EP1492990A2

EP1492990A2 - Heat exchanger and cooling system

Info

Publication number: EP1492990A2
Application number: EP03732268A
Authority: EP
Inventors: Peter Ambros; Wolfgang Dreher; Bruno Knauf
Original assignee: Behr GmbH and Co KG
Current assignee: Mahle Behr GmbH and Co KG
Priority date: 2002-03-20
Filing date: 2003-03-17
Publication date: 2005-01-05
Also published as: JP2005527764A; WO2003078911A3; CN1643326A; US20050092475A1; AU2003239792A1; WO2003078911A2; CN100573016C; DE10212249A1; BR0303650A

Abstract

The invention relates to a heat exchanger (2), especially for use in a cooling circuit (1) of a vehicle, comprising at least one heat transfer network (3), whereby the heat transfer network (3) comprises at least one tubular device, and a cooling circuit (1) associated therewith. According to the invention, the tubular device has a characteristic hydraulic diameter which is smaller than or equal to 2.0 mm.

Description

Heat exchanger and cooling system

The invention relates to a heat exchanger and a cooling system, in particular for use in a motor vehicle. Although the invention is described below with regard to use on a motor vehicle, it should be pointed out that the heat exchanger and the cooling system according to the invention can also be used for other cooling processes.

In the prior art, cooling systems are used in motor vehicles, e.g. to give off the waste heat of the internal combustion engine to the environment. Water is generally used as the coolant, which contains additives such as antifreeze. In the coolant cooler, the heat given off by the engine in the cooling circuit is dissipated to the environment by passing an air stream past the surfaces of the coolant cooler.

For the purposes of this application, the medium in the interior of the cooling circuit is always referred to as coolant. By contrast, cooling medium is the (external) medium to which the cooling capacity of the cooling circuit is given. For example, in a cooling circuit of an internal combustion engine, the waste heat of the engine is absorbed by the coolant contained in the cooling circuit. In the coolant cooler, the waste heat from the cooling then removed to the coolant flowing through the coolant cooler. In a conventional cooling circuit, this is the cooling air.

To increase the heat flow, the heat-emitting surface of the coolant cooler is increased by dividing the coolant flow into a number of parallel coolant pipes, on which cooling fins are arranged, in order to effectively dissipate the heat to the environment.

In practice, a motor vehicle manufacturer or the manufacturer of the cooling system specifies technical conditions that the heat exchanger of the cooling system must comply with. The thermodynamic data such as thermal output at given operating temperatures of the coolant and the environment and maximum pressure loss at a given mass flow of the coolant in the operating state are specified.

The pressure loss is limited with regard to the performance and dimensioning of the coolant pump.

Typically, the individual components of the cooling circuit are designed for so-called critical operating states of the vehicle, in which, for example, when driving uphill under certain load conditions and prevailing outside temperatures, a predetermined amount of heat must be able to be released to the environment without exceeding the permissible limit temperatures.

In order to achieve the aforementioned criteria, heat exchangers for coolant circuits in motor vehicles are designed in such a way that the flow resistance is low at maximum heat transfer performance. Maximum flow losses on the coolant side in the heat exchanger or heat transfer network are specified, which must not be exceeded,

Although the maximum pressure loss in the coolant cooler is low, the pressure loss in the entire cooling circuit is high overall. It is the object of the invention to provide a heat exchanger which, with the same heat transfer properties, allows the power requirement for the pump of the cooling circuit to be reduced. It is furthermore preferably an aspect of the object of the invention to provide a cooling system in which overall a smaller pump output is possible to overcome the flow resistances in the coolant circuit.

The heat exchanger according to the invention is the subject of claim 1. The cooling system according to the invention is the subject of claim 9. Preferred developments are the subject of the dependent claims.

The heat exchanger according to the invention is particularly, but not only, intended for use in a cooling circuit of a motor vehicle. At least one heat transfer network is provided between the coolant inlet and the coolant outlet of the heat exchanger. At least one heat transfer network comprises at least one pipe device, wherein the pipe device or the pipe devices are provided to transport coolant through the heat exchanger from the coolant inlet to the coolant outlet while the waste heat is dissipated. A characteristic hydraulic diameter of a pipe device is less than or equal to 2.0 mm in a heat exchanger according to the invention.

For the purposes of this application, the hydraulic diameter d _hy d _{r is} defined as four times the cross-sectional area divided by the inner circumferential area, as is also used, for example, in the VDI Heat Atlas when designating multi-chamber pipes. In the VDI Heat Atlas, the hydraulic diameter d _{hy r} denotes the hydraulic diameter of all parallel flowed through chambers. In the following, the cross-sectional area denotes the inner cross-sectional area that is available for the flow of the coolant. The inner circumference is the circumference around the flow channel device inside the pipe device.

This means that with a circular and ideally designed pipe device, the hydraulic diameter d _{hydr is} equal to the diameter of the Pipe device d is. In the case of a pipe device with a square inner surface, the hydraulic diameter d _{hydr corresponds to} an inner side length of the pipe device.

The heat exchanger according to the invention has many advantages.

Heat exchangers for coolant circuits of motor vehicles that have become known in the prior art have a hydraulic diameter d _hy dr, which is, for example, 2.8 mm and larger. Such hydraulic diameters are chosen in order to minimize flow losses in the heat exchanger.

On the other hand, smaller hydraulic diameters, as in accordance with this invention, have greater pressure losses at a constant flow rate of the cooling medium, so that heat exchangers according to the invention lead to higher pressure losses with a constant mass flow. Therefore, experts did not consider researching heat exchangers for the coolant circuit of a motor vehicle, in which there are smaller hydraulic diameters, since these may be outside the standards specified by the motor vehicle manufacturers.

Surprisingly, however, it has been shown that the use of smaller hydraulic diameters not only leads to higher flow resistances and thus pressure losses, but at the same time improves the heat transfer on the inside of the cooling medium in such a way that a smaller coolant mass flow is required overall in order to obtain an equal amount of heat to the surroundings evacuate.

It is therefore possible with a heat exchanger according to the invention to reduce the coolant mass flow in the cooling circuit of a motor vehicle. In the case of incompressible cooling media, such as cooling water, a reduction in the coolant mass flow leads directly to a proportional reduction in the flow rate of the coolant in the cooling circuit. Since the flow loss in the cooling circuit is proportional to the square of the speed of the coolant, a reduction in the Coolant mass flow to about 70%, a halving of the flow losses in the overall circuit and a reduction in the hydraulic delivery capacity DP times the volume flow to approx. 35%.

When using a heat exchanger according to the invention, the flow resistance in the heat exchanger is increased by the small hydraulic diameter of the pipe devices in the heat exchanger. The flow rate of the coolant can be reduced while the heat output remains the same outside the cooler, so that the flow rate of the coolant in the periphery is significantly lower than with a conventional heat exchanger.

Peripherals are understood here to mean all components and component areas in the cooling circuit through which coolant flows, with the exception of the pipe devices.

Although the pump output required to overcome the flow loss in the heat exchanger can be considerably higher (e.g. factor 2 or 4) than with a conventional heat exchanger, the higher heat transfer capacity of the heat exchanger can reduce the amount of coolant circulating required for cooling. By reducing the pressure losses in the periphery, the required pump output of the pump device is reduced, so that the present invention allows overall energy to be used to operate the pump (e.g. factor 1, 5 or 2).

If the invention is used in a circuit with an electrical pump, the saving effect on primary energy is greater, since the conversion losses from mechanical energy into electrical energy are also lower.

Another advantage of the heat exchanger according to the invention is that when an electrical pump device is used, an electrical pump with a significantly lower electrical output can be used, so that costs for the pump, battery, alternator, etc. can be saved. Gas and particularly preferably air are used as coolants to absorb the amount of heat removed.

It should be pointed out that the invention or an advantageous development of the invention can also be used in a heating circuit or in any cooling circuit. The invention can also be used in parallel circuits or in multi-circuit systems.

In a preferred development of the invention, at least a multiplicity of essentially identical pipe devices are provided. It is also possible for a first type of tube device to be provided with a first plurality and for a second or even more type of tube devices with a second (third etc.) number in each case.

In a further development of the invention, at least the cross section of at least one type of pipe device is taken from a group of cross-sectional shapes, which comprises round and circular, elliptical, oval, angular, rectangular, triangular, square and rounded modifications of the cross-sectional shapes mentioned.

The cross section of at least one pipe device is preferably substantially constant over at least one length section along the pipe device. It is preferred to use flat pipe devices in which the flow channel of the cooling medium has a relatively small width and a relatively high depth (in each case transverse to the flow direction of the cooling medium).

To improve the pressure resistance, webs can be provided in the flow channel, which, for example, divide a flat flow channel into rectangular or square or round or circular segments. Then a pipe device contains, for example, pipe segments. In the following, reference is always made to the dimensions of the pipe device, even if the pipe device should contain segments. In a preferred development, the pipe devices can also be provided with turbulence insert devices or fin devices in the pipe devices in order to increase the turbulence and the heat transfer. The characteristic hydraulic diameter is not changed by these fin devices and turbulence insert devices.

In a preferred development of the invention, essentially all the pipe devices are arranged essentially parallel to one another, the cooling medium passing transversely through the pipe devices arranged essentially in parallel. Rib devices are preferably provided on the pipe devices, which can have gill devices in order to increase the heat transfer on the outside of the pipe devices.

In a preferred development of the invention, a characteristic cross-sectional ratio of the depth of a pipe device in the flow direction of the cooling medium to the height of a pipe device is between 1 and 100 and preferably between 7 and 50, particularly preferably between 15 and 50, particularly preferably between 20 and 30.

This means that the pipe devices have a substantially greater extent in the flow direction of the cooling medium than in a direction perpendicular thereto and to the flow direction of the coolant. The numerical values mentioned can relate to the external or internal dimensions of the pipe devices.

In a preferred development of one or more of the developments described above, the coolant contains water as an essential constituent, wherein the coolant can also have additives such as antifreeze and other additives. It is equally possible that the coolant contains waterless or only a small amount of water. The invention can also be used in radiators. It is also possible for the invention to be used for cooling or heating engine oil, gear oil or fuel, for example of a motor vehicle. is set. Depending on the application, the coolant may contain oil or other coolants known in the prior art.

Gas and particularly preferably air are preferably used as the cooling medium on the outside of the pipe devices.

The cooling system according to the invention has at least one pump device, at least one heat source device (such as a motor device) and at least one heat exchanger device, the heat exchanger device comprising at least one heat transfer network device. The pump device, the heat exchanger device and the heat source device are connected to form an essentially closed cooling circuit and are flowed through by at least one coolant.

The pressure loss of the heat transfer network device of the heat exchanger device in relation to the pressure loss in the entire coolant circuit, assessed before and after the pump device, is at least 12%, preferably more than 15%.

The cooling system according to the invention has many advantages.

The pressure loss of the heat transfer network device is preferably in the range between 15 and 90% of the pressure loss in the entire coolant circuit in the operating state and particularly preferably in the range between 20% and 70%. It is preferably at least 30%.

In a preferred development of the cooling system according to the invention, the heat transfer network device comprises pipe devices, at least one pipe device being taken from a group of pipe devices, the pipe devices with hydraulic diameters <2 mm and in particular in the hydraulic diameter range between 1 and 1.8 mm, and dimple pipes , Pipe devices with turbulence insert devices and the like includes more. Turbulence deposits can For example, be (metal) spirals or foils or threads that are introduced into the pipe devices.

In a preferred development of the invention, the coolant flow in the heat exchanger device is deflected at least once. The coolant flow in the heat exchanger device can also be deflected 2, 3, 4, 5, 6 or more times.

A heat exchanger device particularly preferably has a fin density in the range between 50 and 120 per decimeter length of the pipe device, the thickness of the individual fins being between 0.01 and 0.5 mm, preferably between 0.05 and 0.2 mm. The greater the fin density, the more heat can be transferred in principle, however, a large fin density, in particular in connection with a large thickness of the fins, reduces the cross-sectional area available for the cooling medium, such as the cooling air flow. The optimum results from the number, thickness and length of the cooling fins for a given material for the cooling fins.

The total flow resistance in the cooling circuit is usually essentially determined by the flow resistances in the connecting hoses, the water boxes, the heat transfer network of the heat exchanger, the thermostats connected in series and the engine block.

It should be pointed out that in the sense of this application, the pressure loss portion of the heat transfer network in the entire cooling circuit is determined by this definition. If other components are added or missing in another system, the numerical values mentioned here may need to be converted or adjusted accordingly.

Further advantages and possible uses of the present invention will now be illustrated on the basis of exemplary embodiments with reference to the drawings. In it show: 1 shows a schematically illustrated cooling circuit according to the invention;

2 shows a first heat transfer network for the heat exchanger according to the invention;

3 shows a heat transfer network for a second heat exchanger according to the invention;

4 shows a heat transfer network for a third heat exchanger according to the invention;

5 shows a diagram for determining an optimal hydraulic diameter for a first tube wall thickness;

6 shows a further diagram for determining an optimal hydraulic diameter with a second tube wall thickness;

7 shows a diagram for determining an optimal fin density for a given tube wall thickness;

8 shows the cooling air throughput over the fin density for different pipe divisions; and

9 shows the proportion of the pressure loss in the cooler to the total pressure loss in the cooling circuit above the hydraulic diameter.

1 shows an exemplary embodiment of a cooling system 1 according to the invention. The coolant system 1 is intended for use in a motor vehicle and is used to cool the engine 5. The heated coolant emerging from the engine is passed through the thermostat 7 and enters a water tank 4 of the heat exchanger 2.

If, for example, the operating temperature of the engine is not reached shortly after the vehicle has started, the coolant can also be discharged via the bypass 8 are guided past the heat exchanger 2 and are passed back into the motor 5 via the pump 6.

The heat exchanger 2 has a heat transfer network 3. Such a heat transfer network 3 according to the invention is shown in different embodiments in FIGS. 2, 3 and 4, in which the arrangement and the dimensions of the individual components differ.

The heat transfer network shown in FIG. 2 has tubes 11 in the form of flat tubes, which have a depth 12 in the direction of flow of the cooling medium, which is 32 mm in the selected exemplary embodiment. Depending on the required dimension of the heat transfer network, the depth of the flat tubes 11 can also be 10, 12, 16, 20, 24, 32 or even 40 or 48 mm or values in between. However, other values are also possible if the requirements for the heat exchanger so require.

The flat tubes 11 according to FIG. 2 have a width 13 of 1.3 mm with a wall thickness 17 which is essentially constant over the circumference of the flat tube and is only 0.26 mm. This corresponds to clear internal dimensions of 31, 48 mm in depth and 0.78 mm in width.

With the free cross-sectional area 23 and the inner circumference 24, a hydraulic diameter D _h y _d is obtained (assuming a rectangular inner cross section) with the above-mentioned dimensions _. = 4 x inner surface / inner panel = 1.52 mm. When using a tube with a depth of 16 mm instead of a depth of 32 mm, the hydraulic diameter of D _hy dr = _1.48 mm results under the conditions mentioned.

If, on the other hand, the depth is doubled to 64 mm, the hydraulic diameter is Dh _y r = 1.54. This means that the hydraulic diameter D _hy dr is essentially influenced by the inner width of the individual flat tubes, while a greater or lesser extent in depth affects the value of the hydraulic diameter only over a wide range to a small extent. By increasing the expansion of the flat tubes in depth, the larger heat transfer area on the one hand increases the transferred heat output, while on the other hand the flow rate of the coolant decreases because the cross-sectional area increases. If the flow rate of the coolant remained constant, a larger mass flow of the coolant would be transported.

In order to increase the cooling capacity while the coolant throughput remains the same, it is better if the hydraulic diameter is reduced by reducing the clear flow width in the tubes 11. This leads to a higher heat transfer on the inside of the tubes, which under certain circumstances increases the overall heat transfer performance considerably.

The coolant mass flow can then be reduced with the same heat output of the cooler. If the flow cross-section remains the same, this leads to a reduction in the flow rate of the coolant and thus to lower flow losses.

The distance 14 of the individual tubes 11 is 9.3 mm in the exemplary embodiment according to FIG. 2. The rib height of the ribs 15 is 8 mm. To increase the heat transfer capacity, the fins are provided with gills 16, so that the boundary layers are always newly formed.

2, the wall thickness 17 of the tubes 11 is 0.26 mm. Smaller or larger wall thicknesses are also possible, such as. B. 0.35 mm. The tendency is to reduce the wall thickness in order to save weight and material and to improve the heat conductor resistance. However, the minimum wall thickness also depends on the pressure within the system.

In the exemplary embodiment according to FIG. 2, the ribs 15 are on the tubes. 11 soldered, while in the exemplary embodiment according to FIG. 3 they are mechanically fastened or clamped. Rib elements 15 are plugged onto the circular tubes 21 in the exemplary embodiment according to FIG. 3. Anschlies- Send the tubes 21 are stretched so that there is a larger outer diameter. The fin elements 15 are held firmly on the tubes 21.

Heat transfer is usually better for soldered connections. Therefore, in the exemplary embodiment according to FIG. 3, when using circular tubes in the heat exchanger, a soldered connection between the fin and the tube can be provided.

In the exemplary embodiment according to FIG. 3, the inside diameter 18 of the pipes 21 is equal to the hydraulic diameter Dhydr. The pipe devices 23 have an inside diameter 18. The wall thickness 25 is shown in FIG. 4.

So-called radius pipes 21 are used in the heat transmission network shown in FIG. 4. The tubes 21 have a depth 12 (in the flow direction of the cooling air) and a maximum width 13. The hydraulic diameter is again calculated from the inner flow surface 23 and the inner circumference 24 from D _hy dr = four x flow cross section 23 / inner circumferential surface 24. The inner circumferential surface 24 and the flow cross section 23 can be determined with knowledge of the depth 12 and the maximum width 13 as well as the tube wall thickness 25 and the geometric contour. ,

The individual tubes 21 are arranged at a lateral distance 14. As in the exemplary embodiment according to FIG. 3, two rows of heat transfer tubes 21 are provided which have a tube row spacing 19.

There may also be fewer, that is to say one row of pipes or even more, for example 3, 4 or 5 rows of pipes. The individual rows of pipes can be aligned in the air flow direction or offset.

5 shows the hydraulic power of the cooler circuit over the pressure loss of the heat transfer block in a diagram. The Heat transfer block consists of the heat transfer network and the floors. The heat transfer network consists of the coolant tubes with the cooling fins.

The diagram was made with reference to a medium-sized car with e.g. B. created a ^" 1.7 liter diesel engine.

The measurement point 30 shown marks a current production model in which, in a certain operating state, a hydraulic delivery capacity of approximately 270 W is required in order to provide the necessary coolant mass throughput for the engine 5 and cooler 7.

All measuring points shown have been determined for a constant cooling capacity. The wall thickness of the tubes of the heat transfer network was 0.35 mm for FIG. 5 and 0.26 mm for FIG. 6. The measuring points 33, 34 and 35 were derived for a cooling system in which the heat exchanger was provided with different heat transfer networks.

While measuring point 30 indicates the current state of the art, in which the hydraulic diameter of the tubes used is large and here is approximately 2.5 mm, the hydraulic diameter for measuring point 33 has been reduced to 1.94 mm. The hydraulic diameter at measuring point 34 is 1.56 mm and at measuring point 35 1.3 mm.

Seen from the measuring point 30, the pressure loss of the heat exchanger block with a decreasing hydraulic diameter increases considerably. While the pressure loss for the hydraulic diameter of 2.5 mm is approximately 120 mbar, it is almost twice for the hydraulic diameter of 1.56 mm with approximately 200 mbar.

On the other hand, the fact that the heat transfer capacity in the pipes can be increased significantly and thus the total coolant mass flow can be reduced, the flow rate in the other components of the cooling circuit is reduced, so that instead of about 270 W at the base point, only about 120 W hydraulic power are required. This is made possible by using a heat exchanger according to the invention without having to change other components in the cooling circuit. The result is surprising, since an increase in pressure loss in the heat transfer block can significantly reduce the pump output and thus the use of primary energy to about half.

If, in addition to changing the heat transfer network, the periphery is adapted by reducing the flow resistances, for example in the engine, the thermostat, the hoses, the water boxes or the like, the lines drawn in FIGS. 5 and 6 result in accordance with the percentages of the peripheral pressure loss 37, where 100% corresponds to the series condition.

With a hydraulic diameter of 1, 56 mm and a pressure loss in the periphery of 80% compared to today's series, this results in measuring point 38, at which the necessary hydraulic power in the cooler circuit is now around 100 watts.

If the pressure loss in the other peripheral components is further reduced to a value of 40%, measurement point 39 results, at which the necessary hydraulic power in the cooler circuit is only between 60 and 70 watts.

Also interesting is the entered state point 40, which was calculated with a pressure loss in the periphery of 0%. Measuring point 40 indicates a pressure loss of approximately 200 mbar in the heat transfer block. It follows from the diagram that the hydraulic power required to overcome the flow resistance in the heat transfer block is just under 30 watts.

From the difference in hydraulic power at operating point 34 and

Operating point 40 results in the necessary hydraulic power in order to

Flow resistances in the. To overcome periphery. With about 120 watts as total power and about 30 watts as flow loss in Heat transfer block determines the flow loss for a cooling circuit according to point 34 in the periphery to about 90 watts.

In a conventional cooling system, the power loss in the periphery is approximately 270 watts total loss minus 25 watts block loss and thus approximately 250 watts.

While the flow power loss in the heat transfer block has been significantly increased by reducing the hydraulic diameter from about 2.5 to about 1.5 mm, the total power loss in the heat transfer network drops considerably since the flow loss in the overall system can be reduced by the required lower coolant mass flows in the cooling system according to the invention.

An electrically operated pump can also be used here to transport the coolant in the cooling circuit.

The dimensioning of the pump 6 in a cooling circuit 1 depends on the critical operating conditions. The pump must be designed in such a way that it can ensure reliable cooling of the motor even in critical operating situations.

If an electric pump is provided, a pump that would be one or two classes larger than would be a pump that can be used in a cooling system according to the invention would have to be used with the hydraulic outputs required today.

High pump outputs of up to 400 or even 600 watts require a larger alternator in the motor vehicle and possibly a change in the on-board voltage used from 12 to 24 volts or 42 volts. In addition, the cross-sections of the wiring and connector plugs and the strength of the fuses have to be adapted to the high electrical currents. The use of an electric pump in the cooling circuit further enables the designer of a motor vehicle to arrange the pump independently of the engine. This leads to design freedom and reduces the volume and weight of the engine block itself. This is important, among other things, with regard to the size, shape and location of crumple zones in automobiles.

The power that an electric pump can provide is independent of the engine speed, so that reliable cooling can be ensured even at low engine speeds.

If an electric coolant pump is installed in future motor vehicles for such cases, this can be dimensioned smaller when using a heat transfer network according to the invention and the mechanical pump can be omitted.

The heat transfer network according to the invention can also be used in secondary circuits, such as in heating or oil circuits. A smaller dimensioning of the main coolant pump or an additional pump located in the corresponding secondary circuit is also advantageous here.

While the required hydraulic pump capacity of the cooling circuit against the pressure loss of the heat transfer block for a tube wall thickness of 0.35 mm was shown in FIG. 5, a similar diagram is shown in FIG. 6, the tube wall thickness for all operating points shown being 0.26 mm is.

It is clearly recognizable that there is a similar course of the pump output required in the overall circuit compared to the pressure loss of the block or the hydraulic diameter of the pipes used in the heat transfer network. When using conventional pipes with a hydraulic diameter of 2.8 mm, as corresponds to the base point 30 shown in this diagram, a high hydraulic power is required.

If the hydraulic diameter is reduced from 2.8 mm in base point 30 to 2.27 mm in measuring point 33, the required hydraulic pump output drops from approx. 300 to approx. 130 watts. If the hydraulic diameter is further reduced to 1.52 mm, the operating state results, as shown in measuring point 34.

The required hydraulic power is around 95 watts and has been reduced to a third of the hydraulic power in base point 30. The drawn curve 48 shows the optimal conditions for different peripheral pressure losses, which were varied between 40% and 120%.

Just as in the diagram according to FIG. 5, the diagram according to FIG. 6 also shows that for hydraulic diameters <approximately 2 mm, the hydraulic power required in the cooling circuit decreases sharply until it reaches an optimum, while the overall losses again occur for even smaller hydraulic diameters increase.

The reason for this is a disproportionately increasing flow loss in the heat exchanger, which can no longer be compensated for by a lower flow loss in the periphery, which is caused by a lower mass flow and thus a lower flow velocity.

The result is an optimal hydraulic diameter <2 mm, and in particular an optimal hydraulic diameter range between approximately 0.5 and 2 mm. The range between about 1 and 1.7 mm is particularly suitable.

5 shows a boundary line 71, which in the application example shows a temperature difference of the coolant of 10 K above the engine anzeichnet. Operating states with higher pressure losses in block than indicated by line 71 indicate temperature differences of the coolant above the engine greater than 10 K. State points on the left (in the orientation of FIG. 5) of line 71 indicate operating states with temperature differences less than 10 K).

Likewise, a boundary line 71 is drawn in FIG. 6, which, as in FIG. 5, indicates a temperature difference of the coolant above the engine of greater than or equal to 10 K. In addition, a boundary line 72 is shown in FIG. 6, which indicates a temperature difference or a temperature gradient of the coolant over the engine of 8 K.

Operating conditions with pressure losses greater than those indicated by the boundary lines 71 and 72 lead to temperature differences over the engine above the respectively specified temperature differences (8 K or 10 K or the like.).

Nowadays, the permissible temperature differences are specified by the manufacturers of the engines or the motor vehicles, so that in general, my operating conditions with pressure losses smaller than the respective temperature limit line are permitted by the manufacturers. Therefore, possible operating states with pressure losses arise (J ^e 6 on the chosen temperature difference limit corresponding to line 71 or 72 in Fig.) Under the direction indicated by the respective boundary line.

At border line 71 according to FIG. 6, this is approximately 340 mbar above the block. With a maximum temperature difference of the coolant above the engine of 8 K according to line 72, it is about 210 mbar in the example shown here. However, it should be noted that at other temperature differences or other exemplary embodiments, other values can occur.

The diagram shown in FIG. 7 was created under the same conditions (cooling capacity, tube wall thickness = 0.26 mm). It shows the hydraulic power in the cooling circuit above the fin density on the outside of the pipes for three different pipe spacings or pipe divisions in the heat transfer network.

The operating status line 41 was determined for pipe distances from one pipe to the next of 9.3 mm, the operating status line 42 for pipe distances of 7.3 mm and the operating status line 43 for pipe distances of 5.8 mm. If the fin density is increased to more than 65 fins per dm pipe length, then with a fin spacing of 9.3 mm there is initially a reduction in the required pump output at about 70 to 75 fins per dm, while it rises again to higher fin densities.

The required hydraulic output then rises again because, among other things, the free flow cross-section of the cooling air is restricted due to the high fin density, and therefore the heat output transferred is reduced. As a result, the coolant flow in the heating network must be increased so that the flow losses in the cooling circuit increase.

Conversely, if the fin density is reduced, the heat output transmitted to the outside is reduced and the flow rate of the coolant must be increased again.

For pipe distances of 7.3 and 9.3 mm there is an optimal range between about 70 and 75 fins per dm pipe length.

In a ^'pipe spacing of 5.8 mm, the available free

■ Cross-section for the cooling air, even with small fin densities, is smaller in comparison to the other pipe spacings, namely 5.8 mm minus 1.3 mm the same 4.5 mm. The free distance is 7.3 mm with a pipe distance of 6.0 mm.

In order to keep the heat transfer performance constant, the flow rate in the coolant pipes must be increased so that the heat transfer coefficient becomes higher. This leads to a necessary hydraulic performance that is above the performance of the other pipe distances.

For the operating state shown in FIG. 7, the optimal result would be an outer fin density of about 73 fins per dm with a pipe spacing of 9.3 mm.

However, the other parameters must also be considered when designing a cooling system. This also includes the cooling air throughput.

In a motor vehicle, the cooling air flowing through the radiator is not only used to cool the coolant cooler, but can also be used to cool other circuits, such as the air conditioning circuit.

It is therefore important that changes to the heat exchanger do not change the cooling air throughput too much, in particular reduce it. An increase can (but does not have to) be positive.

An overview of the change in the cooling air throughput is shown in FIG. 8. In this diagram, the cooling air throughput was plotted against the outer fin density for the pipe distances shown in FIG. 7.

Condition point 54 corresponds to condition point 44 in FIG. 7 and was determined for a pipe spacing of 9.3 mm and a number of fins of 65 per dm pipe length.

At higher fin densities, the cooling air throughput drops sharply at all pipe gaps. It follows that a rib density of 65 per In the example shown here, dm is optimal when using a pipe spacing of 9.3 mm, since on the one hand the cooling air throughput increases by a little more than 1.5%, the hydraulic power in the corresponding condition point 45 being approximately 105 watts.

In summary, there is an optimal design point if a hydraulic diameter that is significantly reduced compared to the prior art is selected in the range between approximately d _{hy r} = 1 mm and 2 mm, preferably 1.1 to 1.8 mm, particularly preferably 1.1 mm to 1.7 mm.

If the fin density, the wall thickness and the pipe spacing are selected appropriately on the basis of such a reduction in the hydraulic diameter, an optimal design point results (possibly also iteratively) for a cooling system according to the invention, which requires significantly less hydraulic pump power than conventionally known in the prior art ,

The pump output can be reduced by 20, 50, 75 or even more percent, depending on the design of the heat transfer network and the other components of a cooling circuit.

The choice of a hydraulic diameter is limited by the fact that on the one hand the flow losses in the transmission network become too large and on the other hand that the coolant mass flow drops too much. With low coolant mass flows, the temperature difference between the outlet and inlet temperatures at the engine to be cooled increases because the overall heat output transmitted remains constant.

So far, the engine manufacturers have set predetermined limits in the range from about 8 to about 10 K, so that the minimum coolant mass flow, as shown in FIG. 6, is fixed. Under certain circumstances, however, this limit can be increased without any risk to the motors, so that smaller hydraulic diameters also appear possible. 9 shows the proportion of the pressure loss in the heat transfer network, based on the total pressure loss of the cooling circuit (is equal to the head of the coolant pump) in the operating state over the hydraulic diameter of the heat transfer pipes.

Lines 61 to 65 each indicate the relative pressure loss in the periphery of the cooling circuit, that is to say the pressure loss outside the heat transfer network.

Line 64 marks the periphery of a current system, while line 61 represents a peripheral pressure loss of 40% compared to a current system. Line 62 represents 60%, line 63 80% and line 65 a loss of 120% in the periphery compared to today's system.

Today's heat transfer networks usually have pipe devices with a hydraulic diameter greater than or equal to 2.8 mm. In today's cooling system, the pressure loss of the heat transfer network measured at the total pressure loss in the cooling circuit is 10% (d _hy dr = 2.8). When using a heat transfer network according to the invention, with a hydraulic diameter of 2 mm, the pressure loss when using today's peripheral components is at least 12% and thus significantly higher.

Also advantageous is a share of the heat transmission network in the pressure drop of the entire cooling circuit of 20%, the share being particularly advantageous is 25%, 30%, 40% or more.

Systems according to the invention have relatively large pressure losses in the heat transmission network. This enables smaller coolant mass flows at higher coolant speeds in the heat transfer network and at the same time higher heat transfer on the inside in the heat transfer network. Outside the heat exchanger, the flow rate can be reduced, so that the total hydraulic power required is reduced. For the optimal test vehicle resulting from FIGS. 6 to 8, there is an optimal design point of a hydraulic diameter d _h ydr = 1.52 and a loss share of 30% if the losses in the periphery are reduced to 80% of the current level. This is possible, for example, with changes to the pipes, water boxes, etc.

The pressure loss in the cooling circuit of today's series is, for example, 4% in the hoses, 15% in the water tanks, 9% in the cooler network, 21% in the series-connected thermostat and 51% in the engine block.

When using a heat exchanger according to the invention without changing the periphery, the losses in the engine block, the series-connected thermostat, the water tanks and the series-connected hoses would remain the same at the same flow speeds.

Due to the improvement in heat transfer, the coolant mass flow can be reduced, as a result of which the flow rate is reduced. This reduces the absolute loss in the periphery, while the pressure loss in the cooler network increases from, for example, 9% to 20% or even 30%.

Not only can the relative pressure drop in the cooler network increase, but also the absolute pressure drop. This increase in pressure loss is more than compensated for in an optimal system by the lower pressure losses in the other components.

Claims

Patent claims

1. heat exchanger (2), in particular for use in a cooling circuit (1) of a motor vehicle, with at least one heat transfer network (3), the heat transfer network comprising at least one pipe device (11), and wherein the pipe device has a characteristic hydraulic diameter, which is less than or equal to 2.0 mm.

2. Heat exchanger according to claim 1, characterized in that at least a plurality of substantially similar pipe devices (11) is provided.

3. Heat exchanger according to claim 1 and / or claim 2, characterized in that at least the cross section (23) at least one pipe device is taken from a group of cross-sectional shapes, which are round and circular, elliptical, oval, angular, triangular, rectangular, square and rounded Modifications of the cross-sectional shapes mentioned includes.

4. Heat exchanger according to at least one of the preceding claims, characterized in that the cooling medium flows substantially perpendicular to the direction of flow of the coolant.

5. Heat exchanger according to at least one of the preceding claims, characterized in that a characteristic cross-sectional ratio of the depth of a pipe device (12) to the width of a

Pipe device (13) is between 1 and 100, preferably between 7 and 50.

6. Heat exchanger according to at least one of the preceding claims, characterized in that the coolant contains water as an essential component.

7. Heat exchanger according to at least one of the preceding claims, characterized in that the cooling medium is a gas and preferably air.

8. Heat exchanger according to at least one of the preceding claims, characterized in that at least one pipe device (11) has a characteristic hydraulic diameter which is less than or equal to 1.8 mm or less than or equal to 1.7 mm or less than or equal to 1.6 mm or less.

9. cooling system (1) with at least one pump device (6); at least one heat exchanger device (2) which comprises at least one heat transfer network device (3); at least one heat source device (5), wherein the pump device (6), the heat exchanger device (2) and the heat source device (5) are arranged in a substantially closed cooling circuit (1) through which a coolant flows;

wherein a pressure loss in the heat transfer network device (3) of the heat exchanger device is at least 12% of the pressure loss in the coolant circuit.

10. Cooling system according to claim 9, characterized in that the pressure loss of the heat transfer network device (3) is greater than 20% of the pressure loss in the cooling circuit (1).

11. Cooling system according to, claim 9 and / or claim 10, characterized in that the heat transfer network device (3) comprises pipe devices (11), at least one pipe device (11) being taken from a group of pipe devices, the pipes with hydraulic diameters <2 mm , Dimple pipes,

Pipe devices with turbulence insert devices and the like includes more.

12. Cooling system according to at least one of claims 9 to 11, characterized in that the coolant flow is deflected at least once.

13. Cooling system according to at least one of claims 9 to 12, characterized in that an electrically operated pump device is provided for conveying the coolant.