Title:
Heat exchanger and cooling system
Kind Code:
A1


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.



Inventors:
Ambros, Peter (Kornwestheim, DE)
Dreher, Wolfgang (Gerlingen, DE)
Knauf, Bruno (Mount Pleasant, SC, US)
Application Number:
10/508252
Publication Date:
05/05/2005
Filing Date:
03/17/2003
Assignee:
BEHR GmbH & CO. KG
Primary Class:
International Classes:
F01P7/14; F28D1/053; F28F1/02; F01P5/12; (IPC1-7): F28F1/00
View Patent Images:
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Primary Examiner:
CIRIC, LJILJANA V
Attorney, Agent or Firm:
FOLEY & LARDNER LLP (WASHINGTON, DC, US)
Claims:
1. A heat exchanger (2), in particular for use in a cooling circuit (1) of a motor vehicle, with at least one heat transmission network (3), the heat transmission network comprising at least one tube device (11), and the tube device having a characteristic hydraulic diameter which is smaller than or equal to 2.0 mm.

2. The heat exchanger as claimed in claim 1, characterized in that at least a multiplicity of essentially identical tube devices (11) is provided.

3. The heat exchanger as claimed in claim 1, characterized in that at least the cross section (23) of at least one tube device is taken from a group of cross-sectional shapes which comprises round and circular, elliptic, oval, angular, triangular, rectangular, square and rounded modifications of the cross-sectional shapes mentioned.

4. The heat exchanger as claimed in claim 1, characterized in that the cooling medium flows essentially perpendicularly to the flow direction of the coolant.

5. The heat exchanger as claimed in claim 1, characterized in that a characteristic cross-sectional ratio of the depth of a tube device (12) to the width of a tube device (13) is between 1 and 100, preferably between 7 and 50.

6. The heat exchanger as claimed in claim 1, characterized in that the coolant contains water as an essential constituent.

7. The heat exchanger as claimed in claim 1, characterized in that the cooling medium is a gas and preferably air.

8. The heat exchanger as claimed in claim 1, characterized in that at least one tube device (11) has a characteristic hydraulic diameter which is smaller than or equal to 1.8 mm or smaller than or equal to 1.7 mm or smaller than or equal to 1.6 mm or smaller.

9. A cooling system (1) with at least one pump device (6); with at least one heat exchanger device (2) which comprises at least one heat transmission network device (3); and with at least one heat source device (5), the pump device (6), the heat exchanger device (2) and the heat source device (5) being arranged in an essentially closed cooling circuit (1) through which a coolant flows; a pressure loss in the heat transmission network device (3) of the heat exchanger device being a proportion of at least 12% of the pressure loss in the coolant circuit.

10. The cooling system as claimed in claim 9, characterized in that the pressure loss of the heat transmission network device (3) is higher than 20% of the pressure loss in the cooling circuit (1).

11. The cooling system as claimed in claim 9, characterized in that the heat transmission network device (3) comprises tube devices (11), at least one tube device (11) being taken from a group of tube devices which comprises tubes with hydraulic diameters <2 mm, dimple tubes, tube devices with turbulence insert devices and the like.

12. The cooling system as claimed in claim 9, characterized in that the coolant stream is deflected at least once.

13. The cooling system as claimed in claim 9, characterized in that an electrically operated pump device is provided for conveying the coolant.

Description:

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 in terms of use on a motor vehicle, it may be pointed out that the heat exchanger according to the invention and the cooling system according to the invention may also be used for other cooling processes.

In the prior art, cooling systems are used in motor vehicles in order, for example, to discharge the waste heat from the internal combustion engine into the surroundings. The coolant used is generally water which contains additives, such as, for example, antifreeze agents. In the coolant cooler, the heat transferred into the cooling circuit by the engine is discharged into the surroundings, in that an air stream is conducted past the surfaces of the coolant cooler.

In the context of this application, the term “coolant” always designates the medium inside the cooling circuit. By contrast, the term “cooling medium” designates the (outer) medium to which the cooling capacity of the cooling circuit is transferred. In the case, for example, of a cooling circuit of an internal combustion engine, the waste heat from the engine is absorbed by the coolant contained in the cooling circuit. In the coolant cooler, the waste heat from the coolant is then discharged to the cooling medium flowing through the coolant cooler. In a conventional cooling circuit, this is the cooling air.

To increase the heat flow, the heat-transferring surface of the coolant cooler is enlarged, in that the coolant stream is apportioned to a number of parallel coolant tubes, on which cooling ribs are arranged, in order to transfer the heat effectively to the surroundings.

In practice, a motor vehicle manufacturer or the manufacturer of the cooling system stipulates technical conditions which must be adhered to by the heat exchanger of the cooling system. The thermodynamic data, such as the heat capacity at given operating temperatures of the coolant and of the surroundings, and the maximum pressure loss in the case of a predetermined mass flow of the coolant in the operating state are stipulated.

The pressure loss is limited in light of the power output and dimensioning of the coolant pump.

Typically, the individual components of the cooling circuit are designed for what are known as critical operating states of the vehicle, in which, for example in uphill travel, under defined load conditions and at prevailing outside temperatures, a predetermined heat quantity must be capable of being transferred to the surroundings, without the permissible limit temperatures being exceeded.

In order to achieve the abovementioned criteria, heat exchangers for coolant circuits in motor vehicles are designed in such a way that, at the maximum heat transmission capacity, the flow resistance is low. Maximum flow losses on the coolant side in the heat exchanger or heat transmission network are stipulated and should not be exceeded.

Although the maximum pressure loss in the coolant cooler is stipulated to be low, the pressure loss in the entire cooling circuit is high overall.

The object of the invention is to make available a heat exchanger which makes it possible, while having the same heat transmission properties, to lower the power requirement for the pump of the cooling circuit. Preferably, furthermore, one aspect of the object of the invention is to make available a cooling system in which, overall, a lower pump power output for overcoming the flow resistances in the coolant circuit is possible.

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 subclaims.

The heat exchanger according to the invention is provided particularly, but not only, for use in a cooling circuit of a motor vehicle. At least one heat transmission network is provided between the coolant inlet and the coolant outlet of the heat exchanger. At least one heat transmission network comprises at least one tube device, the tube device or tube devices being provided for transporting coolant through the heat exchanger from the coolant inlet to the coolant outlet while the waste heat is being discharged. In a heat exchanger according to the invention, a characteristic hydraulic diameter of a tube device is smaller than or equal to 2.0 mm.

In the context of this application, the hydraulic diameter dhydr is in this case defined as four times the cross-sectional area, divided by the inner circumferential area, as also used, for example, in the VDI heat atlas when multichamber tubes are designated. In the VDI heat atlas, the hydraulic diameter dhydr designates the hydraulic diameter of all the chambers through which the flow passes in parallel. The cross-sectional area here designates below the inner cross-sectional area which is available for the flow of the coolant. The inner circumference is the circumference around the flow duct device inside the tube device.

This means that, in the case of a circular and ideally configured tube device, the hydraulic diameter dhydr is equal to the diameter of the tube device d. In the case of a tube device with a square inner area, the hydraulic diameter dhydr corresponds to an inner side length of the tube device.

The heat exchanger according to the invention has many advantages.

Heat exchangers which have become known in the prior art for coolant circuits of motor vehicles have a hydraulic diameter dhydr which is, for example, around 2.8 mm and above. Hydraulic diameters of this kind are selected in order to minimize the flow losses in the heat exchanger.

By contrast, smaller hydraulic diameters, as according to this invention, have higher pressure losses in the case of a constant flow velocity of the cooling medium so that heat exchangers according to the invention lead to higher pressure losses in the case of a constant mass flow. The experts therefore did not consider researching heat exchangers for the coolant circuit of a motor vehicle in which there are smaller hydraulic diameters, since these may lay outside the standards stipulated by the motor vehicle manufacturers.

Surprisingly, however, it was shown that the use of smaller hydraulic diameters not only leads to higher flow resistances and therefore pressure losses, but at the same time improves heat transmission on the inside of the cooling medium in such a way that, overall, a lower coolant mass flow is necessary in order to transport an identical heat quantity into the surroundings.

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, for example, cooling water, a reduction in the coolant mass flow leads directly to a proportional reduction in the flow velocity of the coolant in the cooling circuit. Since the flow losses in the cooling circuit are proportional to the square of the velocity of the coolant, a reduction in the coolant mass flow to about 70% means a halving of the flow losses in the overall circuit and a lowering of the hydraulic conveying capacity □P times volume flow to approximately 35%.

When a heat exchanger according to the invention is used, the flow resistance in the heat exchanger is increased due to the small hydraulic diameter of the tube devices in the heat exchanger. With the heat capacity remaining the same, the flow velocity of the coolant can be reduced outside the cooler, so that the flow velocity of the cooling medium on the periphery is markedly lower than in a conventional heat exchanger.

Periphery is to be understood, in this context, as meaning all components and component regions through which the coolant flows in the cooling circuit, with the exception of the tube devices.

Although the necessary pump power output for overcoming the flow loss in the heat exchanger can be considerably higher (for example, a factor 2 or 4) than in a conventional heat exchanger, overall the coolant circulation quantity necessary for cooling can be reduced owing to the higher heat transmission capacity of the heat exchanger. By the reduction in the pressure losses on the periphery, the necessary pump power output of the pump device is reduced, so that the present invention makes it possible overall to save energy for operating the pump (for example, a factor 1.5 or 2).

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

A further advantage of the heat exchanger according to the invention is that, when an electric pump device is used, an electric pump with a markedly lower electrical power can be used, so that costs for the pump, battery, dynamo, etc. can be saved.

Gas and, particularly preferably, air are used as coolant for absorbing the discharged heat quantity.

It may be pointed out that the invention or an advantageous development of the invention may also be employed in a heating circuit or in any cooling circuit. The invention may likewise be employed in the case of parallel circuits or even in multicircuit systems.

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

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

Preferably, the cross section of at least one tube device is essentially constant over at least one length portion along the tube device. It is preferred to use flat tube devices in which the flow duct of the cooling medium has a relatively small width and a relatively large depth (in each case transversely to the flow direction of the cooling medium).

To improve the pressure resistance, webs may be provided in the flow duct, which, for example, divide a flat flow duct into rectangular or square or round or circular segments. A tube device then contains, for example, tube segments. Reference is always made below to the dimensions of the tube device, even should the tube device contain segments.

In a preferred development, the tube devices may also be provided with turbulence insert devices or rib devices in the tube devices, in order to increase the turbulence and the heat transfer. The characteristic hydraulic diameter is not changed by these rib devices and turbulence insert devices.

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

In a preferred development of the invention, a characteristic cross-sectional ratio of the depth of a tube device in the flow direction of the cooling medium to the height of a tube 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 tube devices have a substantially greater extent in the flow direction of the cooling medium than in a direction perpendicular to this and to the flow direction of the coolant. The numerical values mentioned may refer to the outer or even the inner dimensions of the tube devices.

In a preferred development of one or more of the developments described above, the coolant contains water as the essential constituent, in which case the coolant may also have additives, such as antifreeze agents and other additives. It is possible, just as well, for the coolant to be waterless or to contain only a small fraction of water. The invention may also be used in the case of heating bodies. It is likewise possible for the invention to be employed for the cooling or heating of engine oil, transmission oil or the fuel of, for example, a motor vehicle. Depending on the application, the coolant may have as a constituent oil or other coolants known in the prior art.

Preferably, gas and, particularly preferably, air are used as cooling medium on the outside of the tube devices.

The cooling system according to the invention has at least one pump device, at least one heat source device (such as, for example, an engine device) and at least one heat exchanger device, the heat exchanger device comprising at least one heat transmission network device. The pump device, the heat exchanger device and the heat source device are interconnected to form an essentially closed cooling circuit and have at least one coolant flowing through them.

The pressure loss of the heat transmission network device of the heat exchanger device in relation to the pressure loss in the entire coolant circuit, evaluated upstream and downstream of the pump device, is at least 12%, preferably more than 15%.

The cooling system according to the invention has many advantages.

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

In a preferred development of the cooling system according to the invention, the heat transmission network device comprises tube devices, at least one tube device being taken from a group of tube devices which comprises tube devices with hydraulic diameters <2 mm and, in particular, in the hydraulic diameter range of between 1 and 1.8 mm, and also dimple tubes, tube devices with turbulence insert devices and the like. Turbulence inserts may be, for example, (metal) spirals or foils or filaments which are introduced into the tube devices.

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

Particularly preferably, a heat exchanger device has a rib density in the range of between 50 and 120 per decimeter length of the tube device, the thickness of the individual ribs being between 0.01 and 0.5 mm, preferably between 0.05 and 0.2 mm. The greater the rib density is, the more heat can basically be transmitted, although a large rib density, particularly in conjunction with a large thickness of the ribs, reduces the available cross-sectional area for the cooling medium, such as, for example, the cooling air stream. An optimum is obtained from the number, thickness and length of the cooling ribs, with the stipulated material for the cooling ribs.

Conventionally, the overall flow resistance in the cooling circuit is determined essentially by the flow resistances in the connecting hoses, the water boxes, the heat transmission network of the heat exchanger, the series-connected thermostats and the engine block.

It may be pointed out that, in the context of this application, the pressure loss fraction of the heat transmission network in the entire cooling circuit is determined by this definition. If, in another system, further components are added or are absent, the numerical values mentioned here must, if appropriate, be recalculated or adapted accordingly.

Further advantages and possibilities of use of the present invention are illustrated by means of exemplary embodiments, with reference to the drawings in which:

FIG. 1 shows a diagrammatically illustrated cooling circuit according to the invention;

FIG. 2 shows a first heat transmission network for the heat exchanger according to the invention;

FIG. 3 shows a heat transmission network for a second heat exchanger according to the invention;

FIG. 4 shows a heat transmission network for a third heat exchanger according to the invention;

FIG. 5 shows a graph for determining an optimum hydraulic diameter in the case of a first tube wall thickness;

FIG. 6 shows a further graph for determining an optimum hydraulic diameter in the case of a second tube wall thickness;

FIG. 7 shows a graph for determining an optimum rib density in the case of a given tube wall thickness;

FIG. 8 shows the cooling air throughput against the rib density for different tube divisions; and

FIG. 9 shows the fraction of the pressure loss in the cooler in the overall pressure loss in the cooling circuit against the hydraulic diameter.

FIG. 1 illustrates an exemplary embodiment of a cooling system 1 according to the invention. The coolant system 1 is provided for use in a motor vehicle and serves for cooling the engine 5. The heated coolant emerging from the engine is conducted through the thermostat 7 and enters a waterbox 4 of the heat exchanger 2.

If the operating temperature of the engine is not yet reached, for example shortly after the starting of the vehicle, the coolant may also be routed past the heat exchanger 2 via the bypass 8 and is conducted into the engine 5 again via the pump 6.

The heat exchanger 2 has a heat transmission network 3. Such a heat transmission network 3 according to the invention is illustrated in various embodiments in FIGS. 2, 3 and 4, in which the arrangement and dimensions of the individual components differ from one another.

The heat transmission network illustrated in FIG. 2 has tubes 11 in the form of flat tubes which have a depth 12 in the flow direction of the cooling medium which is 32 mm in the selected exemplary embodiment. Depending on the required dimension of the heat transmission network, the depth of the flat tubes 11 may also be 10, 12, 16, 20, 24, 32 or else 40 or 48 mm or values between these. However, other values are also possible when the requirements to be met by the heat exchanger necessitate this.

The flat tubes 11 according to FIG. 2 have a width 13 of 1.3 mm with a wall thickness 17 which, essentially constant over the circumference of the flat tube, is only 0.26 mm. This corresponds to clear inner dimensions of 31.48 mm in depth and 0.78 mm in width.

With the dimensions mentioned, the free cross-sectional area 23 and the inner circumference 24 results (on the assumption of a rectangular inner cross section), in a first approximation, in a hydraulic diameter Dhydr=4×inner area/inner circumference=1.52 mm.

When a tube with a depth of 16 mm instead of a depth of 32 mm is used, this results, under the preconditions mentioned, in a hydraulic diameter of Dhydr=1.48 mm.

If, by contrast, the extent in depth is doubled to 64 mm, a hydraulic diameter of Dhydr=1.54 is obtained. This means that the hydraulic diameter Dhydr is influenced essentially by the inner width of the individual flat tubes, whereas a greater or smaller extent in depth influences the value of the hydraulic diameter over wide ranges to only a slight extent.

By an increase in the extent of the flat tubes in depth, on the one hand, the greater heat transmission area achieves an increase in the transmitted heat capacity, while, on the other hand, the flow velocity of the coolant decreases, since the cross-sectional area increases. If the flow velocity of the coolant were to remain constant, a greater mass flow of the coolant would be transported.

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

Then, for the same heat capacity of the cooler, the coolant mass flow can be reduced. With the flow cross section remaining the same, this leads to a reduction in the flow velocity of the coolant and therefore to lower flow losses.

In the exemplary embodiment according to FIG. 2 the spacing 14 between the individual tubes 11 is 9.3 mm.

The rib height of the ribs 15 is 8 mm. To increase the heat transmission capacity, the ribs are provided with gills 16, so that the boundary layers are repeatedly reformed.

In the exemplary embodiment according to FIG. 2, the wall thickness 17 of the tubes 11 is 0.26 mm. Smaller or larger wall thicknesses are also possible, such as, for example, 0.35 mm. The tendency is to reduce the wall thickness in order to save weight and material and to improve the heat conduction 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 soldered onto the tubes 11, whereas, in the exemplary embodiment according to FIG. 3, they are fastened mechanically or clamped. Rib elements 15 are slipped onto the tubes 21 which are circular in the exemplary embodiment according to FIG. 3. The tubes 21 are subsequently expanded, so that a larger outside diameter is obtained. The rib elements 15 are held firmly on the tubes 21.

Conventionally, heat transmission is better in the case of soldered connections. Consequently, even in the exemplary embodiment according to FIG. 3, a soldered connection between rib and tube may be provided if circular tubes are used in the heat exchanger.

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

In the heat transmission network illustrated in FIG. 4, radii tubes 21, as they are known, are used. 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 area 23 and the inner circumference 24 as Dhydr=four×flow cross section 23/inner circumferential area 24. The inner circumferential area 24 and the flow cross section 23 can be determined with a knowledge of the depth 12 and the maximum width 13 and also of the tube wall thickness 25 and the geometric contour.

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

There may also be fewer, that is to say one tube row, or else more, for example 3, 4 or 5 tube rows. The individual tube rows may be arranged so as to be in alignment or in each case offset in the air flow direction.

FIG. 5 illustrates a graph of the hydraulic capacity of the cooler circuit against the pressure loss of the heat transmission block. The heat transmission block consists of the heat transmission network and of the tubesheets. The heat transmission network consists of the coolant tubes with the cooling ribs.

The graph was set up with reference to a motor vehicle of medium size with, for example, a 1.7 liter diesel engine.

The depicted measurement point 30 identifies a present-day series model in which, in a specific operating state, a hydraulic conveying capacity of about 270 W is required, in order to make available the necessary coolant mass throughput for the engine 5 and cooler 7.

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

While the measurement point 30 indicates the current prior art in which the hydraulic diameter of the tubes used is large and is about 2.5 mm here, the hydraulic diameter for the measurement point 33 was reduced to 1.94 mm. The hydraulic diameter is 1.56 mm at the measurement point 34 and 1.3 mm at the measurement point 35.

As seen from the measurement point 30, the pressure loss of the heat exchanger block increases considerably with a decrease in hydraulic diameter. Whereas the pressure loss in the case of the hydraulic diameter 2.5 mm is about 120 mbar, in the case of the hydraulic diameter of 1.56 mm it already amounts to almost double at about 200 mbar.

On the other hand, since the heat transmission capacity in the tubes can be markedly increased and therefore the overall coolant mass flow can be reduced, the flow velocity in the remaining components of the cooling circuit is reduced, so that a hydraulic capacity of only about 120 W is required instead of about 270 W at the base point.

This is made possible by the use of a heat exchanger according to the invention, without further components in the cooling circuit having to be changed. The result is surprising, since, by an increase in the pressure loss in the heat transmission block, the pump power output and consequently the use of primary energy can be markedly reduced to about half.

If, in addition to the change in the heat transmission network, the periphery is adapted in that, there, the flow resistances, for example, in the engine, the thermostat, the hoses, the waterboxes or the like, are reduced, this results in the lines, depicted in FIGS. 5 and 6, corresponding to the percentages of the periphery pressure loss 37, 100% corresponding to the series state.

In the case of a hydraulic diameter of 1.56 mm and a pressure loss on the periphery of 80% in comparison with the present-day series, this results in the measurement point 38 at which the necessary hydraulic capacity in the cooler circuit is about 100 Watt.

With a further reduction in the pressure loss in the further components of the periphery to a value of 40%, this results in the measurement point 39 at which the necessary hydraulic capacity in the cooler circuit is only between 60 and 70 Watt.

The plotted state point 40, which was calculated in the case of a pressure loss on the periphery of 0%, is also useful. The measurement point 40 indicates a pressure loss of approximately 200 mbar in the heat transmission block. It follows from the graph that the necessary hydraulic capacity for overcoming the flow resistance in the heat transmission block is barely 30 Watt.

The difference in the hydraulic capacities in the operating point 34 and the operating point 40 yields the necessary hydraulic capacity for overcoming the flow resistances on the periphery. With an overall capacity of 120 Watt and with a flow capacity loss in the heat transmission block of about 30 Watt, the flow capacity loss for a cooling circuit according to point 34 on the periphery is determined at approximately 90 Watt.

In a conventional cooling system, the capacity loss on the periphery is approximately 270 Watt overall loss minus a block loss of 25 Watt and consequently about 250 Watt.

Whereas the flow capacity loss in the heat transmission block increased markedly due to a reduction in the hydraulic diameter from about 2.5 to about 1.5 mm, the overall capacity loss in the heat transmission network falls considerably, since the flow loss in the overall system can be reduced owing to the lower coolant mass flows necessary in the cooling system according to the invention.

It is also appropriate here to use an electrically operated pump in order to transport the coolant in the cooling circuit.

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

If an electric pump is provided, it would be necessary, in the case of the hydraulic capacities required today, to use a pump which would have to be one to two classes larger than a pump which can be used in a cooling system according to the invention.

High pump capacities of up to 400 or even 600 Watt require a larger dynamo in the motor vehicle and, under certain circumstances, a changeover of the on-board voltage used from 12 to 24 volt or even 42 volt. Moreover, the cross sections of cabling and connecting plugs and the strength of the fuses must be adapted to the high electrical currents.

The use of an electric pump in the cooling circuit, furthermore, enables the designer of a motor vehicle to arrange the pump independently of the engine. This leads to freedoms in structural terms and reduces the volume and weight of the engine block per se. This is important, inter alia, as regards the size, shape and position of crumple zones on the motor vehicle.

The power which an electric pump can make available is independent of the engine rotational speed, so that reliable cooling can be ensured even in the case of low engine rotational speeds.

If an electric coolant pump is installed for such cases in future motor vehicles, this can have smaller dimensioning and the mechanical pump can be dispensed with when a heat transmission network according to the invention is used.

The heat transmission network according to the invention may also be used in secondary circuits, such as, for example, in heating or oil circuits. In this case, too, a smaller dimensioning of the main coolant pump or of an additional pump located in the corresponding secondary circuit is advantageous.

While the required hydraulic pump capacity of the cooling circuit has been illustrated in FIG. 5 against the pressure loss of the heat transmission block for a tube wall thickness of 0.35 mm, FIG. 6 illustrates a similar graph in which the tube wall thickness is 0.26 mm for all the depicted operating points.

It can be seen clearly that a similar profile of the required pump capacity in the overall circuit with respect to the pressure loss of the block or to the hydraulic diameter of the tubes which is used in the heat transmission network is obtained.

When conventional tubes with a hydraulic diameter of 2.8 mm, such as corresponds to the depicted base point 30 in this graph, are used, a high hydraulic capacity is required.

With a reduction of the hydraulic diameter from 2.8 mm at the base point 30 to 2.27 mm at the measurement point 33, the required hydraulic pump capacity falls from approximately 300 to about 130 Watt. When the hydraulic diameter continues to be lowered to 1.52 mm, the operating state, such as is illustrated at the measurement point 34, is obtained.

The required hydraulic capacity is about 95 Watt and was reduced to one third of the hydraulic capacity at the base point 30. The depicted curve 48 shows the optimum conditions for different peripheral pressure losses which were varied between 40% and 120%.

Exactly as in the graph according to FIG. 5, the graph according to FIG. 6 also shows that, for hydraulic diameters <about 2 mm, the required hydraulic capacity in the cooling circuit decreases sharply until it reaches an optimum, whereas, toward even smaller hydraulic diameters, the overall losses rise again.

The cause of this is a flow loss in the heat exchanger which rises superproportionally and which can no longer be compensated by a lower flow loss on the periphery brought about by a lower mass flow and consequently a lower flow velocity.

This results in an optimum hydraulic diameter <2 mm and, in particular, an optimum hydraulic diameter range of between about 0.5 and 2 mm. The range of between about 1 and 1.7 mm is particularly suitable.

FIG. 5 depicts a boundary line 71 which, in the application example, marks a temperature difference of the coolant of 10 K in relation to the engine. Operating states with higher pressure losses in the block than indicated by the line 71 identify temperature differences of the coolant in relation to the engine which are greater than 10 K. State points on the left (in terms of FIG. 5) of the line 71 identify operating states with temperature differences lower than 10 K).

FIG. 6 likewise depicts a boundary line 71 which, as in FIG. 5, indicates a temperature difference of the coolant in relation to the engine of greater than or equal to 10 K. In addition, FIG. 6 also illustrates a boundary line 72 which indicates a temperature difference or a temperature gradient of the coolant in relation to the engine of 8 K.

Operating states with pressure losses greater than those indicated by the boundary lines 71 and 72 lead to temperature differences in relation to the engine which are above the temperature differences indicated in each case (8 K or 10 K or the like).

Nowadays, the permissible temperature differences are stipulated by the manufacturers of the engines or of the motor vehicles, so that, in general, operating states with pressure losses lower than the respective temperature boundary line are permitted by the manufacturers. Consequently, (depending on the selected temperature difference limit according to line 71 or 72 in FIG. 6) possible operating states are obtained, with pressure losses below those indicated by the corresponding boundary line.

In the case of the boundary line 71 according to FIG. 6, this pressure loss amounts to about 340 mbar in relation to the block. In the case of a maximum temperature difference of the coolant in relation to the engine of 8 K according to the line 72, it is about 210 mbar in the example illustrated here. It may be pointed out, however, that other values may occur in the case of other temperature differences or other exemplary embodiments.

The graph illustrated in FIG. 7 was set up under the same preconditions (cooling capacity, tube wall thickness=0.26 mm). This illustrates the hydraulic capacity in the cooling circuit against the rib density on the outside of the tubes for three different tube spacings or tube divisions in the heat transmission network.

The operating state line 41 was determined for tube spacings from one tube to the next of 9.3 mm, the operating state line 42 for tube spacings of 7.3 mm and the operating state line 43 for tube spacings of 5.8 mm. If the rib density is increased to more than 65 ribs per dm tube length, this initially results, in the case of a rib spacing of 9.3 mm, in a lowering of the required pump capacity with about 70 to 75 ribs per dm, whereas it rises again toward higher rib densities.

The required hydraulic capacity then rises again because, inter alia due to the high rib density, the free flow cross section of the cooling air is contracted and therefore the transmitted heat capacity becomes lower. The coolant stream in the heat network must consequently be increased, so that the flow losses in the cooling circuit rise.

If, conversely, the rib density is reduced, the heat capacity transmitted outward is reduced and the flow velocity of the coolant must again be increased.

This results, for tube spacings of 7.3 and 9.3 mm, in an optimum range of between about 70 and 75 ribs per dm tube length.

In the case of a tube spacing of 5.8 mm, the available free cross section for the cooling air, even with low rib densities, is smaller, as compared with the other tube spacings, to be precise 5.8 mm minus 1.3 mm which is equal to 4.5 mm. By contrast, for the tube spacing of 7.3 mm, the free spacing is 6.0 mm.

In order to keep the heat transmission capacity constant, the flow velocity in the coolant tubes must be increased so that the heat transmission coefficient becomes higher. This leads to a necessary hydraulic capacity which lies above the capacity of the other tube spacings.

For the operating state illustrated in FIG. 7, with a tube spacing of 9.3 mm, an optimum result would be an outer rib density of about 73 ribs per dm.

However, the further parameters must also be taken into account in the design of a cooling system. These also include, inter alia, the cooling air throughput.

In a motor vehicle, the cooling air flowing through the cooler is not only used for cooling the coolant cooler, but may also be employed for cooling further circuits, such as, for example, the air conditioning system circuit.

It is therefore important that the cooling air throughput is not varied to too great an extent, in particular reduced, owing to changes to the heat exchanger. An increase may be (but is not necessarily) positive.

An overview of the variation in the cooling air throughput is shown in FIG. 8. In this graph, the cooling air throughput was plotted against the outer rib density in each case for the tube spacings illustrated in FIG. 7.

The state point 54 corresponds to the state point 44 in FIG. 7 and was determined for a tube spacing of 9.3 mm and a rib number of 65 per dm tube length.

Toward higher rib densities, the cooling air throughput falls sharply in the case of all tube spacings. The result of this is that a rib density of 65 per dm is optimum in the example illustrated here when a tube spacing of 9.3 mm is used, since, on the one hand, the cooling air throughput increases by somewhat more than 1.5%, the hydraulic capacity at the corresponding state point 45 being approximately 105 Watt.

In summary, an optimum design point is obtained when a hydraulic diameter markedly reduced in comparison with the prior art is selected in the range of between about dhydr=1 mm and 2 mm, preferably 1.1 to 1.8 mm, particularly preferably 1.1 mm to 1.7 mm.

If the rib density, wall thickness and tube spacing are suitably selected on the basis of such a reduction in the hydraulic diameter, this results (also iteratively under certain circumstances) in an optimum design point for a cooling system according to the invention which requires a substantially lower hydraulic pump capacity than those known conventionally in the prior art.

The reduction in the pump capacity may amount to 20, 50, 75 or even more percent, depending on the design of the heat transmission network and of the further components of a cooling circuit.

The choice of a hydraulic diameter has a lower limit, in that, on the one hand, the flow losses in the transmission network become too great and, on the other hand, the coolant mass flow falls too sharply. In the case of low coolant mass flows, the temperature difference between the outlet and the inlet temperature of the engine to be cooled increases, since the overall transmitted heat capacity remains constant.

The engine manufacturers set stipulated limits here in the range of about 8 to approximately 10 K, so that the minimum coolant mass flow is fixed, as illustrated in FIG. 6. However, under certain circumstances, this limit may be increased without risk to the engines, so that even smaller hydraulic diameters appear possible.

In FIG. 9, the fraction of the pressure loss in the heat transmission network with respect to the overall pressure loss of the cooling circuit (equal to the lift of the coolant pump) in the operating state is illustrated against the hydraulic diameter of the heat transmission tubes.

The lines 61 to 65 here in each case identify the relative pressure loss on the periphery of the cooling circuit, that is to say the pressure loss outside the heat transmission network.

The line 64 identifies the periphery of a present-day system, while the line 61 illustrates a peripheral pressure loss of 40% with respect to a present-day system. The line 62 represents 60%, the line 63 80% and the line 65 a loss of 120% on the periphery, as compared with a present-day system.

Present-day heat transmission networks normally have tube devices with a hydraulic diameter larger than/equal to 2.8 mm. In a present-day cooling system, the pressure loss of the heat transmission network, measured with reference to the overall pressure loss in the cooling circuit, is around 10% (dhydr=2.8). When a heat transmission network according to the invention is used, with a hydraulic diameter of 2 mm, the pressure loss in the case of the use of present-day peripheral components is at least 12% and consequently markedly higher.

A fraction of the heat transmission network in the pressure drop of the overall cooling circuit of 20% is also advantageous, and the fraction amounts particularly advantageously to 25%, 30%, 40% or more.

Systems according to the invention have relatively high pressure losses in the heat transmission network. This makes it possible to have lower coolant mass flows along with a higher coolant velocity in the heat transmission network and, at the same time, with a higher heat transmission on the inside in the heat transmission network. Outside the heat exchanger, the flow velocity can be reduced, so that, overall, the required hydraulic capacity is reduced.

For the optimum test vehicle resulting from FIGS. 6 to 8, an optimum design point with a hydraulic diameter dhydr=1.52 and a loss fraction of 30% are obtained when the losses on the periphery are lowered to 80% of the present-day level. This is possible, for example, by means of modifications to the lines, waterboxes, etc.

The pressure loss in the cooling circuit of a present-day series is, for example, 4% in the hoses, 15% in the waterboxes, 9% in the cooler network, 21% in the series-connected thermostat and 51% in the engine block.

When a heat exchanger according to the invention is used without any modification on the periphery, for the same flow velocities, the losses in the engine block, in the series-connected thermostat, in the waterboxes and in the series-connected hoses will remain the same.

Owing to the improvement in heat transmission, the coolant mass flow can be reduced, with the result that the flow velocity is reduced. As a result, there is a decrease in the absolute loss on the periphery, while the pressure loss in the cooler network increases from, for example, 9% to 20% or even 30%.

In this case, not only the relative pressure loss in the cooler network can increase, but also the absolute pressure loss. This pressure loss increase is overcompensated in an optimum system by the lower pressure losses in the further components.