DE102010008175A1 - Heat transmitter i.e. exhaust gas vaporizer, for use in energy recovery plant of motor car, has pipes combined into tube bundle, where flow pass extends over tube length so that two differently shaped heat transmitter portions are defined - Google Patents

Abstract

The transmitter has a small pipe (10) arranged coaxially in a large pipe (20). A flow pass (3) is formed inside the small pipe and another flow pass (4) is arranged in a space between the small pipe and the large pipe. The small pipe and large pipe are combined into a tube bundle (12). The latter flow pass extends over tube length, so that two differently shaped heat transmitter portions are defined and are connected to each other. The latter flow pass is represented by spiral-like structures in a pipe wall of the large pipe or a pipe wall of the small pipe.

Description

The invention relates to a heat exchanger, comprising a smaller tube, which is arranged in a larger tube and with flow trains for at least two fluid streams, wherein a first flow train is formed in the interior of the smaller tube and a second flow train in an annular space between the smaller and the larger Pipe is arranged, (addition to DE 10 2009 041 406.1 ).

A heat exchanger is out of the DE 199 44 951 A1 known. There is a so-called "internal" heat exchanger in an air conditioner in the high-pressure side and low-pressure side refrigerant are in heat exchange to achieve performance advantages of the air conditioning. The known heat exchanger was designed as a single meandering or spiral wound multi-channel pipe. The second flow trains were represented by webs running along the inner wall of the larger pipe or on the outer wall of the smaller pipe, which in the first case are supported on the inner wall of the outer pipe and abut on the outer wall of the inner pipe in the second case. There are also designs without webs, wherein the inner tube and the outer tube are formed as round tubes and the annular space formed between the two tubes is occupied as a second flow train with a heat exchange rib.

From the DE 196 24 030 A1 already a heat exchanger of the same type is known. This heat exchanger also has a single helical coaxial tube, which is arranged in a container. There special beads are proposed, which result in a compact heat exchanger. Spacers were provided between the inner tube and the outer tube.

The Applicant has already filed a patent application for a heat exchanger containing the file number DE 10 2009 041 406.1 had received. The older application is specifically geared to the use of coaxial flat tubes, which can be produced from sheet metal strips. In the second flow trains are inserted ribs. In the earlier application, in contrast to the publications cited above, the coaxial flat tubes were combined to form a tube stack in order to provide a compact heat exchanger, preferably an evaporator.

An object of the invention is to pick up the initially presented construction principle and further develop it for other heat transfer purposes. Another object is to improve the subject of the earlier application. The solution according to the invention is carried out with a heat exchanger having the features of claim 1. Because the coaxial tubes are combined to form a tube bundle and because the second flow trains of the coaxial tubes are designed differently over the tube length and thus define at least two differently shaped heat exchanger sections, z. As a particularly production-friendly and efficient very advantageous evaporator, which can be used for example in a motor vehicle and, for example, as part of a Rankine cycle to achieve energy recovery z. B. represent from the heat energy of the exhaust gases or the hot charge air.

In the heat exchanger according to the invention, used as an evaporator, z. B. a first fluid, for. As the hot exhaust gas of the engine, heat withdrawn and this is a second fluid, which is the working fluid of a Rankine cycle process, for. As a refrigerant or water, heated, evaporated and superheated. The hot exhaust gas flows through the first flow train inside the smaller tubes, the working fluid flows through the second flow train in an annular space between the smaller and the larger tubes. The inventors have found that it is thermodynamically advantageous for such exhaust gas evaporators if the second flow trains of the first heat exchanger section are structured to be finer than those of the second heat exchanger section. The finely structured heat exchanger section is associated with the incoming fluid to be evaporated. In this flow section, it is advantageous to impart a high flow pressure loss to the still supercooled and therefore fluid fluid. This leads to a more homogeneous distribution of the fluid flow to a plurality of flow trains or tubes connected in parallel in the tube bundle and also to a stabilization of the evaporation process. To generate a desired high pressure loss z. B. the flow-through cross-sectional areas are made small or by special designs of Strömungszuggeometrie flow resistances such. B. deflection measures are introduced. If the fluid reaches the boiling limit, an increase in volume occurs as the proportion of vapor increases. The steam can be in the second heat exchanger section, which is coarser structured, so compared to the first heat exchanger section z. B. has larger spaces, better expand, whereby the heat exchange is improved at a reasonable pressure loss. Of course, more than two differently structured heat exchanger sections can be provided by multiple different configurations of the second flow trains, in which the rooms gradually increase. In particular, there must be no abrupt transition between rather, the sections give a fluid transition, in which the addressed spaces gradually become larger or generally have different geometries. It is understood that all coaxial tubes of the tube bundle are formed identically, which is why the determination of the heat exchanger sections in the heat exchanger is possible.

If the geometry is formed by the inner tubes, for. Example, by rolling, pressing or forming a smooth inner tube as the starting component, or a tube is used, which was subjected to a hydroforming process, the additional advantage of a structured wall surface for the internal flow trains to a generation of turbulence and thus to a Improvement of the heat transfer in the first flow train lead. The performance of the heat exchanger is thereby improved. As explained in more detail below, this design is also particularly advantageous for receiving different thermal elongations and thus to reduce component voltages.

It is advantageous if the fluids in the first and in the second flow train are conducted countercurrently through the heat exchanger. Through the smaller tubes, so by the first flow trains, the exhaust gas or the hot charge air of an internal combustion engine can flow. Through the second flow trains flows to be evaporated liquid. But can also be advantageous power guides in DC or Kreuzgegen- or cross-DC to z. B. to avoid overheating of the fluid to be evaporated beyond its thermal limit temperature and to use the available installation space while maintaining the predetermined pressure loss limits for the fluids in the first and in the second flow train optimal.

To improve the heat transfer and thus to increase the power density of the evaporator, the first flow trains may additionally comprise turbulence inserts.

In addition, it is advantageous if the smaller tubes are not connected metallically with the larger pipes over the pipe length. The smaller pipes are inserted into the larger pipes with a minimal mounting clearance and fixed in position. In operation, the inner tubes generally have higher material temperatures compared to the outer tubes. Due to the greater thermal expansion of the inner tubes in the radial direction of the manufacturing gap is closed and the desired dense current conduction is achieved. On the other hand, but this also makes the production much easier and cheaper to design because z. B. deleted a soldering process. In addition, this structural design is very advantageous for thermal stresses, which are very critical especially in evaporators and often lead to damage. Occurring different thermal expansions in the tube longitudinal direction, z. B. between adjacent tubes with different temperatures can be collected over the entire tube length, whereby the resulting thermal stresses are reduced. By introducing suitable structures in the tube, a flexibility in the tube longitudinal direction can also be realized as they z. B. of bellows knows.

To minimize the gap in operation, the inner tubes z. B. after installation with the help of a mandrel or by internal pressure plastically expanded, or the outer tubes are nestled by a plastic deformation of the inner tubes.

The design of the heat exchanger according to the invention as a tube bundle also offers the advantage of a high space flexibility by only the outer contour of the manifolds must be adapted to the installation space. The outer geometry of the tube bundle is thus adaptable to rugged installation spaces. The length of the entire heat exchanger is arbitrarily adjustable by different tube lengths.

Other features will be apparent from the dependent claims, which are to be considered as individually listed at this point. In addition, these and other features and advantages thereof will become apparent from the following description of embodiments with reference to the accompanying drawings.

But also advantageous is the use of the heat exchanger according to the invention for other cooling or heating tasks. Conceivable z. As the use as intercooler, oil cooler or as an indirect coolant radiator, in which the two flow of coolant at high temperature level (integration of the heat exchanger in a high-temperature coolant circuit) or low temperature level (integration of the heat exchanger in a low-temperature coolant circuit) are flowed through. In an embodiment as intercooler or oil cooler (cooled by air or coolant), in which advantageously the charge air or the oil is passed through the second flow trains, the geometry of these flow trains z. B. to the local density, viscosity, pressure loss and performance requirements. But even a leadership of these media through the inner or through the first flow train may be advantageous.

The 1 shows an overall view of the heat exchanger according to the invention.

The 2 shows a smaller tube.

The 3 shows a smaller tube coaxially arranged in a larger tube.

The 4 shows a similar representation as 1 ,

The 5 shows a representation in the region of the tube ends of the heat exchanger.

The 6 shows one of the smaller tubes on the outside of heat exchanger sections are defined.

The 7 - 10 show tubes with differently designed heat exchanger sections.

The 11a . 11b and 12 show a tube bundle or a heat exchanger with an intermediate collecting space.

The 13 shows several details 12 ,

The 14 shows a smaller tube with a second flow train forming wire winding

The 15 shows a smaller pipe with an insert.

The heat exchanger of the embodiment is an evaporator, which is included in an energy recovery system of a motor vehicle. In the evaporator are the exhaust gases of the motor vehicle engine (empty arrows) and a liquid to be evaporated, such as water (patterned arrows, 1 ) or the like, in the heat exchange. The liquid to be evaporated flows through the in the 2 and 3 visible second flow trains 4 , The first flow trains 3 that are inside the smaller tubes 10 are located, are flowed through by the exhaust gases. Like from the 5 or 12 can be recognized, are - based on their cross section - smaller tubes 10 a little longer than the larger pipes 20 so that they have a projection on both ends 7 ( 12 ). At the ends are collection boxes 30 for the exhaust gas, the tube plates 100 are attached. ( 4 or 5 ) The water can be side of the tubesheets 200 at the entrance 300 Enter, on the flow trains 4 distribute and exit at the opposite end as steam.

The second flow trains 4 are represented by spiral-like structures, for example in the manner of beads in the tube wall of the outer tubes or in the tube wall of the inner tubes, wherein in the one heat exchanger section A, the spiral-like structures are made narrower than in the other heat exchanger section B (FIG. 7 ). The spiral-like structures can be understood as having a variable pitch thread. The tighter structures, or tighter threads, are those that have a smaller pitch. Preferably, the second flow trains designed differently over the pipe length become 4 due to the different design of the tube walls of the smaller and / or larger tubes 10 . 20 shown. However, it is within the scope of this invention that the design of the second flow trains 4 also by means of wound wire windings or similar additional elements 43 around the smaller pipes 10 are arranged around, can be executed, 14 , In this case, the tube walls then remain largely undeformed, so preferably just nested tubes are used.

As a rule, the fluid to be evaporated in a Rankine cycle process is under a high operating pressure. The exhaust gas of the internal combustion engine has an overpressure. Particularly advantageous is therefore the design of the tubes as round tubes, which allows a significant reduction in the wall thickness due to the particularly advantageous pressure-resistant geometry.

6 shows a preferred embodiment in which the second flow trains 4 are represented by spiral-like structures in the tube wall of the inner tubes. These structures are similar to turns, as z. B. from trapezoidal threads are known. In the one heat exchanger section A, structures in the form of single twist windings and in the other heat exchanger section C structures in the form of interrupted single twist windings are provided in the exemplary embodiment. In the heat exchanger section B there is a continuous transition from single twist to interrupted single twist turns.

The structures can be catchy or multi-threaded, as they are z. B. knows of catchy or multi-start spirals or threads, overlays of left and right spirals or threads are conceivable.

To the variety of ways to provide differently designed second flow trains 4 to show, be on the 8th directed. In the 8th an embodiment is shown in which only a portion of the tube wall of the smaller tube 10 equipped with structures with variable pitch. The bigger pipe 20 is also equipped with structures, however go over the entire pipe length. There are sections of constant slope and sections of increasing slope. The structures at the smaller pipe 10 intersect with those at the larger tube 20 and form in this area z. B. a cross-spiral channel guide. In the area where the inner tube 10 has no structures, there is a simple swirl-shaped channel geometry. As this example shows, more than two heat exchanger sections can be seen here.

Like from the 9 and 10 can be seen, the second flow trains 4 alternatively, by spaced annular webs 41 are formed on the inner wall of the outer tubes or on the outer wall of the inner tubes, which are fluid-permeable in the tube longitudinal direction, in which a heat exchanger section B larger distances between the webs - seen in the longitudinal direction of the tubes - are provided as in the other heat exchanger section A. The fluid permeability by means of interruptions 42 of the bridges 41 shown. By shaping the shape and size of the interruptions 42 can the flow passage in the adjacent rooms 50 be made easier or more difficult, ie adapted to the desired turbulence and the pressure loss. 9a shows the design of this geometry through the inner tube. By positioning these breaks 42 a deflection of the fluid flow and thus an extension of the flow paths can be achieved. Also z. B. 2, 3 or more interruptions 42 in a jetty 41 be executed (not shown).

Like in the 10 is shown in a particularly advantageous embodiment, such annular webs with interruptions 42 also in combination with spiral structures for the design of the second flow trains 4 be provided, with the result that a plurality of differently configured heat exchanger sections A, B, C, ... can be achieved, in which the slope of the structures is made variable or remains constant. According to 10 are the annular webs with interruptions 42 in section A. Section B has annular webs with breaks 42 in combination with spiral structures. Advantageously, the formation of the annular webs with interruptions through the inner tube, to achieve additional advantages in terms of turbulence generation and heat transfer in the inner flow train and to allow thermal elongation.

As a result of thermodynamic calculations, it has been found that for the size ratio of the spaces created in the different heat exchanger sections A, B 50 between the spaces at the fluid inlet and the spaces at the fluid outlet, for example, the ratio from 1: 3, for example, 1: 3, 1: 5, 1:10 should be selected to 1:80 or greater, because then in terms of stability and pressure loss particularly favorable heat transfer performance occur 9 . 9a , Advantageously, a continuous or stepwise enlargement of the flow-through cross-sectional area, adapted to the change in the material values or to the heat transfer characteristic variables such. As the density, viscosity, Reynolds or Nusseltzahlen.

In the case of a channel geometry, as known from the prior art and which remains constant over the entire fluid flow length, low flow velocities and thus low Reynolds numbers and thus heat transfers are established at the fluid inlet region. Due to the low flow rates, the pressure loss in this area is very low, which in particular adversely affects the stability of the evaporation. As a result of the strong volume increase of the fluid due to evaporation and overheating, very high flow velocities and pressure losses then result at the outlet region. Since the flow here is highly turbulent, the heat transfer in this area is no longer significantly increased. It is therefore desirable to increase the flow velocity at the fluid inlet to increase the heat transfer, which can be represented with only a slight increase in the pressure loss. On the steam outlet side, however, the flow rate can be significantly reduced, which is hardly disadvantageous in terms of heat transfer performance, but significantly reduces the pressure loss. Overall, the variable heat transfer performance can be maximized by a variable duct geometry without exceeding the maximum pressure drop of the evaporator on the fluid side.

By choosing different local flow cross-sectional areas, but also by other design measures, such. B. locally different surface structures, flow baffles, deflections or Abrißkanten, the local prevailing flow rates and relevant to the heat transfer characteristics such. B. the Reynolds numbers and Nusseltzahlen and the local pressure losses are set to the target sizes.

Thermodynamic simulation calculations have shown that it is advantageous to design a channel geometry which is variable over the run length of the fluid or over the pipe length, at which the local pressure loss (dp / dx, measurable in unit [Pa / m]), ie the pressure loss per flowed through Length for each flow segment, approximately constant. This turns out at the exemplary application, if the area ratio of the traversed trains in the heat exchanger sections at the fluid inlet and outlet something of a ratio of 1: 3, 1: 4, 1: 5 to 1: 8 corresponds. A design of the current-carrying geometry over the pipe length constant pressure losses (pressure loss per unit length) leads to a stable evaporation process. In addition to optimizing the heat transfer performance achievable under the pressure loss limitations, this design therefore additionally leads to an improvement in the stability of the evaporation process. A further change in the area ratio to 1:10, 1:15, 1:20, 1:30, 1:40 and 1:80 can lead to further optimization, in particular if a geometry is shown in which the local Reynolds number remains constant over the flow-through length. This consideration also applies to the above-described design of the fluid guide geometry of spiral-like structures in the single spin or interrupted single twist turns, with one or more generous design, 3 . 6 . 7 . 8th and 9 , as well as their combinations with ring-shaped web structures, 10 ,

It is advantageous that the coaxial tubes 10 and 20 are not connected metallically over the pipe length, so that the smaller pipes 10 due to its greater thermal expansion, caused by the high exhaust gas temperature, to the inner wall of the larger pipes 20 nestle and are not significantly impeded in their elongation.

In the exemplary embodiments shown, it is provided that, as already mentioned, the length of the smaller tubes 10 is greater than the length of the larger tubes 20 . 5 , Furthermore, it is provided that at least one pipe end designed as a perforated plate second floor 200 for the larger pipes 20 and a likewise designed as a perforated plate first floor 100 for the smaller pipes 10 are arranged, which have a distance from each other, which is about half the difference in length of the tubes 10 . 20 corresponds, particularly advantageously the distance is smaller fluid inlet side and the distance steam outlet side made larger. By means of the second floor 200 become the second flow trains 4 of the tube bundle 12 summarized and by means of the first soil 100 the first flow trains 3 , Between the second floor 200 and the first floor 100 is therefore a collection room 40 for the second flow trains 4 running fluid flow, in the case shown for the liquid to be evaporated, is formed. An inlet / outlet 300 for the liquid is at the collection room 40 arranged. At the first floor 100 There are more collection rooms 30 for the exhaust, 1 , which form the guidance of the exhaust gas to the inlet and outlet ports.

It leads to further thermodynamic advantages when an intermediate collection space 400 for the fluid to be evaporated in the tube bundle 12 is arranged, as also from the 1 . 4 . 11a and 12 is apparent. In the exemplary embodiments shown, the intermediate collecting space is located 400 advantageously at the point where the fluid has almost completely evaporated. The intermediate collection chamber is also covered by embedded perforated plates 201 formed, which are covered by a final wall, 13 , The bigger pipes 20 own in the area of the intermediate collection room 400 Pipe wall breakthroughs 401 , In the intermediate collection room 400 can not settle vaporized liquid and z. B. be recycled into the circulation. Also, any unequal distributions of the fluid flow rates of the individual flow trains can compensate each other. In addition, the intermediate collection room fulfills 400 the function of supporting the tubes against each other or to a surrounding housing (not shown). To support and avoid vibration damage additional spacers are distributed over the pipe length 201 arranged in the form of perforated plates whose hole diameter approximately equal to the diameter of the larger tubes (not shown).

For reasons of cost reduction, it may be useful to use a soldering process for the production of the heat exchanger to the fluid-tight connections of the smaller pipes 10 with the first floor 100 and the connection of the larger pipes 20 with the second floor 200 and the fluid-tight seal of the fluid collecting spaces 30 and 40 manufacture. At the same time by a soldering process, the corresponding connections of the pipes 20 to the intermediate collection room 400 and also the attachment of one or more perforated plates 201 be prepared for the support. Either the smaller tubes are used in this soldered version of the heat exchanger 10 not with the larger pipes 20 soldered to get the described thermal change advantages. Alternatively, but here also a solder joint between the smaller pipes 10 with the larger pipes 20 be created if z. B. a vibration problem makes this necessary.

The entire tube bundle is advantageously surrounded on its outer contour by an insulation to reduce heat loss from the outer flow trains to the environment. This insulation can be combined with a housing. Alternatively, for example, by foaming the spaces in the tube bundle, a thermal insulation and at the same time a support of the tubes are realized against each other (not shown).

The 15 shows a smaller tube 10 with an insert 101 , The use 101 is, here only exemplified, for swirling through the first flow trains 3 thought flowing gas. It can be made of a sheet metal and be provided with wing-like exhibited elements that can be geometrically entangled and twisted. Such inserts are provided in particular in those embodiments in which the tube wall of the smaller tubes 10 is not significantly deformed or structured.

LIST OF REFERENCE NUMBERS

3

first flow trains

4

second flow trains

7

Pipe length difference (overhang)

10

smaller pipes

12

tube bundle

20

larger pipes

30

Collecting space (exhaust gas)

40

Collecting space (fluid)

41

Stege

42

interruptions

43

additional elements

50

Rooms at A and B

100

first floor for the smaller pipes 10 (Gas)

101

commitment

200

second floor for the larger pipes 20 (Fluid)

201

Perforated plates for the larger tubes 20 (Fluid)

300

Inlet / outlet

400

Intermediate collection space (fluid)

401

Pipe wall breakthroughs

A

Wärmeübertragerabschnitt

B

other heat exchanger section

C

next heat exchanger section

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 102009041406 [0001, 0004]
• DE 19944951 A1 [0002]
• DE 19624030 A1 [0003]

Claims (18)

Heat exchanger, comprising a smaller tube ( 10 ) which is coaxial in a larger tube ( 20 ) and with flow trains ( 3 . 4 ) for at least two fluid streams, wherein the first flow train ( 3 ) inside the smaller tube ( 10 ) is formed and the second flow train ( 4 ) in a space between the smaller and the larger pipe ( 10 . 20 ), wherein a plurality of coaxial tubes ( 10 . 20 ) to a tube bundle ( 12 ) and wherein the second flow trains ( 4 ) are designed differently over the tube length, so that at least two differently shaped, adjoining heat exchanger sections (A, B) are defined. Heat exchanger according to claim 1, characterized in that the second flow trains ( 4 ) are represented by spiral-like structures in the tube wall of the outer tubes or in the tube wall of the inner tubes, wherein in the one heat exchanger section, the spiral-like structures are made narrower than in the other heat exchanger section. Heat exchanger according to claim 1, characterized in that the second flow trains ( 4 ) are shown by spiral-like structures in the tube wall of the outer tubes or in the tube wall of the inner tubes, wherein in the one heat exchanger section (A) structures in the form of Einfachdrallwindungen and in the other heat exchanger section (B) structures in the form of interrupted or partially interrupted Einfachdrallwindungen are provided. Heat exchanger according to claim 1, characterized in that the second flow trains ( 4 ) are represented by spiral-like turns in the tube wall of the outer tubes or in the tube wall of the inner tubes, wherein the spiral-like structures have an identical pitch over the tube length, in which one heat exchanger section the structures are substantially fluid-impermeable and in the other Heat exchanger section are fluid-permeable. Heat exchanger according to claim 1, characterized in that the second flow trains ( 4 ) are formed by spaced annular webs on the inner wall of the outer tubes or on the outer wall of the inner tubes, which are fluid-permeable in the tube longitudinal direction, in which a heat exchanger section larger distances between the webs - seen in the longitudinal direction of the tubes - are provided as in the other heat exchanger section. Heat exchanger according to claim 1, characterized in that by the first flow trains ( 3 ) a hotter fluid, for example, the exhaust gas or the charge air of an internal combustion engine flows and through the second flow trains ( 4 ) flows a liquid to be evaporated. Heat exchanger according to claims 1-6, characterized in that the coaxial tubes ( 10 . 20 ) are not metallically connected over the tube length, so that the smaller tubes ( 10 ) due to their thermal expansion to the inner wall of the larger tubes ( 20 nestle). Heat exchanger according to one of the preceding claims, characterized in that the length of the smaller tubes ( 10 ) is greater than the length of the larger tubes ( 20 ). Heat exchanger according to claims 1 to 8, characterized in that at least one pipe end designed as a perforated plate first floor ( 100 ) for the smaller tubes ( 10 ) and designed as a perforated plate second floor ( 200 ) for the larger pipes ( 20 ) are provided, which have a distance from each other, which is about half the difference in length of the tubes ( 10 . 20 ), wherein by means of the first bottom, the first flow trains of the tube bundle and by means of the second bottom, the second flow trains of the tube bundle are summarized. Heat exchanger according to one of the preceding claims, characterized in that an inlet / outlet for the one fluid with the first floor ( 100 ) cooperates and another inlet / outlet for the second fluid with the second floor ( 200 ) cooperates. Heat exchanger according to one of the preceding claims, characterized in that between the first floor ( 100 ) and the second floor ( 200 ) a collection room ( 40 ) for the second flow trains ( 4 ) running fluid flow is formed. Heat exchanger according to one of the preceding claims, characterized in that the fluid volume of the collecting space ( 40 ) on the fluid outlet side is greater than the fluid volume of the collecting space ( 40 ) on the fluid inlet side. Heat exchanger according to one of the preceding claims, characterized in that an intermediate collecting space ( 400 ) in the tube bundle ( 12 ) is arranged. Heat exchanger according to one of the preceding claims, characterized in that distributed over the tube length spacers can be arranged in the form of perforated plates whose Hole diameter about the diameter of the larger pipes correspond. Heat exchanger according to one of the preceding claims, characterized in that the heat exchanger can have an insulation, for example by foaming the intermediate spaces in the tube bundle. Heat exchanger according to one of the preceding claims, characterized in that the smaller and the larger tubes ( 10 . 20 ) have an approximately round cross-section. Heat exchanger according to claim 1 and 5, characterized in that for the size ratio of the created in the different heat exchanger sections spaces ( 50 ) is chosen between the spaces at the fluid inlet and the spaces at the fluid outlet about the ratio 1: 3 to 1:80 or greater. Heat exchanger according to claims 1-4, characterized in that the Durchströmungsflächenverhältnis between fluid inlet and fluid outlet in the range of 1: 3 to 1:80 or greater.

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