DE102010008175B4 - Heat exchanger - Google Patents

Abstract

Heat exchanger, comprising a smaller tube (10) which is arranged coaxially in a larger tube (20) and forms an annular space with it, with flow lines (3, 4) for at least two fluid flows, the first flow line (3) inside the smaller pipe (10) is formed and the second flow passage (4) is arranged in the annular space between the smaller and the larger pipe (10, 20), with several coaxial pipes (10, 20) being combined to form a pipe bundle (12), wherein the second flow lines (4) over the length of the tube have different flow cross-sectional areas, seen in the direction of flow, which are formed by helical beads that extend from the tube wall of the inner tubes (10) to the outer tube (20) and / or from the tube wall of the outer tubes (20) extending towards the inner tube (10), so that at least two differently designed, adjoining heat exchanger sections (A, B) are defined, whereby ei the first heat exchanger section (A) has smaller and the second heat exchanger section (B) has larger throughflow cross-sectional areas.

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 B3 ).

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.

By the DE 43 45 045 A1 a single tube for a heat exchanger with a built-in wire spiral installation element has been known, wherein the wire spiral has a variable pitch and rests against the inner wall of the single tube. The externally applied wire helix imposes a tangentially directed velocity component on the outer flow, ie a swirl flow results, but only in the region close to the wall - the core flow, ie the flow inside the tube is not influenced by the action of the wire helix. There is no coaxial tube with an annulus between an inner and an outer tube.

By the DE 199 55 939 A1 a single tube for a heat exchanger was known, which - to increase the turbulence - directed into the interior of the tube surveys in the form of beads, which are annular or helical. The elevations or beads are embossed from the outside as depressions in the pipe wall, preferably with a roller embossing tool so that they protrude into the interior of the tube and reduce its diameter.

By the DE 94 18 641 U1 a single tube for a heat exchanger with various bead shapes was known. Among other things, the single tube has indentations in the form of simple to fourfold spiral turns and Ausdellungen in the form of spiral beads of the same and opposite slope. The inputs and Ausdellungen are arranged in groups and at constant intervals in the flow direction of the single tube.

By the US 4 171 634 A was according to a first embodiment ( 4 and 5 ) a pipe-in-tube heat exchanger consisting of a coaxial tube, the coaxial tube comprising a smooth tube and a twisted tube having spiral-like turns and variable pitch. The coaxial arrangement of the smooth and the twisted tube spiral-shaped flow channels are formed with increasing flow cross-section, wherein the gap between the flow channels on the outside thereof grows proportionally with the flow cross-section. Another embodiment ( 6 ) shows a housing-and-tube heat exchanger with a bundle of twisted single tubes, ie in this embodiment, no coaxial tubes - as in 4 and 5 shown - used. The production of such a twisted tube by applying a torsional and tensile stress is extremely problematic, especially in terms of dimensional stability and repeatability.

The older application of the Applicant DE 10 2009 041 406 B3 is specially designed for the use of coaxial flat tubes that 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 object is achieved with a heat exchanger according to the independent claims 1, 2 or 7.

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 does not have to be an abrupt transition between the sections, but rather a fluid transition in which the spaces addressed become progressively 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 beneficial For thermal stress, which are very critical especially in evaporators and often cause 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, hereinafter referred to as helical 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, in which a heat exchanger section A the helical structures are made narrower than in the other heat exchanger section B (FIG. 7 ). The helical structures may 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 helical 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-threaded threads, overlays of left and right hand 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, but go over the entire tube 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 , in the following beads 41 are formed in the inner wall of the outer tubes or in 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 beads - seen in the longitudinal direction of the tubes - are provided as in the other heat exchanger section A. Die Fluid permeability is interrupted 42 the beads 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. two, three or more interruptions 42 in a bead 41 be executed (not shown).

Like in the 10 is shown in a particularly advantageous embodiment, such annular beads with interruptions 42 also in combination with helical 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 beads intermittently 42 in section A. Section B has annular beads with breaks 42 in combination with helical structures. Advantageously, the formation of the annular beads 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 about the ratio from 1: 3, for Example with 1: 3, 1: 5, 1:10 until I am to 1:80 or greater should be chosen, because then occur in terms of stability and pressure loss particularly favorable heat transfer performance, 9 . 9a , Advantageously, a continuous or stepwise enlargement of the flow-through cross-sectional area, adapted to the change of 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 demolition edges, the locally 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 occurs in the exemplary application when the area ratio of the traversed trains in the heat exchanger sections at the fluid inlet and outlet corresponds approximately to a ratio of 1: 3, 1: 4, 1: 5 to 1: 8. 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 design of the fluid guidance geometry described above, consisting of helical structures in the single spin or of interrupted single twist windings, with single or multi-slide design, 3 . 6 . 7 . 8th and 9 , as well as their combinations with annular bead 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 boxes 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 tubes 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 collection boxes 30 and collection rooms 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 shown here only by way of example, 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 box (exhaust gas)

40

Collecting space (fluid)

41

Bead (bridge)

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

Spacer for the larger pipes 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

Claims (23)

Heat exchanger, comprising a smaller tube ( 10 ) which is coaxial in a larger tube ( 20 ) is arranged and forms an annular space with this, 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 the annulus between the smaller and the larger tube ( 10 . 20 ), wherein a plurality of coaxial tubes ( 10 . 20 ) to a tube bundle ( 12 ), the second flow trains ( 4 ) have different flow cross-sectional areas, seen in the flow direction, over the tube length, which are formed by helical beads extending from the tube wall of the inner tubes ( 10 ) to the outer tube ( 20 ) and / or extending from the tube wall of the outer tubes ( 20 ) to the inner tube ( 10 ), so that at least two differently shaped, adjoining heat exchanger sections (A, B) are defined, the first heat exchanger section (A) having smaller and the second heat exchanger section (B) larger flow cross-sectional areas. Heat exchanger, comprising a smaller tube ( 10 ) which is coaxial in a larger tube ( 20 ) is arranged and forms an annular space with this, 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 the annulus between the smaller and the larger tube ( 10 . 20 ), wherein a plurality of coaxial tubes ( 10 . 20 ) to a tube bundle ( 12 ), the second flow trains ( 4 ) over the tube length have different flow cross-sectional areas, viewed in the flow direction, which are provided by additional annular elements arranged in the helical elements ( 43 ) are formed, so that at least two differently shaped, adjoining heat exchanger sections (A, B) are defined, wherein the first heat exchanger section (A) has smaller and the second heat exchanger section (B) larger flow cross-sectional areas. Heat exchanger according to claim 1 or 2, characterized in that the helical beads or the additional elements ( 43 ) in the first of the at least two heat exchanger sections (A, B) are designed narrower than in the second heat exchanger section. Heat exchanger according to claim 1, 2 or 3, characterized in that the helical beads or the additional elements ( 43 ) are designed to be catchy or more continuous. Heat exchanger according to claim 1, 2, 3 or 4, characterized in that the helical beads or the additional elements ( 43 ) have a variable slope. Heat exchanger according to one of claims 1 to 5, characterized in that the helical beads in the first of the at least two heat exchanger sections (A, B) are interrupted continuously and in the second heat exchanger section. Heat exchanger, comprising a smaller tube ( 10 ) which is coaxial in a larger tube ( 20 ) is arranged and forms an annular space with this, 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 the annulus between the smaller and the larger tube ( 10 . 20 ), wherein a plurality of coaxial tubes ( 10 . 20 ) to a tube bundle ( 12 ), the second flow trains ( 4 ) over the tube length annular, partially circumferentially formed and in the tube longitudinal direction fluid-permeable beads ( 41 ) extending from the tube wall of the inner tubes ( 10 ) to the outer tube ( 20 ) and / or extending from the tube wall of the outer tubes ( 20 ) to the inner tube ( 10 ), so that at least two differently shaped, adjoining heat exchanger sections (A, B) are defined, wherein the axial spacing of the beads (FIG. 41 ) is smaller in the first heat exchanger section (A) and larger in the second heat exchanger section (B). Heat exchanger according to claim 7, characterized in that the annular beads ( 41 ) Interruptions ( 42 ), which are different in shape and size. Heat exchanger according to one of claims 1 to 8, except claim 2, characterized in that the beads are made by rolling, pressing, forming or hydroforming. Heat exchanger according to one of the preceding claims, 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 one of the preceding claims, characterized in that the coaxial tubes ( 10 . 20 ) are slid over the pipe length with a mounting gap, so that the smaller pipes ( 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 one of the preceding claims, characterized in that on at least one tube end a first plate designed as a perforated plate ( 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 ) corresponds. Heat exchanger according to claim 13, 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 claim 14, characterized in that the first floor ( 100 ) an inlet / outlet for the one fluid and the collecting space ( 40 ) are associated with another inlet / outlet for the second fluid. Heat exchanger according to claim 14 or 15, 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 in the form of perforated plates ( 201 ), whose hole diameters correspond approximately to the diameter of the larger tubes. Heat exchanger according to one of the preceding claims, characterized in that the heat exchanger has 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 one of claims 1 to 6, characterized in that the Durchströmungsflächenverhältnis between fluid inlet and fluid outlet in the range of 1: 3 to 1:80 or greater. Heat exchanger according to claim 7 or 8, characterized in that for the distances of the first annular beads ( 41 ) at the fluid inlet and the last annular beads ( 41 ) is selected at the fluid outlet about the ratio 1: 3 to 1:80 or greater. Heat exchanger according to at least one of the preceding claims, characterized in that the coaxial tubes ( 10 . 20 ) both helical and annular beads ( 41 ) exhibit.

Images