in a combustion gas recycling system.

Below, FIG. 19A.2 to FIG. 111B.2 describe, by way of example, constructive designs of a cross section of different preferred flow channels. It should be equally clear thereby that modifications of the same and an arbitrary combination of features of the embodiments specifically described in the figures are possible, and a hydraulic diameter in the range between 1.5 mm and 6 mm, preferably between 2 mm and 5 mm, preferably between 2.8 mm and 3.8 mm, can nevertheless be obtained. In particular, with the embodiments shown in the following figures, a modification is shown in which a channel jacket thickness and a web thickness d are the same or similar, and another modification is shown in which a ratio of a web thickness d and a channel jacket thickness s is less than 1.0 mm. Accordingly, it is also possible to vary and adapt the wall thicknesses of partial webs or similar dimensions, depending on the objective to be attained.

In particular, the following true-to-scale figures show embodiments of flow channels for a combustion gas recycling system or a heat interchanger, for example, instead of the flow channels in the combustion gas heat interchanger. In particular, the flow channels explained below all fulfill the prerequisites of a hydraulic diameter in accordance with the concept of the invention.

FIG. 19A.2 shows two modifications of a flow channel 1061, wherein the jacket thickness s and the web thickness d are essentially the same. Moreover, the same reference symbols are used for the same features.

The flow channel 1061 is formed as a profile which is, as a whole, extruded—that is, as an extruded channel jacket together with the extruded webs. Accordingly, the flow channel 1061 has a channel jacket 1063 with an interior space 1067 that is surrounded by a channel jacket inner side 1065, which in this case is designed for the heat-exchanging conduction of the first fluid in the form of a combustion gas. Furthermore, the flow channel 1061, in this case, has five webs 1069, situated in the interior space 1067 on the channel jacket inner side 1065, which are formed together with the channel jacket 1063 as an integral extruded profile. A web 1069 runs entirely parallel to a flow channel axis, standing perpendicular to the drawing plane, uninterrupted along the flow path formed in the housing of a heat interchanger. The throughflow cross section shown, transverse to the flow channel axis, is designed to conduct the combustion gas in the interior space 1067. Dimensioning is effected based on the hydraulic diameter d_hfor the flow channel profile 1061 under consideration, with reference to the distances a, b. The hydraulic diameter turns out to be four times the ratio of the of the throughflow cross-sectional area to a circumference which can be wetted by the combustion gas. The area of the throughflow cross section is in this case a multiple of the product of a and b. The wettable circumference is in this case also the multiple of double the sum of a and b. In this case a gives the width of the free cross section of a flow path 1074 that is subdivided by the webs 1069 in the flow channel, and b in this case gives the free height of the flow path 1074.

Explained in more detail, in this flow channel 1063 and in the following flow channels, a wall thickness s is in the range between 0.2 mm and 2 mm, preferably in the range between 0.8 mm and 1.4 mm. A height b of a line of flow 1074 or a height of the interior space 1067 is, in this case, in the range between 2.5 mm and 10 mm, preferably in the range between 4.5 mm and 7.5 mm. A width a of a line of flow 1074 is in the range between 3 mm and 10 mm, preferably in the range between 4 mm and 6 mm.

FIG. 100A.2 shows a modification of a particularly preferred embodiment of a flow channel 1071, which—as explained previously—differs merely in the wall thickness of the channel jacket 1073 relative to the wall thickness of a web 1079. The flow channel 1071 also has the webs 1079 in the form of whole webs and in addition, partial webs 1079′, situated alternately relative to the whole webs 1079. The flow channel 1071 is in turn formed entirely as an extruded profile, wherein a line of flow 1074 is formed in turn by the distance between two whole webs 1079. In this case, two partial webs 1079′ are situated with front ends 1076 opposite one another.

In FIG. 111A.2 and FIG. 111B.2, two other modifications 1081, 1081′ of a particularly preferred embodiment of a flow channel 1081, 1081′ are shown in which two partial webs 1089′ are located with front ends 1086 that are staggered laterally with respect to one another.

A ratio of a distance a₃, from a first partial web 1089′ to a whole web 1089, to a distance a₄, from a second partial web 1089′ to the whole web 1089, lies in a range between 0.5 mm and 0.9 mm, preferably in a range between 0.6 mm and 0.8 mm. Basically, the distance e between two opposite partial webs 1079′ and/or between two partial webs 1089′, staggered with respect to one another, to a height b of the tube cross section is in a range below 0.8 mm, in particular, in a range between 0.3 mm and 0.7 mm.

The extruded parts described in FIGS. 1, 2, 3A, 3B, 3C, 4, 5, 6, 7A, 7B, 8, 9, 15.2, 17.2, 18.2, 19A.2, 100A.2, 111A.2, and 111B.2 are in particular made from aluminum. In particular, the extrusion materials have the following percentages by mass, especially for corrosion protection.

Silicon: Si<1%, in particular Si<0.6%, in particular Si<0.15%

Iron: Fe<1.2%, in particular Fe<0.7%, in particular Fe<0.35%

Copper: Cu<0.5%, in particular Fe<0.2%, in particular Cu<0.1%

Chromium: Cr<0.5%, in particular 0.05%<Cr<0.25%, in particular 0.1%<Cr<0.25%

Magnesium: 0.02%<Mg<0.5%, in particular 0.05%<Mg<0.3%

Zinc: Zn<0.5%, in particular 0.05%<Zn<0.3%

Titanium: Ti<0.5%, in particular 0.05%<Ti<0.25%

To obtain a high corrosion resistance of the aluminum alloys, the grain sizes, measured in the section of the component in the extrusion direction, <250 micrometers, in particular, <100 micrometers, in particular, <50 micrometers.

HEAT EXCHANGER

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