JP2005527764A - Heat exchanger and cooling system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a heat exchanger capable of reducing the power demand for a pump of a cooling circuit with the same heat transfer characteristics. In addition, a cooling system is provided that allows for lower overall pump power to overcome the flow resistance in the refrigerant circuit.
A heat exchanger, in particular a heat exchanger (2) for use in a motor vehicle cooling circuit (1), comprising at least one heat transfer network (3), the heat transfer network comprising: A heat exchanger comprising at least one tube mechanism (11), the tube mechanism having a characteristic hydrodynamic diameter of 2.0 mm or less.

Description

  The present invention relates to heat exchangers and cooling systems, particularly for use in motor vehicles. It will be pointed out that although the invention is described below in connection with its use in motor vehicles, the heat exchanger according to the invention and the cooling system according to the invention can also be used in other cooling processes.

  In the state of the art, cooling systems are used in motor vehicles, for example, to release waste heat of internal combustion engines to the surroundings. Water is generally used as the refrigerant, and this water contains additives such as antifreeze. The heat released from the engine to the cooling circuit is exhausted to the surroundings by passing an air flow in the refrigerant cooler near the surface of the refrigerant cooler.

  In the meaning of the present application, the refrigerant always represents a medium inside the cooling circuit. In contrast, the cooling medium represents a (external) medium into which the cooling performance of the cooling circuit is released. For example, in an internal combustion engine cooling circuit, engine waste heat is absorbed by refrigerant contained in the cooling circuit. Next, the waste heat of the refrigerant is discharged in a refrigerant cooler to a cooling medium flowing through the refrigerant cooler. In the conventional cooling circuit, the cooling medium is cooling air.

  In order to increase the heat flow, the heat radiating surface of the refrigerant cooler is enlarged, and the refrigerant flow is distributed to a large number of parallel refrigerant tubes provided with cooling fins so that heat is efficiently released to the surroundings.

  In practice, the cooling system heat exchanger must comply with the technical conditions specified by the car manufacturer or the cooling system manufacturer. Thermodynamic data such as the heat output at a predetermined operating temperature of the refrigerant and the surroundings, the maximum pressure loss at a predetermined refrigerant mass flow rate in the operating state, and the like are designated.

  The pressure loss is limited in consideration of the power to be consumed by the refrigerant pump and the dimensional design.

  Typically, the individual components of the cooling circuit are designed for so-called vehicle critical operating conditions, where, for example, when driving uphill under specific load conditions and dominant outside temperatures, the allowable limit temperature is set. Without exceeding, it must be possible to release a certain amount of heat to the surroundings.

  In order to achieve the above criteria, the heat exchanger for the refrigerant circuit in the automobile is designed so that the flow resistance is small at the maximum heat transfer capacity. The maximum flow loss that should not be exceeded is provided on the refrigerant side in the heat exchanger or heat transfer network.

  Although the maximum pressure loss in the refrigerant cooler is set to be small, the pressure loss in the entire cooling circuit is high as a whole.

  The subject of this invention is providing the heat exchanger which can reduce the power demand for the pumps of a cooling circuit in the same heat transfer characteristic. Furthermore, one aspect of the subject of the present invention is primarily to provide a cooling system that allows for smaller pump power as a whole to overcome the flow resistance in the refrigerant circuit.

  The heat exchanger according to the invention is a heat exchanger, in particular a heat exchanger for use in a cooling circuit of a motor vehicle, having at least one heat transfer network, which heat transfer network is at least one tube. A heat exchanger including a mechanism, the tube mechanism having a characteristic hydrodynamic diameter of 2.0 mm or less. The cooling system according to the present invention includes at least one pump mechanism, at least one heat transfer mechanism including at least one heat transfer network mechanism, and at least one heat source mechanism, and includes the pump mechanism, the heat transfer mechanism, and the heat source. Cooling in which the mechanism is disposed in a substantially sealed cooling circuit through which the refrigerant flows and the pressure loss in the heat transfer network mechanism of the heat transfer mechanism is at least 12% of the pressure loss in the refrigerant circuit System. Preferred configurations are the subject of the dependent claims.

  The heat exchanger according to the invention is intended for use in the automotive cooling circuit in particular, but not exclusively. At least one heat transfer network is provided between the refrigerant inlet and the refrigerant outlet of the heat exchanger. The at least one heat transfer network includes at least one pipe mechanism, and the one or more pipe mechanisms are provided to transport the refrigerant in the heat exchanger from the refrigerant inlet to the refrigerant outlet while the waste heat is discharged. ing. In the heat exchanger according to the present invention, the characteristic hydrodynamic diameter of the pipe mechanism is 2.0 mm or less.

In the sense of the present application, the hydrodynamic diameter _dhydrr is defined as four times the cross-sectional area divided by the inner peripheral surface so that it can be used, for example, in a VDI thermal map when a multi-chamber tube is displayed. In the VDI thermal map, the hydrodynamic diameter _dhydrr represents the hydrodynamic diameter of all chambers flowing in parallel. The cross-sectional area here represents the inner cross-sectional area available for refrigerant flow in the following. The inner peripheral surface is a circumference around the flow passage mechanism inside the pipe mechanism.

This means that for a circular ideally formed tube mechanism, the hydrodynamic diameter d _hydrr is equal to the tube mechanism diameter d. In the case of a tube mechanism having a square inner surface, the hydrodynamic diameter d _hydrr corresponds to the inner length of the tube mechanism.

  The heat exchanger according to the invention has many advantages.

An automotive refrigerant circuit heat exchanger known in the state of the art has a hydrodynamic diameter _dhydrr of, for example, 2.8 mm or more. Such a hydrodynamic diameter is selected to minimize flow loss in the heat exchanger.

  In contrast, a relatively small hydrodynamic diameter corresponding to the present invention has a larger pressure loss when the flow rate of the cooling medium is constant, and the heat exchanger according to the present invention has a constant mass flow rate. Results in higher pressure loss. Therefore, the professional industry has not considered studying heat exchangers for automotive refrigerant circuits with relatively small hydrodynamic diameters. This is because the relatively small hydrodynamic diameter can be outside the standards specified by the automobile manufacturer.

  Surprisingly, however, the use of a relatively small hydrodynamic diameter not only results in excessively high flow resistance and thus pressure loss, but at the same time heat transfer inside the cooling medium is improved and the same amount of heat is carried out to the surroundings. It has been found that the refrigerant mass flow rate required for this is reduced as a whole.

  Therefore, the refrigerant mass flow rate in the automobile cooling circuit can be reduced with the heat exchanger according to the present invention. In the case of an incompressible cooling medium such as cooling water, for example, a reduction in refrigerant mass flow directly results in a relatively reduction in refrigerant flow rate in the cooling circuit. Since the flow loss in the cooling circuit is proportional to the square of the refrigerant velocity, reducing the refrigerant mass flow rate to approximately 70% reduces the flow loss by half in the total circuit, resulting in a hydrodynamic discharge rate □ Px means that the volumetric flow rate drops to about 35%.

  When utilizing the heat exchanger according to the present invention, the flow resistance in the heat exchanger is increased by the small hydrodynamic diameter of the tube mechanism in the heat exchanger. The flow rate of the refrigerant can be reduced outside the cooler when the heat output is constant, and the coolant flow rate at the periphery is much lower than in a conventional heat exchanger.

  In this case, the peripheral portion refers to all members and all member regions that allow the refrigerant in the cooling circuit to flow except for the pipe mechanism.

  The pump power that is essential to overcome flow losses in the heat exchanger may be significantly (eg, 2 or 4 times) higher than in conventional heat exchangers, but the higher heat transfer capacity of the heat exchanger As a whole, the amount of refrigerant circulation required for cooling can be reduced. By reducing the pressure loss at the periphery, the required pump power of the pump mechanism is reduced, and the present invention makes it possible to save the energy for operating the pump as a whole (for example, 1.5 times or 2 times). To do.

  When the present invention is used in a circuit having an electric pump, there is little conversion loss from mechanical energy to electric energy, so the primary energy saving effect is great.

  As another advantage of the heat exchanger according to the present invention, when using an electric pump mechanism, an electric pump having a remarkably small electric output can be used, and costs for pumps, batteries, generators, etc. can be saved. .

  A gas, particularly preferably air, is used as a refrigerant for absorbing the amount of heat discharged.

  It should be pointed out that the invention or the advantageous configuration of the invention can also be applied in heating circuits or any cooling circuit. Similarly, the present invention can be applied in parallel circuits or in multi-circuit systems.

  In a preferred configuration of the invention, at least a number of substantially similar tube mechanisms are provided. It is also possible to provide a first majority of the first type of pipe mechanisms and provide a second (third, etc.) majority of the second or more types of pipe mechanisms.

  In another configuration of the invention, at least the cross-section of at least one type of tube mechanism is spherical, circular, elliptical, oval, square, rectangular, triangular, square cross-sectional shape, or rounding modification of the cross-sectional shape. It is selected from a group of cross-sectional shapes including embodiments.

  Mainly, the cross-section of the at least one tube mechanism is substantially constant over at least one length along the tube mechanism. It is preferred to utilize a flat tube mechanism in which the cooling medium flow passages (each across the flow direction of the cooling medium) have a relatively small width and a relatively high depth.

  An abdomen can be provided in the flow passage to improve the compressive strength, for example, the abdomen divides the flat flow passage into rectangular or square or spherical or circular segments. The tube mechanism then includes, for example, a tube segment. Even if the tube mechanism includes segments, it will always be related to the dimensions of the tube mechanism below.

  In a preferred configuration, the tube mechanism can include a turbulence mechanism or fin mechanism within the tube mechanism to enhance turbulence and heat transfer. These fin and turbulence mechanisms do not change the characteristic hydrodynamic diameter.

  In a preferred configuration of the invention, substantially all the pipe mechanisms are arranged substantially parallel to one another, and the cooling medium flows through the pipe mechanisms arranged substantially parallel. A fin mechanism is mainly provided in the tube mechanism, and the fin mechanism can have a housing mechanism in order to enhance heat transfer outside the tube mechanism.

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

  This means that in the flow direction of the cooling medium, the pipe mechanism has a considerably larger extent than the direction perpendicular to this and the direction perpendicular to the refrigerant flow direction. The numerical value can be related to the outer dimension of the tube mechanism or to the inner dimension.

  In one preferred configuration of the one or more of the above-described configurations, the refrigerant may contain water as a main component, and the refrigerant may also have an additive such as an antifreeze liquid or another additive. It is entirely possible that the refrigerant is anhydrous or contains only a small proportion of water. The present invention can also be applied to a heater. Similarly, the present invention can be used to cool or heat, for example, automobile engine oil, transmission oil or fuel. Depending on the application, the refrigerant can have oil or other refrigerant known in the state of the art as a component.

  Mainly gas, particularly preferably air, is used as the cooling medium outside the tube mechanism.

  The cooling system according to the present invention has at least one pump mechanism, at least one heat source mechanism (such as an engine mechanism) and at least one heat transfer mechanism, and the heat transfer mechanism includes at least one heat transfer network mechanism. . The pump mechanism, the heat transfer mechanism, and the heat source mechanism are connected to a substantially sealed cooling circuit, and at least one refrigerant flows therethrough.

  The pressure loss of the heat transfer network mechanism of the heat transfer mechanism with respect to the pressure loss in the entire refrigerant circuit is at least 12%, mainly more than 15%, evaluated before and after the pump mechanism.

  The cooling system according to the present invention has many advantages.

  Mainly, the pressure loss of the heat transfer network mechanism is in the range of 15 to 90%, particularly preferably in the range of 20% to 70% of the pressure loss in the entire refrigerant circuit during operation. Preferably it is at least 30%.

  In a preferred configuration of the cooling system according to the invention, the heat transfer network mechanism comprises a tube mechanism, at least one tube mechanism having a hydrodynamic diameter <2 mm, in particular a hydrodynamic diameter range of 1 to 1.8 mm, It is selected from a group of tube mechanisms including dimple tubes, tube mechanisms with turbulence mechanisms, and the like. Turbulization can be, for example, a (metal) helix or thin film or thread that is brought into the tube mechanism.

  In a preferred configuration of the invention, the refrigerant flow is redirected at least once in the heat transfer mechanism. The refrigerant flow in the heat transfer mechanism can be diverted two times, three times, four times, five times, six times or more.

  Particularly preferably, the heat transfer mechanism has a fin density in the range of 50-120 per decimeter length of the tube mechanism, the thickness of the individual fins being 0.01-0.5 mm, mainly 0.05-0. .2 mm. The higher the fin density, basically, the more heat that can be transferred. However, the large fin density, particularly in combination with the large thickness of the fins, reduces the cross-sectional area available for a cooling medium such as a cooling air flow. The optimum state is obtained from the number, thickness and length of cooling fins for a given cooling fin material.

  Usually the total flow resistance in the cooling circuit is substantially determined by the flow resistance in the connecting hoses, tanks, heat exchanger network, series connected thermostats and engine block.

  It should be pointed out that in this sense the ratio of the pressure loss of the heat transfer network in the total cooling circuit is determined by this definition. If other components are added or missing in another system, the numerical values pointed out here must be converted or adapted accordingly.

  Hereinafter, other advantages and possible applications of the present invention will be described based on embodiments with reference to the drawings.

  An embodiment of a cooling system 1 according to the present invention is shown in FIG. The refrigerant system 1 is scheduled to be used in an automobile and is useful for cooling the engine 5. The warmed refrigerant flowing out of the engine is passed through the thermostat 7 and flows into the tank 4 of the heat exchanger 2.

  For example, when the engine operating temperature is not yet achieved immediately after starting the vehicle, the refrigerant can be guided to the side of the heat exchanger 2 via the bypass 8 and is again introduced into the engine 5 via the pump 6.

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

  The heat transfer network shown in FIG. 2 has a tube 11 in the form of a flat tube, which has a depth 12 in the flow direction of the cooling medium, which is 32 mm in the case of the selected embodiment. Depending on the required dimensions of the heat transfer network, the depth of the flat tube 11 can be 10, 12, 16, 20, 24, 32 mm, or it can be 40 or 48 mm or a value therebetween. However, other values are possible if the requirements for the heat exchanger require it.

  The flat tube 11 of FIG. 2 has a width 13 of 1.3 mm and a wall thickness 17 of −substantially constant over the circumference of the flat tube−only 0.26 mm. This corresponds to the inner inner dimension with a depth of 31.48 mm and a width of 0.78 mm.

With a free cross-sectional area 23 and an inner peripheral surface 24 (assuming the inner cross section is rectangular), the hydrodynamic diameter in the first order approximation of the above dimensions is D _hydrr = 4 × inner surface / inner peripheral surface = 1. 52 mm. When using a tube of depth 16mm instead of depth 32 mm, hydrodynamic diameter under the precondition is the _D hydr = 1.48 mm.

On the other hand, when the spread in the depth direction is doubled to 64 mm, the hydrodynamic diameter becomes D _hydro = 1.54. This is because the hydrodynamic diameter D _hydrr is substantially affected by the inner width of the individual flat tubes, while large or small depth spreads only affect the hydrodynamic diameter to a small extent over a wide range. Means that.

  By increasing the depth spread of the flat tube, on the one hand, an increase in the heat output transmitted is achieved by means of a larger heat transfer surface, while on the other hand the cross-sectional area increases and thus the refrigerant flow rate decreases. If the refrigerant flow rate remains constant, a larger refrigerant mass flow rate will be transported.

  In order to increase the cooling performance at a constant refrigerant flow rate, it is preferable if the hydrodynamic diameter is reduced by reducing the internal flow width in the tube 11. This increases the heat transfer on the inner surface of the tube, which in some circumstances increases the overall heat transfer capacity.

  In that case, the refrigerant mass flow rate can be reduced at the same heat output of the cooler. This results in a decrease in the flow rate of the refrigerant and thus a reduction in the flow loss in a constant flow cross section.

  The spacing 14 between the individual tubes 11 is 9.3 mm in the case of the embodiment of FIG. The fin height of the fin 15 is 8 mm. In order to increase the heat transfer capability, the fin is provided with a housing 16, and thus a boundary layer is constantly formed anew.

  In the embodiment of FIG. 2, the wall thickness 17 of the tube 11 is 0.26 mm. Smaller or larger wall thicknesses such as 0.35 mm are possible as well. Wall thickness tends to be reduced in order to save weight and material and improve heat transfer resistance. However, the minimum wall thickness also depends on the pressure inside the system.

  In the embodiment of FIG. 2, the fins 15 are brazed to the tube 11, while in the embodiment of FIG. 3, the fins are mechanically secured or tightened. The fin element 15 is mounted on a circular tube 21 in the embodiment of FIG. Subsequently, the tube 21 is stretched to obtain a larger outer diameter. The fin element 15 is firmly held by the tube 21.

  Usually heat transfer is improved for brazed joints. Therefore, in the embodiment of FIG. 3, a brazed joint can be provided between the fin and the tube by using a circular tube in the heat exchanger.

In the embodiment of FIG. 3, the inner diameter 18 of the tube 21 is equal to the hydrodynamic diameter _Dhydrr . The tube mechanism 23 has an inner diameter 18. The wall thickness 25 is shown in FIG.

In the heat transfer network shown in FIG. 4, a so-called radius pipe 21 is used. The tube 21 has a depth 12 and a maximum width 13 (in the direction of cooling air flow). The hydrodynamic diameter is also calculated from the inner flow area 23 and the inner peripheral surface 24, and D _hydrr = 4 × flow cross-sectional area 23 ÷ inner peripheral area 24. The inner peripheral area 24 and the flow cross-sectional area 23 can be calculated by knowing the depth 12, the maximum width 13, the tube thickness 25, and the geometric contour.

  The individual tubes 21 are arranged at intervals 14 in the lateral direction. As in the embodiment of FIG. 3, two rows of heat transfer tubes 21 having a tube row spacing 19 are provided.

  It is also possible to have fewer, i.e. one tube row, or more, e.g. 3, 4 or 5 tube rows. The individual tube rows can be arranged in a straight line in the direction of air flow or shifted from each other.

  FIG. 5 is a diagram showing the relationship between the hydrodynamic power of the cooling circuit and the pressure loss of the heat transfer block. The heat transfer block consists of a heat transfer net and a bottom. The heat transfer network is composed of a refrigerant pipe provided with cooling fins.

  This diagram was created for a medium-sized vehicle equipped with, for example, a 1.7 l diesel engine.

  The station 30 written represents today's series model, where about 270 W of hydrodynamic power is required to provide the refrigerant mass flow that is essential for the engine 5 and the cooler 7 during certain operating conditions. .

  All written stations are calculated for a constant cooling performance. The wall thickness of the heat transfer net tube was 0.35 mm in FIG. 5 and 0.26 mm in FIG. Stations 33, 34, 35 are derived for a cooling system in which the heat exchangers have different heat transfer networks.

  Station 30 represents the current state of the art, where the hydrodynamic diameter of the tube utilized is large, in this case about 2.5 mm, whereas the hydrodynamic diameter of station 33 is 1. Reduced to 94 mm. The hydrodynamic diameter at station 34 is 1.56 mm and 1.3 mm at station 35.

  From the point of view of the station 30, the pressure loss of the heat transfer block increases significantly as the hydrodynamic diameter decreases. The pressure loss at a hydrodynamic diameter of 2.5 mm is about 120 mbar, whereas the pressure loss at a hydrodynamic diameter of 1.56 mm is already about double, about 200 mbar.

  On the other hand, the ability to significantly increase heat transfer in the tube, and thus reduce the total refrigerant mass flow, reduces the flow velocity in the remaining components of the cooling circuit, instead of about 270 watts at the base point. Only about 120 W of hydrodynamic power is required.

  This is possible by utilizing the heat exchanger according to the present invention without having to change other components in the cooling circuit. This result is surprising. This is because, by increasing the pressure loss in the heat transfer block, the pump power, and therefore the primary energy input, can be significantly reduced to almost half.

  Complementing the heat transfer network, the perimeter is adapted, that is, for example, the flow resistance in an engine, thermostat, hose, tank or the like is reduced, and the lines written in FIGS. Depending on the percentage value of the pressure loss 37, 100% corresponds to the series state.

  If the hydrodynamic diameter is 1.56 mm and the pressure loss in the periphery is 80% compared to today's series, a station 38 where the required hydrodynamic power in the cooling circuit is only about 100 watts is can get.

  Lowering the pressure loss in the other components of the periphery further to a value of 40% gives a station 39 where the required hydrodynamic power in the cooling circuit is only 60-70 watts.

  The written state point 40 is also of interest and was calculated with 0% pressure loss in the periphery. Station 40 shows a pressure loss of about 200 mbar in the heat transfer block. It can be seen from the diagram that the hydrodynamic power required to overcome the flow resistance in the heat transfer block is only 30 watts.

  The difference between the hydrodynamic power at operating point 34 and the hydrodynamic power at operating point 40 reveals the hydrodynamic power necessary to overcome the flow resistance in the periphery. With about 120 watts as total power and about 30 watts as flow loss power in the heat transfer block, the flow loss power for the cooling circuit after point 34 in the periphery is estimated to be about 90 watts.

  In conventional cooling systems, the power lost in the periphery is about 270 watts total minus 25 watts block loss, thus about 250 watts.

  By reducing the hydrodynamic diameter from about 2.5 to about 1.5 mm, the flow loss power in the heat transfer block is significantly increased while the total power loss in the heat transfer network is significantly reduced. This is because the flow loss in the total system can be reduced by reducing the required refrigerant mass flow rate in the cooling system according to the present invention.

  The use of an electric pump is also considered here for transporting the refrigerant in the cooling circuit.

  The dimensional design of the pump 6 in the cooling circuit 1 depends on the critical operating conditions. The pump must be designed to ensure reliable cooling of the engine even in critical operating conditions.

  If an electric pump is provided, it will be one or two grades larger than the pump available in the cooling system according to the present invention that must be utilized in the hydrodynamic power required today.

  High pump power, less than 400 watts or even less than 600 watts, requires a larger generator in the car and, depending on the situation, requires changing the onboard voltage from 12 volts to 24 volts or 42 volts. In addition, harness and connector cross sections and fuse strength must be adapted to high currents.

  In addition, the use of an electric pump in the cooling circuit allows the vehicle designer to place the pump regardless of the engine. This provides design freedom and reduces the volume and weight of the engine block itself. This is especially important when considering the size, shape and position of the vehicle deformation zone.

  The power that can be used by the electric pump is not affected by the engine speed, and reliable cooling can be ensured even at a low engine speed.

  If an electric refrigerant pump is to be assembled in a future vehicle for such a case, the pump can be further dimensioned using the heat transfer network according to the present invention and the mechanical pump can be omitted. it can.

  The heat transfer network according to the present invention can be used in a sub circuit such as a heating circuit or an oil circuit. Even in that case, it is advantageous to design the main refrigerant pump or the auxiliary pump in the sub-circuit to a smaller size.

  FIG. 5 shows the required hydrodynamic pump power of the cooling circuit in relation to the pressure loss of the heat transfer block for a tube wall thickness of 0.35 mm, whereas FIG. 6 shows all the operating points written. A similar diagram is shown with a tube thickness of 0.26 mm.

  As can be clearly seen, a similar curve of the required pump power in the entire circuit is obtained for the block pressure drop or the hydrodynamic diameter of the tube used in the heat transfer network.

  When using a conventional tube having a hydrodynamic diameter of 2.8 mm corresponding to the base point 30 on which this diagram is written, high hydrodynamic power is required.

  As the hydrodynamic diameter is reduced from 2.8 mm at base point 30 to 2.27 mm at station 33, the required hydrodynamic pump power is reduced from about 300 to about 130 watts. If the hydrodynamic diameter is further reduced to 1.52 mm, the operating condition as shown at station 34 is obtained.

  The required hydrodynamic power was about 95 watts and was reduced to 1/3 of the hydrodynamic power at the base point 30. The written curve 48 shows the optimal conditions for different ambient pressure losses varying between 40% and 120%.

  As in the diagram of FIG. 5, in the diagram of FIG. 6, for the hydrodynamic diameter <about 2 mm, the required hydrodynamic power is strongly reduced in the cooling circuit until it reaches an optimum state. It becomes clear that for even smaller hydrodynamic diameters, the total loss rises again.

  The cause is a flow loss that rises proportionally in the heat exchanger, which can no longer be offset by a lower mass loss caused by a lower mass flow rate and hence a lower flow rate in the periphery.

  As a result, the optimum hydrodynamic diameter is <2 mm, in particular the optimum hydrodynamic diameter range is about 0.5-2 mm. A range of about 1 to 1.7 mm is particularly suitable.

  The limit line 71 written in FIG. 5 shows the refrigerant temperature difference of 10K on the engine in the application example. An operating condition with a higher pressure loss in the block than indicated by line 71 indicates that the refrigerant temperature difference on the engine is greater than 10K. The left state point (in the direction of FIG. 5) of line 71 represents an operating state with a temperature difference of less than 10K.

  Similarly, the limit line 71 written in FIG. 6 indicates a refrigerant temperature difference of 10K or more on the engine, as in FIG. In addition, the limit line 72, which is still shown in FIG. 6, indicates an 8K refrigerant temperature difference or temperature gradient on the engine.

  An operating condition with a greater pressure loss than that due to the limit lines 71, 72 results in a temperature difference above the specified temperature difference (such as 8K or 10K) on the engine each time.

  Today, the allowable temperature difference is specified by the engine or automobile manufacturer, and operating conditions that have a pressure drop below each temperature limit line are generally allowed by the manufacturer. Therefore, an operating state is possible that has a pressure loss equal to or less than that indicated by the limit line (depending on the temperature difference limit selected for line 71 or 72 in FIG. 6).

  This is about 340 mbar on the block for the limit line 71 in FIG. When the maximum temperature difference of the refrigerant on the engine is 8K by the line 72, the pressure loss is about 210 mbar in the example shown here. However, it should be pointed out that other temperature differences or other values may appear in other embodiments.

  The diagram shown in FIG. 7 was created under the same preconditions (cooling performance, tube thickness = 0.26 mm). There, the hydrodynamic power in the cooling circuit is shown in relation to the fin density on the tube outer surface for three types of tube spacing or tube pitch in the heat transfer network.

  The operation state line 41 is calculated for a tube interval from one tube to the next tube, 9.3 mm, the operation state line 42 is calculated for a tube interval of 7.3 mm, and the operation state line 43 is calculated for a tube interval of 5.8 mm. When the fin density is increased to more than 65 fins per tube length / diameter, the required pump power decreases at about 70-75 fins / diameter at fin spacing of 9.3 mm, while the fin density increases again to a higher level. To do.

  The required hydrodynamic power then rises again. This is because, inter alia, due to the high fin density, the free-flow cross section of the cooling air is narrowed and therefore the heat output transferred is reduced. This has to increase the refrigerant flow in the heat transfer network, thus increasing the flow loss in the cooling circuit.

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

  When the tube interval is 7.3 mm and 9.3 mm, the optimum range is about 70 to 75 fins / diameter / tube length.

  At a tube spacing of 5.8 mm, the available cooling air free cross section is small compared to another tube spacing even at low fin density, ie 5.8 mm-1.3 mm = 4.5 mm. On the other hand, when the tube interval is 7.3 mm, the free interval is 6.0 mm.

  In order to keep the transmission capacity constant, the heat transfer coefficient must be increased by increasing the flow velocity in the refrigerant pipe. As a result, the required hydrodynamic power is above the power of the other tube spacing.

  In the operating state shown in FIG. 7, an outer fin density of about 73 fins / diameter will result in an optimal result when the tube spacing is 9.3 mm.

  However, other parameters must be considered when designing the cooling system. This includes, inter alia, the cooling air flow rate.

  In an automobile, the cooling air flowing in the cooler is not only used to cool the refrigerant cooler, but can also be used to cool other circuits such as an air conditioner circuit, for example.

  Therefore, it is important that the flow rate of the cooling air does not change excessively and is not particularly reduced due to various changes in the heat exchanger. The increase may be positive (but not necessarily).

  FIG. 8 shows an outline of the change in the cooling air flow rate. In this diagram, the cooling air flow rate is shown in relation to the outer fin density for each of the tube intervals shown in FIG.

  The state point 54 corresponds to the state point 44 in FIG. 7, and is obtained with respect to the tube interval of 9.3 mm and the number of fins 65 / diameter / tube length.

  With high fin density, the cooling air flow rate decreases strongly at all tube intervals. As will become apparent, the 65 / diameter fin density in the example shown is optimal when using a tube spacing of 9.3 mm. This is because, on the other hand, the cooling air flow rate is increased by a little over 1.5%. The hydrodynamic power at the state point 45 is about 105 watts.

If summary, significantly reduced hydrodynamic diameter as compared to the state of the art is about _d hydr = 1 mm to 2 mm, the range preferably 1.1～1.8Mm, special Preferably the 1.1~1.7mm The optimal design point is obtained when selected within.

  Starting from such a reduction in hydrodynamic diameter, if fin density, wall thickness and tube spacing are suitably selected, the state of the art requires much less hydrodynamic pump power than previously known. The optimum design point for the cooling system according to the present invention can be obtained (repeatedly depending on the situation).

  The pump power reduction can be 20%, 50%, 75% or more depending on the design of the heat transfer network and other components of the cooling circuit.

  The choice of hydrodynamic diameter is limited on the lower side, on the one hand by excessive flow losses in the heat transfer network and on the other hand by excessively reducing the refrigerant mass flow. When the refrigerant mass flow rate is small, the temperature difference between the outlet temperature and the inlet temperature of the engine to be cooled increases because the heat output transmitted as a whole is constant.

  In this regard, conventionally, the engine manufacturer has set a specified limit within the range of about 8 to about 10K, and the minimum refrigerant mass flow rate as shown in FIG. 6 has been established. However, this limit can be increased without danger to the engine in some circumstances, and even smaller hydrodynamic diameters are possible.

  FIG. 9 shows the ratio of the pressure loss in the heat transfer network to the total pressure loss of the cooling circuit in the operating state (this is equal to the discharge head of the refrigerant pump) in relation to the hydrodynamic diameter of the heat transfer tube. is there.

  Lines 61 to 65 represent the pressure loss rate at the periphery of the cooling circuit, that is, the pressure loss outside the heat transfer network.

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

Today's heat transfer networks usually have a tube mechanism with a hydrodynamic diameter of 2.8 mm or more. For today's refrigeration systems, the pressure loss in the heat transfer network is 10% (d _hydr = 2.8) measured in the total pressure loss in the cooling circuit. Using the heat transfer network according to the invention, the pressure loss is at least 12% when using today's peripheral members at a hydrodynamic diameter of 2 mm and is therefore quite high.

  A proportion of 20% of the heat transfer network in the total cooling circuit pressure drop is also advantageous, with particular preference being 25%, 30%, 40% or more.

  The system according to the present invention has a relatively large pressure loss in the heat transfer network. This allows a smaller refrigerant mass flow rate when the refrigerant speed in the heat transfer network is high and at the same time the heat transfer inside the heat transfer network is high. The flow rate can be reduced outside the heat exchanger, reducing the overall required hydrodynamic power.

6-8, when the loss at the periphery is reduced to 80% of today's present situation, the optimum design point of hydrodynamic diameter _dhydrr = 1.52 and the loss of 30% A percentage is obtained. This is possible, for example, by various changes such as pipes and tanks.

  The pressure loss in today's series cooling circuits is, for example, 4% in the hose, 15% in the tank, 9% in the cooling network, 21% in the thermostat connected in series, and 51% in the engine block.

  When using the heat exchanger according to the present invention without changing in the periphery, the pressure loss is constant in the engine block, in the thermostat connected in series, in the tank and in the hose connected in series at the same flow rate. Let's stay at.

  Due to improved heat transfer, the refrigerant mass flow rate can be reduced, thereby reducing the flow rate. This reduces the absolute loss at the periphery, while the pressure loss in the cooling network increases, for example from 9% to 20% or even 30%.

  At that time, not only the pressure loss rate increases in the cooling net, but also the absolute pressure loss may increase. This increase in pressure loss is overcompensated in the optimum system by a small pressure loss in other components.

1 is a simplified cooling circuit according to the present invention. 1 shows a first heat transfer network for a heat exchanger according to the present invention. The heat transfer network for the 2nd heat exchanger concerning the present invention is shown. The heat transfer network for the 3rd heat exchanger concerning the present invention is shown. It is a diagram for calculating the optimum hydrodynamic diameter in the first tube thickness. It is another diagram for calculating the optimum hydrodynamic diameter in the second pipe wall thickness. It is a diagram for calculating the optimum fin density at a predetermined tube thickness. The relationship between a cooling air flow rate and the fin density of a different pipe pitch is shown. The relationship between the ratio of the pressure loss in the cooler to the total pressure loss in the cooling circuit and the hydrodynamic diameter is shown.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Cooling system 2 Heat exchanger 3 Heat transfer network 4 Tank 5 Engine 6 Pump 7 Thermostat 8 Bypass 11 Flat tube 12 Depth 13 Maximum width 14 Spacing 15 Fin 16 Housing 17, 25 Thickness 21 Pipe 24 Inner peripheral surface 41, 42 , 43 Operating state line 44, 54 State point 71, 72 Boundary line

Claims (13)

A heat exchanger, for example a heat exchanger (2) for use in a motor vehicle cooling circuit (1), having at least one heat transfer network (3), the heat transfer network having at least one tube A heat exchanger comprising a mechanism (11), wherein the tube mechanism has a characteristic hydrodynamic diameter of 2.0 mm or less. 2. A heat exchanger according to claim 1, characterized in that at least a number of substantially similar tube mechanisms (11) are provided. A group of crossings, wherein at least the cross-section (23) of the at least one tube mechanism comprises a spherical, circular, elliptical, oval, square, triangular, rectangular, square cross-sectional shape, rounding variant of said cross-sectional shape The heat exchanger according to claim 1 and / or claim 2, wherein the heat exchanger is selected from a surface shape. The heat exchanger according to at least one of the preceding claims, characterized in that the cooling medium flows substantially perpendicular to the flow direction of the refrigerant. Heat exchange according to at least one of the preceding claims, characterized in that the characteristic cross-sectional ratio between the depth of the tube mechanism (12) and the width of the tube mechanism (13) is 1 to 100, mainly 7 to 50 vessel. The heat exchanger according to at least one of the preceding claims, characterized in that the refrigerant contains water as a main component. Heat exchanger according to at least one of the preceding claims, characterized in that the cooling medium is a gas, mainly air. The at least one tube mechanism (11) according to at least one of the preceding claims, characterized in that it has a characteristic hydrodynamic diameter of 1.8 mm or less or 1.7 mm or less or 1.6 mm or less or less. Heat exchanger. A cooling system (1) comprising at least one pump mechanism (6); at least one heat transfer mechanism (2) comprising at least one heat transfer network mechanism (3); and at least one heat source mechanism (5) And a pump mechanism (6), a heat transfer mechanism (2), and a heat source mechanism (5) are disposed in a substantially sealed cooling circuit (1) for circulating the refrigerant; A cooling system in which the pressure loss in the heat net mechanism (3) is at least 12% of the pressure loss in the refrigerant circuit. The cooling system according to claim 9, characterized in that the pressure loss of the heat transfer network mechanism (3) is greater than 20% of the pressure loss in the cooling circuit (1). The heat transfer network mechanism (3) includes a tube mechanism (11), and at least one tube mechanism (11) includes a tube with a hydrodynamic diameter <2 mm, a dimple tube, a tube mechanism with a turbulent mechanism, etc. 11. Cooling system according to claim 9 and / or 10, characterized in that it is selected from a group of tube mechanisms. 12. Cooling system according to at least one of claims 9 to 11, characterized in that the refrigerant flow is diverted at least once. The cooling system according to at least one of claims 9 to 12, wherein an electric pump mechanism for transporting the refrigerant is provided.

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