Heat exchanger and heat pump comprising at least one such heat exchanger

Abstract

The invention relates to a heat exchanger (2, 4), in particular a tube, shell-and-tube, finned-tube and plate heat exchanger, with at least one elongated flow channel (8) through which, during operation, a fluid is guided in a main flow direction (7) corresponding to the longitudinal extent of the flow channel (8), characterized in that the at least one flow channel (8) has internal components and / or design features that impart a swirl to the fluid flowing in the main flow direction (7) in a circumferential direction of the flow channel (8). The invention further relates to a heat pump (1) with at least one such heat exchanger (2, 4).

Description

[0001] The invention relates to a heat exchanger, wherein the heat exchanger has at least one elongated flow channel through which, during operation, a fluid is guided in a main flow direction corresponding to the longitudinal extent of the flow channel. The invention further relates to a heat pump with at least one such heat exchanger.

[0002] Such heat exchangers are used, for example, in heat pump systems and are known in a wide variety of designs in the state of the art. Different heat exchanger types are used, such as tube, shell and tube, finned tube, and plate heat exchangers. One disadvantage of these heat exchangers is that they require a large installation space. Furthermore, they currently achieve only a low COP (Coefficient of Performance). The COP describes the efficiency of the heat pump system. It indicates the ratio of the heat output to the energy input required for its production, which is supplied to the heat pump system in the form of electricity.

[0003] Both WO 20217 / 025151 A1 and US 7 363 769 B2 disclose heat exchangers with a flow channel having internals and / or design features that impart a swirl to the flow.

[0004] Starting from this state of affairs, it is an object of the present invention to create an improved heat exchanger and an improved heat pump of the type mentioned above, which occupy a comparatively small installation space and / or have improved efficiency.

[0005] To solve this problem, the present invention provides a heat exchanger of the type mentioned above, characterized in that the at least one flow channel has internals and / or design features that impart a swirl to the fluid flowing in the main flow direction in a circumferential direction of the flow channel. Investigations have shown that the turbulence of the fluid caused by such a deliberately induced swirl leads to an improvement in the intensity of heat transfer, particularly in the liquid, non-boiling, and gaseous states of the fluid. Furthermore, it has been shown that pressure losses can be minimized thanks to the imposed swirl, especially in the boiling state of the fluid.Accordingly, the efficiency of the heat exchanger according to the invention can be optimized compared to conventional heat exchangers, in which the fluid flows through the flow channel only in the main flow direction, while maintaining the same installation space. Alternatively, the installation space can be reduced while maintaining or improving efficiency.

[0006] According to a first embodiment of the present invention, the at least one flow channel is designed as a pipeline with a particularly circular cross-section, wherein stationary, rigid swirl bodies are inserted into the flow channel as internals. Each of these swirl bodies has a central axis extending in the main flow direction and guide vanes extending radially outwards from this central axis, which impart the desired swirl to the fluid flowing towards the respective swirl body. It has been shown that the intensity of heat transfer in a pipeline with such swirl bodies can be approximately doubled compared to a flow channel without swirl bodies.

[0007] According to a further embodiment of the present invention, the at least one elongated flow channel is designed as a pipe and is divided, at least in sections, into at least two sub-channels extending parallel to each other in the main flow direction, between which a partition wall extends, wherein the first sub-channel is provided downstream and the second sub-channel upstream with a baffle plate extending transversely to the main flow direction, and wherein the partition wall is provided with fluid passage openings through which the fluid introduced into the first sub-channel is directed into the second sub-channel. This forced deflection of the fluid from the first sub-channel into the second sub-channel, caused primarily by the baffle plate of the first sub-channel and the fluid passage openings, imparts a controlled swirl to the fluid in the circumferential direction of the flow channel.With such a design of at least one flow channel, the intensity of the heat transfer of the heat pump system can be increased up to five times compared to conventional heat exchangers, in which the fluid is guided through a simple pipe with a circular cross-section, which in particular enables a significant reduction in installation space.

[0008] Preferably, the at least one elongated flow channel is subdivided, at least partially, into three sub-channels extending parallel to each other in the main flow direction, with a partition extending between each of these sub-channels. The first central sub-channel is provided with a baffle plate extending transversely to the main flow direction downstream, and the second and third sub-channels are each provided with baffles extending upstream. The partitions are provided with fluid passage openings through which the fluid introduced into the first sub-channel is guided into the second and third sub-channels while being subjected to a swirl. With such a design, the greatest increase in heat transfer intensity compared to conventional heat exchangers, in which the fluid is guided through a simple pipe with a circular cross-section, has been achieved.

[0009] Preferably, the first sub-channel has a rectangular or, more preferably, a square cross-section, and the second and third sub-channels each have a semicircular cross-section. This design has proven to be particularly simple, inexpensive, and efficient.

[0010] It is advantageous for the baffle plate of the first sub-channel to be provided with at least one through-hole or, preferably, with at least one through-slot. Thanks to such through-holes and / or through-slots, friction losses can be minimized.

[0011] Preferably, the fluid passage openings are arranged at intervals in the main flow direction, with the distance between adjacent fluid passage openings preferably increasing gradually downstream. This also helps to reduce flow losses.

[0012] According to a further embodiment of the present invention, a plurality of flow channels are provided, wherein each flow channel is formed by a plurality of straight flow channel sections extending in the main flow direction and connected to each other via fluid passage openings, which are arranged overlapping each other in the main flow direction and offset from each other in directions transverse to the main flow direction, wherein each flow channel through which a hot fluid is guided preferably contacts an adjacent flow channel through which a cold fluid is guided over its entire length.This arrangement of the individual flow channel sections, which overlap in the main flow direction and are offset in directions perpendicular to the main flow direction and connected via the fluid passage openings, results in the fluid being subjected to a circumferential swirl as it transitions from one flow channel section to the next. With this design, the best results in terms of increasing heat transfer intensity were achieved in tests, with an improvement of up to seven times compared to conventional heat exchangers where the fluid is guided through a pipe with a circular cross-section.

[0013] Preferably, the flow channel sections are formed by cuboid hollow bars, particularly those with square end faces, each equipped with a fluid passage opening at its free end. This achieves a simple, modular design. The hollow bars can, for example, be bonded together. Alternatively, the hollow bars can be additively manufactured together, so that individual hollow bars exist only virtually and not physically.

[0014] Advantageously, the fluid passage openings are slot-shaped, with the slot width preferably corresponding to 0.1 to 0.3 times the length of an end face of the hollow rod, and in particular 0.25 times. This minimizes friction losses.

[0015] According to a further embodiment of the present invention, the heat exchanger is provided in the form of a finned plate heat exchanger, which has a plurality of flow channels, each of which is bounded by two parallel plates and inclined fins and has a trapezoidal cross-section, wherein at least one end wall of each flow channel is provided with fluid passage openings through which the fluid introduced into a flow channel is directed into an adjacent flow channel while being subjected to a swirl in the circumferential direction of the flow channel.

[0016] Preferably the fluid passage openings (15) are located on the side from which the fluid enters a flow channel. (8)The flow is introduced and each is provided with a cover (27) open on the upstream side. The covers can be produced, for example, by slotting and forming the sheet metal forming the rib, resulting in a very simple design.

[0017] Furthermore, the present invention provides a heat pump with at least one heat exchanger according to the invention.

[0018] Further advantages and features of the present invention will become clear from the following description with reference to the accompanying figures. Therein is Figure 1 a schematic view of a heat pump; Figure 2 a perspective view of a flow channel designed according to a first approach according to the invention, wherein it is a flow channel of a heat exchanger of the in Figure 1 The heat pump shown can be used; Figure 3 an enlarged side view of a Figure 2only schematically represented swirl body; Figure 4 a graph showing the improvement in heat transfer intensity of variants of the in Figure 2 The flow channel shown compares to a reference flow channel as a function of the Reynolds number; Figure 5 a graph showing the increase in hydraulic friction loss of these variants compared to the reference flow channel as a function of the Reynolds number; Figure 6 a diagram showing the improvement in the intensity of heat transfer of these variants compared to the reference flow channel at a Reynolds number of 10,000; Figure 7 a perspective view of a first variant of a flow channel designed according to a second inventive approach, which is a flow channel of a heat exchanger located in Figure 1 The heat pump shown can be used; Figure 8a perspective view of a second variant of a flow channel designed according to a second inventive approach, which is a flow channel of a heat exchanger located in Figure 1 The heat pump shown can be used; Figure 9 a perspective view of a third variant of a flow channel designed according to a second inventive approach, which is a flow channel of a heat exchanger located in Figure 1 The heat pump shown can be used; Figure 10 a perspective view of the in Figure 7 The first variant shown, which exemplifies the deflection of a fluid guided through the flow channel; Figure 11 a graph showing the improvement in the intensity of heat transfer in the Figures 7 to 9 The variants shown are compared to the reference flow channel as a function of the Reynolds number; Figure 12 a graph showing the increase in hydraulic friction loss in the Figures 7 to 9The variants shown are compared to the reference flow channel as a function of the Reynolds number; Figure 13 a diagram showing the improvement in the intensity of heat transfer in the Figures 7 to 9 The variants shown are compared to the reference flow channel at a Reynolds number of 10,000; Figure 14 a perspective view of a heat exchanger, which is one of those in Figure 1 The heat pump shown can be used, in which the flow channels are designed according to a third approach according to the invention; Figure 15 a cross-sectional view along line XV in Figure 14 ; Figure 16 another, partially transparent perspective view of the in Figure 14 depicted heat exchanger; Figure 17 a schematic view of two flow channels of the in Figure 14 depicted heat exchanger; Figure 18 a graph showing the improvement in heat transfer intensity of three variants of the in Figure 17 The flow channels shown are compared to the reference flow channel as a function of the Reynolds number; Figure 19 a graph showing the increase in hydraulic friction loss of three variants of the in Figure 17 The flow channels shown are compared to the reference flow channel as a function of the Reynolds number; Figure 20 a diagram showing the improvement in heat transfer intensity of three variants of the in Figure 17 The flow channels shown compare to the reference flow channel at a Reynolds number of 10,000; Figure 21 a perspective view of a first variant of a flow channel designed according to a fourth inventive approach, which is a flow channel of a heat exchanger located in Figure 1 The heat pump shown can be used; Figure 22 a front view of the in Figure 21depicted flow channel, which shows the deflection of a fluid guided through the flow channel; Figure 23 a perspective view of a second variant of a flow channel designed according to a fourth inventive approach, which is a flow channel of a heat exchanger located in Figure 1 The heat pump shown can be used; Figure 24 a front view of the in Figure 23 depicted flow channel, which shows the deflection of a fluid guided through the flow channel; Figure 25 a perspective view of a third variant of a flow channel designed according to a fourth inventive approach, which is a flow channel of a heat exchanger located in Figure 1 The heat pump shown can be used; Figure 26 a front view of the in Figure 25 depicted flow channel, which shows the deflection of a fluid guided through the flow channel; Figure 27a graph showing the improvement in the intensity of heat transfer in the Figures 21, 23 and 25 The variants shown are compared to the reference flow channel as a function of the Reynolds number; Figure 28 a graph showing the increase in hydraulic friction loss in the Figures 21, 23 and 25 The variants shown are compared to the reference flow channel as a function of the Reynolds number; Figure 29 a diagram showing the improvement in the intensity of heat transfer in the Figures 21, 23 and 25 The variants shown are compared to the reference flow channel at a Reynolds number of 10,000; Figure 30 a graph showing the improvement in the intensity of heat transfer of the flow channels designed according to the four approaches according to the invention compared to the reference flow channel as a function of the Reynolds number; Figure 31a graph showing the increase in hydraulic friction loss of the flow channels designed according to the four approaches according to the invention compared to the reference flow channel as a function of the Reynolds number; Figure 32 a diagram showing the improvement in the intensity of heat transfer of the flow channels designed according to the four approaches of the invention compared to the reference flow channel at a Reynolds number of 10,000; Figure 33 a perspective view showing, by way of example, a conventional heat exchanger and a heat exchanger modified according to the invention.

[0019] The same reference numbers subsequently denote identical or similar components or component areas.

[0020] Figure 1Figure 1 schematically shows a heat pump 1 comprising a first heat exchanger 2, a compressor 3, a second heat exchanger 4, and an expansion valve 5, which are sequentially integrated into a fluid circuit 6 through which a refrigerant fluid is circulated. In the first heat exchanger 2, energy is extracted from the heat source provided by nature (e.g., air, water, or ground) and transferred to the liquid refrigerant, causing it to evaporate. In its vaporous state, the refrigerant is then fed to the compressor 3, which subsequently transfers the compressed refrigerant to the second heat exchanger 4. In the second heat exchanger 4, the refrigerant condenses, and the energy extracted from the refrigerant is transferred to a fluid to be heated, for example, heating water, which is thus heated accordingly.The refrigerant is then fed to the throttle 5 and expanded, after which it is directed back to the first heat exchanger 2.

[0021] In conventional heat exchangers 2, 4, regardless of whether they are designed as tube, shell-and-tube, finned-tube, or plate heat exchangers, the fluids are guided in a straight line through at least one elongated flow channel 8 in a main flow direction 7, the main flow direction 7 corresponding to the longitudinal extent of the flow channel 8. The invention is based on the fundamental idea of ​​providing the at least one flow channel 8 with internal components and / or design features that impart a swirl to the fluid flowing in the main flow direction 7 in a circumferential direction of the flow channel 8. The aim is to thereby increase the intensity of the heat transfer and / or reduce the required installation space of the heat exchanger 2, 4.

[0022] The Figures 2 and 3Figure 1 shows a first approach according to the invention, in which two swirl bodies 9 are inserted as internals in a flow channel 8. The flow channel 8 is a pipe having a smooth inner surface and a circular cross-section with diameter D. The swirl bodies 9, which are fixed and rigidly inserted in the flow channel 8, each comprise a central axis 10 extending in the main flow direction 7 and guide vanes 11 extending radially outwards from this central axis 10, which impart circumferential swirl to the fluid flowing towards the swirl body 9 in the main flow direction 7. For this purpose, the guide vanes 11 are arcuate in this case, with the angle α formed by the main flow direction 7 and a tangent 12 applied to the downstream edge of a guide vane 12 preferably being 60°.This angle α, as well as the number of guide vanes 11, which is eight in this case, can in principle be varied.

[0023] In an initial experiment, a total fluid mass flow rate mt was measured through the following conditions: Figure 2 The flow channel shown is 8: Fluid: Air (ideal gas) Inlet pressure p* in = 1 atm Inlet temperature T* in = 303.15 K Outside temperature T h = 373.15 K Heat transfer coefficient outside HTC h Reynolds number Re = variable Nusselt number Nu based on hydraulic diameter and friction factor f

[0024] In a first series of measurements, the swirl bodies 9 were positioned at a distance L = 6D.

[0025] In a second series of measurements, the swirl bodies 9 were positioned at a distance L = 40D.

[0026] A reference measurement was carried out under analogous conditions using the same flow channel 8 without the swirl bodies positioned therein.9. This reference measurement is marked with the index 0.

[0027] The Figures 4 to 6 show excerpts of the results obtained during the investigations.

[0028] The arrangement of swirl bodies 9 in the flow channel 8 enables, as shown in the Figures 4 and 6 The figures show an increase in the intensity of heat transfer in the flow channel 8 - depending on the Reynolds number - of 1.5 to 4 times, compared to the reference measurement, with a simultaneous increase in hydraulic losses of 2 to 8 times, see Figure 5 . A reduction in the distance L between the two swirling bodies 9 leads to an increase in the intensity of heat transfer as well as an increase in hydraulic losses.

[0029] The Figures 7 to 9Figure 1 shows variants of a flow channel 8 designed according to a second approach of the invention, which is divided, at least in part, into several, in this case three, sub-channels 8a, 8b and 8c extending parallel to each other in the main flow direction 7, between each of which a partition 13 extends. A first central sub-channel 8a of the illustrated flow channel 8 has a square cross-section. The other two sub-channels 8b and 8c flanking the central sub-channel 8a each have a semicircular cross-section.

[0030] The geometries of the in the Figures 7 to 9 The depicted variants of the flow channels 8 are fundamentally identical.

[0031] At the in Figure 7 the first variant of the flow channel 8 shown, which is in the Figures 11 to 13Represented by the numeral "1", the first central sub-channel 8a downstream, and the second sub-channel 8b and the third sub-channel 8c upstream, are each equipped with a completely closed baffle plate 14 extending transversely to the main flow direction 7. Furthermore, the partitions 13 are designed with fluid passage openings 15 through which the fluid introduced into the first sub-channel 8a is guided, imparted with swirl, into the second sub-channel 8b and the third sub-channel 8c. In the illustrated embodiment, the spacing of the individual fluid passage openings 15 in the direction of the main flow direction 7 increases downstream. This geometry is subsequently designated with the index 1.

[0032] The geometry of the in Figure 8 the second variant of the flow channel 8 shown, which is in the Figures 11 to 13 The number "2" represents t, which corresponds in essential parts to that of the one in Figure 7 flow channel shown 8. The baffle plate 14 closing the first middle partial channel 8a is provided here with a central circular through-hole 16.

[0033] The geometry of the in Figure 9 shown third variant of the flow channel 8, which is in the Figures 11 to 13 represented by the number "3", differs in that it is represented by the number in Figure 8 The flow channel 8 shown shows that the baffle plate 14 closing the first middle sub-channel 8a is not provided with a central through-hole 16, but in this case with two horizontal through-slots 17, which are positioned in the upper and lower areas of the baffle plate 14.

[0034] In a second experiment, a total fluid mass flow was measured under the same conditions as in the first experiment through the [unclear text]. Figures 7 to 9 The flow channels shown are 8.

[0035] Figure 10Figure 1 shows how the fluid passing through the fluid passage openings 15 is subjected to a swirl in the circumferential direction of the flow channel 8, as illustrated by the arrows. The fluid velocities are maximal in the arc-shaped outer sections of the second and third sub-channels 8b and 8c.

[0036] The Figures 11 to 13 The following are excerpts of the results obtained during the measurement series, where the simple pipeline served as a reference, index 0.

[0037] The geometries of variants 1 to 3 allow, as described in the Figures 11 and 13 As shown, compared to the reference geometry 0, the intensity of heat transfer in the flow channel 8 increases by a factor of 4 to 10, depending on the Reynolds number, while hydraulic losses increase by a factor of 40 to 100, see Figure 12The provision of a through-hole 16 or through-slots 17 in the baffle plate 14 closing the first partial channel 8a leads to a significant reduction in hydraulic losses with a slight deterioration in heat transfer, see Figures 11 and 12 in summary.

[0038] The Figures 14 to 17 Figure 2 shows a third approach according to the invention for increasing the intensity of heat transfer. The illustrated heat exchanger 2, 4 has a fluid inlet 18 and a fluid outlet 19 for a first fluid, as well as a fluid inlet 20 and a fluid outlet 21 for a second fluid, wherein the fluids are guided through the heat exchanger 2, 4 in counterflow. For this purpose, a plurality of flow channels 8 arranged in a matrix-like cross-section are provided, see Figure 2. Figure 15 , each flow channel having 8, as described in Figure 17As shown schematically, the flow channel 8 is formed by a plurality of straight, flow-direction-extending, and interconnected flow-channel sections via fluid passage openings 22. These sections overlap each other in the main flow direction 7 and are offset from each other in directions transverse to the main flow direction 7. The individual flow-channel sections of a flow channel 8 are positioned such that the flow channel 8 as a whole has a helical shape in the direction of its longitudinal extension. Each flow channel 8, through which a hot fluid is guided, preferably contacts an adjacent flow channel 8, through which a cold fluid is guided, along its entire length, as shown on the left. Figure 17As shown, two flow channels 8 are each helically "twisted". The flow channel sections are each formed by cuboid hollow rods 23 with square end faces, each of which is provided with a fluid passage opening 15 at its free end. The fluid passage openings 15 are each slot-shaped and extend in the main flow direction 7, with the slot width b preferably corresponding to 0.1 to 0.3 times the length d of an end face of a hollow rod 23, in particular 0.25 times. The overlapping arrangement of the individual flow channel sections in the main flow direction 7 and offset in directions transverse to the main flow direction 7 results in the fluid flowing through the flow channel 8 being subjected to a swirl in the circumferential direction of the flow channel 8, as shown in Figure 17as indicated by the dashed lines 24. The flow channels 8 shown can be composed of individual interconnected hollow bars, for example, welded or soldered together. Alternatively, the matrix-like arrangement can also be additively manufactured, so that the hollow bars are only virtual hollow bars.

[0039] In a third experiment, a total fluid mass flow was measured under the same conditions as in the first experiment through the [unclear text]. Figures 14 to 17 The flow channels shown were guided. Here too, the simple pipeline served as a reference, index. 0.

[0040] In a first variant, which is in the Figures 18 to 20In the first variant (represented by the number "1"), the slot width b of the fluid passage openings of the hollow bars was 0.25 times the length d of an end face of the hollow bars, in a second variant (number "2") it was 0.5 times and in a third variant (number "3") it was 1 times.

[0041] The Figures 18 to 20 show excerpts of the results obtained during the investigations.

[0042] The geometries of variants 1 to 3 allow, as described in the Figures 18 and 20 As shown, compared to the reference geometry 0, the intensity of heat transfer in the flow channel 8 increases by a factor of 4 to 7, depending on the Reynolds number, while hydraulic losses increase by a factor of 20 to 45, see Figure 19The geometry with the smallest slot width b exhibits both the greatest increase in intensity and friction. It is assumed that the increased friction, in particular, can be significantly reduced by optimizing the geometry of the flow channels 8.

[0043] The Figures 21 to 26 They show a third approach according to the invention for increasing the intensity of heat transfer in a finned plate heat exchanger. Figures 21, 23 and 25 Figure 1 shows three variants of flow channels with largely identical geometry, bounded by two parallel plates 25 and obliquely angled ribs 26, each with a trapezoidal cross-section. In the case of the Figure 21In the first variant shown, each rib 26 of a flow channel 8 is provided with fluid passage openings 15, through which the fluid introduced into a flow channel 8 is guided into an adjacent flow channel 8 under the influence of a swirl, see Figure 22 The fluid passage openings 15 are each provided on the side from which the fluid is introduced into a flow channel 8 with a cover 27 open on the upstream side, which in this case is produced by slotting and forming the sheet metal forming the rib 26. In the Figure 23 In the second variant shown, both ribs 27 of a flow channel 8 are provided with corresponding fluid passage openings 15, the covers 27 being selected such that the fluid introduced from one flow channel 8 into the adjacent flow channel 8 is then guided again into the next flow channel 8, see Figure 24 . At the in Figure 25 In the third variant shown, fluid passage openings are provided on both ribs 26 of a flow channel 8, wherein the covers 27 are selected such that the fluid from two flow channels 8 is introduced into a third flow channel 8 arranged between them, see Figure 26 .

[0044] In a fourth experiment, a total fluid mass flow was measured under the same conditions as in the first experiment through the [unclear text]. Figures 21, 23 and 25 The flow channels shown were guided. Here too, the simple pipeline served as a reference, index 0.

[0045] The Figures 27 to 29 The following are excerpts of the results obtained during the investigations, where the numbers "1", "2" and "3" again represent the different variants.

[0046] The geometries of variants 1 to 3 allow for an increase in the intensity of heat transfer in the flow channel 8 – depending on the Reynolds number – by a factor of 2.5 to 7, compared to geometry 0 (reference), see the Figures 27 and 29 , with a simultaneous increase in hydraulic losses by a factor of 1.8 to 2.5, see Figure 28 The best results were achieved with the geometry of the second variant.

[0047] The Figures 30 to 32 Approaches 1 to 4, represented by the numbers "1" to "4", are compared. In summary, the second approach is very promising, as it shows the greatest increase in heat transfer intensity. A significant advantage of this second approach, as well as the first and fourth approaches, is that it can be implemented relatively easily in existing heat exchanger designs and, if necessary, even retrofitted. Figure 33The figure on the left shows an example of a conventional heat exchanger 2, 4, whose flow channels 8 are formed by simple smooth pipes with a circular cross-section, and on the right a heat exchanger 2, 4 modified according to the second approach, in which the construction volume is significantly reduced with comparable intensity of heat transfer.

[0048] Figure 32 The intensities of heat transfer of the three approaches are compared at a Reynolds number of 10,000. Although the increase in heat transfer intensity is smaller with the third approach than with the second, the third approach is seen as having the greatest optimization potential. However, the third approach cannot be implemented with existing heat exchangers. Rather, the technical implementation of the third approach requires the construction of a new heat exchanger 2, 4.

[0049] Although the invention has been illustrated and described in detail by the embodiments shown in the figures, the invention is not limited by the disclosed examples and other variations can be derived from them by a person skilled in the art without departing from the scope of protection of the invention as defined by the accompanying claims. It should be noted in particular that the heat exchangers according to the invention can be used advantageously not only in heat pumps but also in other technical fields.

Claims

1. A heat exchanger (2,4), in particular a tube heat exchanger, a shell and tube heat exchanger, a finned tube heat exchanger and a plate heat exchanger, with at least one elongated flow channel (8) through which a fluid is guided through in a main flow direction (7) corresponding to the elongated extension of the main flow channel (8) during operation, characterised in that the at least one flow channel (8) has internals and / or design features which impart a swirl, in a circumferential direction of the flow channel (8), to the fluid flowing in the main flow direction (7).

2. The heat exchanger (2,4) according to claim 1, characterised in that the at least one flow channel (8) is configured as a pipeline, in particular with a circular cross section, and that fixed, rigid swirl bodies (9) are inserted as internals into the flow channel (8), each of which has a central middle axis (10) extending in the main flow direction (7) and guide vanes (11) extending radially outwards starting from said middle axis (10).

3. The heat exchanger (2,4) according to claim 1 or 2, characterised in that the at least one elongated flow channel (8) is configured as a pipeline and subdivided at least in some regions into at least two partial channels (8a,b) extending parallel to one another in the main flow direction (7), between which a partition wall (13) extends, wherein the first partial channel (8a) is provided downstream and the second partial channel (8b) is provided upstream with a baffle plate (14) extending transversely to the main flow direction (7), and wherein the partition wall (13) is provided with fluid passage openings (15) through which the fluid introduced into the first partial channel (8a) is guided into the second partial channel (8b).

4. The heat exchanger (2,4) according to claim 3, characterised in that the at least one elongated flow channel (8) is subdivided at least in some regions into three partial channels (8a,b,c) extending parallel to one another in the main flow direction (7), between each of which a partition wall (13) extends, wherein the first middle partial channel (8a) is provided downstream and the second partial channel (8b) and the third partial channel (8c) are each provided upstream with a baffle plate (14) extending transversely to the main flow direction (7), and wherein the partition walls (13) are provided with fluid passage openings (15) through which the fluid introduced into the first partial channel (8a) is guided, subjected to a swirl, into the second partial channel (8b) and the third partial channel (8c).

5. The heat exchanger (2,4) according to any of claims 4, characterised in that the first partial channel (8a) has a rectangular or preferably square cross-section, and that the second and third partial channels (8b,c) each have a semicircular cross-section.

6. The heat exchanger (2,4) according to any of claims 3 to 5, characterised in that the baffle plate (14) of the first partial channel (a) is provided with at least one through hole (16) or preferably with at least one through slot (17).

7. The heat exchanger (2,4) according to any of claims 3 to 6, characterised in that the fluid passage openings (15) are arranged spaced apart from one another in the main flow direction (7), wherein the distance between adjacent fluid passage openings (15) preferably increases gradually downstream.

8. The heat exchanger (2,4) according to any of the previous claims, characterised in that a plurality of flow channels (8) is provided, wherein each flow channel (8) is formed by a plurality of rectilinear flow channel sections extending in the main flow direction (7), connected to one another via fluid passage openings (22), which are arranged overlapping one another in the main flow direction (7) and offset to one another in directions transverse to the main flow direction (7), wherein each flow channel (8) through which a hot fluid is guided contacts, preferably over its entire length, an adjacent flow channel (8) through which a cold fluid is guided.

9. The heat exchanger (2,4) according to claim 8, characterised in that the flow channel sections are formed by cuboid hollow rods (23) with in particular square end sides, which are each provided with a fluid passage opening (22) at their free ends.

10. The heat exchanger (2,4) according to claim 9, characterised in that the fluid passage openings (22) are formed to be slot-shaped and extend in the main flow direction (7), wherein the slot width (b) preferably corresponds to 0.1 to 0.3 times the length (d) of an end side of the hollow rod (23), in particular 0.25 times.

11. The heat exchanger (2,4) according to claim 1 in the form of a finned plate heat exchanger which has a plurality of flow channels (8) which are each bounded by two parallel plates (25) and obliquely inclined fins (26) and each have a trapezoidal cross-section, wherein at least one end wall of each flow channel (8) is provided with fluid passage openings (15).

12. The heat exchanger (2,4) according to claim 11, characterised in that the fluid passage openings (15), at the side from which the fluid is introduced into a flow channel (8), are each provided with a cover (27) which is configured to be open on the inflow side.

13. A heat pump (1) with at least one heat exchanger (2,4) according to any of the previous claims.

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