EP0268831B1 - Plate fin - Google Patents

Abstract

1. Fin (1) made of aluminium or an aluminium alloy for the joint ribbing of several heat-exchanger tubes (14) of a finned tubular heat exchanger (22) in motor vehicles, in which ambient air as a first heat-exchanger fluid is guided along the surface of the fin (1) and a second heat-exchange fluid is guided into the heat-exchanger tubes (14), with a design of the fin (1) undulated in the direction of flow (2) of the first fluid, with joining sleeves (4) formed on in the fin (1) for connection to the heat-exchanger tubes (14), at least one wave crest (5) extending between two joining sleeves (4) mutually adjacent transversely relative to the direction of flow (2) of the first fluid, and with local air-guide profiles formed on from the undulated surface of the fin (1) in interspaces between the joining sleeves (4), characterized in that the air-guide profiles are bosses (6) in the finned tubular heat exchanger (22), in that the bosses (6) are arranged respectively on one flank (20) of the undulation, in that the particular boss (6) is at a distance from the two ends of the flank (20), and in that the heigt (f) of the bosses (6) is at least 15% and at most 80% of the fin spacing (b) in the finned tubular heat exchanger (22).

Description

The invention relates to a lamella made of AI or an AI alloy for the joint ribbing of several heat exchanger tubes of a tube fin heat exchanger in motor vehicles according to the preamble of claim 1. Heat exchanger fins of this type are known from DE-A 2 530 064. They are acted upon by outside air as the first heat exchange fluid, while a second heat exchange fluid is guided in the heat exchanger tubes, which are ribbed by the fins.

Fins for tube fin heat exchangers of motor vehicles are now generally made of Al or Al alloys, in very thin wall thicknesses between typically 0.08 and 0.15 mm. A production of such lamellae from sheet iron is practically out of the question for reasons of thermal conductivity which is four times worse, corrosion and weight reasons. Stainless steel sheet would be corrosion-resistant, but only accounts for about 10% of the thermal conductivity of an AI lamella. Manufacture of such lamellae from Cu would meet the requirements with regard to corrosion resistance or even heat conduction, which is even better, but differs except in exceptional cases, e.g. with some engine coolers or solderable heating heat exchangers, for reasons of weight and the Cu price in comparison with the AI price.

In this sense, heat exchangers for motor vehicles in development as large series parts have been optimized not only according to performance data, but also according to weight, construction volume, use of materials and the like. result, which does not allow a comparison with heat exchangers from other applications with smaller quantities up to the individual production without further ado.

For lamellae of this type, the highest possible heat transfer coefficient between the lamella on the one hand and the gaseous first fluid acting on the latter on the other hand is to be achieved with simple, permanent means. This increase in the heat transfer value brings savings in the investment area and in operation, since with the same amount of heat to be transferred and the same operating temperatures, the heat exchanger front surface (exposed surface) and the structural depth can be reduced or the distance between the heat exchanger fins can be increased. Since the heat exchanger lamella is used for tubular lamella heat exchangers in motor vehicles, the reduction in the construction volume and the associated reduction in the weight of the heat exchanger are of crucial importance. This applies both to the possible application on motor vehicle coolers or heating heat exchangers and to the preferred applications for condensers or evaporators in motor vehicle air conditioning systems.

The heat exchange between the two fluids takes place by means of heat radiation, heat conduction and convection, but in particular by means of convection, in which the heat is transferred by moving material particles. The heat exchange by convection in particular depends largely on the type of flow of the gaseous first fluid around the tubes and the heat exchanger fins.

beschrieben, worin aa die über die Plattenlänge L gemittelte Wärmeübergangszahl vom gasförmigen Fluid an die Lamellen- oder Plattenoberfläche, w die Strömungsgeschwindigkeit des die Lamellen beaufschlagenden gasförmigen Mediums, c eine sich aus den physikalischen Eigenschaften des strömenden Mediums ergebende Konstante sind. Aus der Gleichung ist ersichtlich, daß zur Verbesserung der Wärmeübergangszahl bei Platten entweder die Strömungsgeschwindigkeit erhöht oder die überströmte Länge der Platte L verringert werden muß.It is known that in flows parallel to a plate, a laminar boundary layer forms on the surface, which becomes thicker with an increasing flow path L and hinders the heat exchange by convection, since the heat can only be transferred through molecular conduction processes through this laminar boundary layer. The external heat transfer coefficient is qualitative through the formula

described in which aa is the heat transfer coefficient, averaged over the plate length L, from the gaseous fluid to the lamella or plate surface, w is the flow rate of the gaseous medium acting on the lamellae, c is a constant resulting from the physical properties of the flowing medium. From the equation it can be seen that in order to improve the heat transfer coefficient for plates, either the flow velocity has to be increased or the overflowed length of the plate L has to be reduced.

A further reduction in the heat exchange also results in the flow dead spaces that build up behind pipe areas in the flow direction of the first fluid, ie in their flow shadows. Due to a low-intensity stationary vortex formed by the flow behind the heat exchanger tubes, the local heat transfer numbers there are considerably smaller than in the areas touched by the main flow. Also, as the lamella thickness progressively decreases, the thermal resistances in the lamella have to be taken into account, which results in the requirement for a heat flow density in the lamella that is as uniform as possible with a constant pipe spacing, which in turn is achieved by adapting the local heat transfer coefficients. In order to achieve these requirements, it is known to profile the lamella in various ways. One of the simplest known profiles consists of a corrugated configuration of the lamella in the flow direction of the first fluid, so that the wave crests and troughs run transversely to this flow direction (see, for example, the generic DE-A 2 530 064, but also, for example, the DE-A 2 756 941). This corrugation on the one hand slightly increases the flow path and thus the flow velocity between the fins, and on the other hand, due to the required deflection of the air in the corrugated fins, an at least partial rebuilding of the laminar boundary layer is achieved after each wave crest, thereby enlarging the boundary layer and thus degradation of the external heat transfer coefficient corresponding to the above equation can be at least partially avoided. This is especially true when the wave crests are relatively sharp are formed, in particular as edges of a straight zigzag corrugation (cf. claim 4); in the context of the invention, however, rounded wave crests should also be included.

However, there are limits to the increase in the external heat transfer coefficient compared to a smooth plate, since from a corrugation angle 8 of 15 to 20 _° the heat transfer coefficient only increases slightly, while the air-side pressure loss increases to an increasing extent (Fig. 3). The increase in the external heat transfer coefficient, which is only achieved with a corrugation, is inadequate in the area of a technically sensible ratio of increase in performance to increase in pressure loss, i.e. corrugation angle up to a maximum of 20 _° , since the degradation of the laminar boundary layer by the corrugation is insufficient and, in addition, the increase in surface area and the increase in Flow velocity at a corrugation angle of maximum 20 _{° are} still very low (6%). Corrugation of the lamella surface also means that there is no significant reduction in the dead space behind the pipes and no optimal distribution of the local heat transfer coefficients with respect to a uniform radial heat flow density. Corrugation angles of approx. 45 _° , as shown in the drawing in DE-A 2 530 064, lead to an enlargement of the boundary layer through which the heat has to be transported as a result of molecular conduction of the air, since the air is only used to a small extent flows parallel to the corrugation and a stationary vortex of low intensity forms in each trough. The high surface area enlargement of 30 to 40%, depending on the surface proportion of the tubular pipe connection points, is largely compensated for by the increase in the boundary layer thickness described at the outset, so that experience shows that the increase in the external heat transfer coefficient is only insignificantly greater than at a corrugation angle of 20 _° (Fig . 7).

For non-corrugated slats, new profile shapes have been developed that are pressed out of the slat itself, with openings in the slat material. All these profiles have in common that the formation of a laminar boundary layer of greater thickness is to be prevented as much as possible by the shape of the lamella. A known profile of this type (DE-U 7 806 410) tries to guide the gaseous fluid in a pressed-out guide channel with a semicircular cross-section into the flow shadow behind the connection points. The thermal and hydraulic flow profiles have to be rebuilt due to the leading edges at the beginning and end of a guide channel. In addition to the fact that it is doubtful that the formation of the flow shadow which reduces the flow is impaired by the guide channels, the section edges are also limited to only a small proportion of the area of the lamella, while a large area of the lamella without profile-reducing profiles is designed as a smooth lamella. A more even distribution of lamellar openings and guide bars is given in the known arrangement (DE-A 2 518 226) of a heat exchanger lamella for the chemical industry, in particular the petroleum industry, in which a large number of guide bars of small width are provided between two connecting points in the lamella which is to be forced to constantly rebuild the laminar boundary layer. However, this measure actually appears to be disadvantageous, since the slots in the heat exchanger lamella associated with the outer guide webs then make the flow of heat from the pipes to be accommodated in the connection points to the outer guide webs more difficult, in some cases by a considerably larger flow path, so that a larger temperature difference is required for heat transport and so that the rib efficiency is reduced. Since the outer guide webs are located close to one another in the same plane, the boundary layer formed in the previous slot is also not completely broken down.

In DE-A 3 131 737, which relates to space heating and cooling systems, guide webs are instead formed as roof-shaped strips in the fins of tubular fin heat exchangers and are arranged in a bridge-like manner with respect to one another. This design of the guide webs is better in terms of boundary layer degradation and lamella stability, but there can be no radial heat flow around the heat exchanger tubes, which is at least approximately invariant in angles. The production tools for this lamella are particularly complicated and there are high pressure losses for a given output, in particular in the case of additional condensation of water vapor as a result of the lamella falling below the dew point. In the latter case, the condensation water adheres like a sponge by adhesion between the many guide bars, so that the heat transfer is even worse than with a smooth slat due to the blocking of the lamella surface with condensation water.

According to an improvement (DE-C 3 336 985) of this basic design, the air-side pressure loss has been reduced and the required manufacturing tools have been simplified with a consistently high heat transfer coefficient by simplifying at least one opening and at least on one edge of the opening, which is transverse to the Extending in rows and adjacent to a connection point, a guide web for the gaseous fluid, which is extended from the lamella plane, is provided. The laminar boundary layer is broken down on the one hand by the guide bar, and on the other hand the air is guided in such a way that the formation of a flow shadow behind the pipes is reduced. Due to the required width of the guide bars in the direction of flow of the gaseous fluid, measured at least three quarters of the outside diameter of the connection points, the lamella stability is reduced, so that with a given lamella stability, the material thickness must be increased, since the use of harder lamella material due to the maximum achievable height of the sleeve-shaped connection points up to 2.4 mm is not possible. Even with a material thickness increased from 0.12 to 0.15 mm, there are certain stability problems in handling during the production process in a lamella package in which the tubes have not yet been inserted into the sleeve-shaped connection points, so that relatively long production times are accepted have to. Also, the problem of contamination and the seizing of water when the fin falls below the dew point, analogous to the heat exchanger fin according to DE-A 3 131 737, has not yet been completely solved.

In the generic known slat according to DE-A 2 530 064, which is intended for use in motor vehicle radiators, attempts have already been made to further improve the heat transfer coefficient of a corrugated slat in that flaps are formed from the slat and are formed by punched tear holes and serve to form spacers of successive individual fins in the fin package of the tube fin heat exchanger, to be used as local air-guiding profiles by angular position for the flow of the first fluid. The grid-like arrangement of these inclined air baffles in the gaps of mutually offset heat exchanger tubes or the connecting sleeves of the fins that receive them is intended to improve the action of the first fluid in the slipstream of the heat exchanger tubes in the slipstream. The corrugation of the lamella is so short-wave that a tear hole with two flaps running parallel to each other occupies two flanks of the corrugation, bridging each of a wave crest. In the area of the tear hole there is no corrugation and its desired effect. This design with the tear holes and the sharply curved lobes, which take up the entire distance between adjacent lamellae, is predestined to hold on to condensation water that forms when the temperature falls below the dew point and to trap dirt. Furthermore, the lobes that occupy the entire lamella spacing are relatively large-area turbulence generators with their own slipstream effect for the gaseous first fluid, with the associated reduction in the heat transfer between the lamella and the gaseous first fluid. The inclination of the flaps in only one possible direction of inclination appears to be unmotivated, since the flow of the first fluid has no preferred lateral direction. A multiplication of the flaps serving as spacers would increasingly block the passage cross section of the plate pack for the first fluid and thus also counteract this desired increase in efficiency. The improvement achieved in the known measure with regard to the heat transfer coefficient should therefore only be minimal at all and only be significant if separate spacers of the lamellae are used in the lamella package from the retaining sleeves of the lamellae. In the context of the invention, instead, such separate spacers are preferably completely dispensed with, since it is known per se from previous lamella constructions of the applicant (cf. DE-C 3 336 985) to provide the connecting sleeves of the lamella to the heat exchanger tubes with collars which engage in corresponding grooves on the back of the next lamella, thus allowing the connection sleeves to be used as spacers in the lamella package. However, the invention is not limited to this case, but also allows the use of separate spacers as an option, although not a preferred option.

Substantially closed bulges from flat fins are already known in tubular fin heat exchangers, which are disclosed for materials other than Al or Al alloys.

This is shown, for example, by DE-C 496 733 from 1930, in which the fins are made of sheet metal and are soldered to the tubes. It is obviously intended for iron sheet or stainless steel sheet, since the construction of the sheets with throttling points between the tubes and the air guide for guiding the air flowing through the plate pack into the air shadow areas behind the tubes is based on material which otherwise lacks the thermal conductivity in the slipstream areas mentioned leads to excessive temperature drop and thus a reduction in efficiency. This would not be the case with copper fins. At that time there were not thin aluminum slats, which also caused slipstream problems. In addition, the expressly desired throttling between the pipes leads to high pressure losses.

Especially for motor vehicle radiators, solderable fins, primarily made of copper, are already known from US-A 1 575 864, which dates back to 1926, and in which protruding bulges protrude from one flat fins on one or both sides, in particular conically, which are closed in the finned plane are. Such bulging of flat slats has been considered again and again since the 1930s, also by the applicant, with the most varied of variants, but was also rejected just as often because the achievable surface enlargement and the turbulence increase for the required power density in comparison with the respective time otherwise known lamella constructions of the type discussed above are not sufficient.

From US-A 2 046 791 a lamella for the common ribbing of a tubular lamella heat exchanger in motor vehicles is known, which in addition to features of the preamble of claim 1 also has the features that the air guide profiles are closed bulges with a smaller height than the lamella spacing in the tubular lamella heat exchanger and that the bulges are each arranged on a flank of the corrugation. In this known lamella, good mechanical stability and easy manufacture can only be achieved while maintaining, not increasing, the thermal efficiency, which is already perceived as good, by leading the air flow along the apex lines of the wave crests or wave troughs, i.e. parallel to the corrugation is so that the corrugation does not increase the air resistance and thus the heat transfer.

The invention has for its object to provide a fin for the common ribbing of several heat exchanger tubes of a tube fin heat exchanger in motor vehicles, which has the excellent heat transfer ratio of slotted fins, but disadvantages of the same, such as great water retention when the dew point is undershot, relatively low stability of the fin and high maintenance requirements Manufacturing tools, largely avoided.

This object is achieved in a generic heat exchanger fin in accordance with the characterizing features of claim 1.

The outstanding feature of the lamella according to the invention is that a further increase in the external heat transfer coefficient a _{a is} achieved by the further local profiling impressed on the flanks of the basic corrugation of the lamella in the form of bulges smaller than the lamella spacing, so that even the previous one only with slotted lamellas or with perforated lamellas (DE-C 3 336 985), which were previously considered to be the optimal heat transfer coefficients, can still be exceeded significantly, for example with approximately 8 to 20%. There is only a slight increase in the air-side (first fluid) pressure loss.

With an unprofiled corrugation of the slat, even with an optimal straight zigzag, on the other hand, only performance increases up to a corrugation angle e of 15 to 20 _° can be achieved. The heat transfer coefficient achieved is, however, significantly below the possible heat transfer coefficient of slotted fins (DE-C 3 336 985).

At a larger angle e, only a slight increase in performance can be expected, but a strong increase in the air-side pressure drop (FIG. 7), since the boundary layer through which the heat has to be transported by means of molecular conduction processes is smaller as a result of stationary eddies Intensity that develops in each wave trough increases. In the lamella according to the invention, it is sufficient to select the corrugation angle ₈ only up to its thermodynamically sensible maximum value of 15 to 20 _° and to achieve a further controlled turbulence increase and surface enlargement through the bulges, behind which no dead spaces can form. Thus, the turbulence-enhancing and boundary layer-reducing effect of the features fully benefits the lamella surface located downstream, without this being compensated for again by local boundary layer enlargements. The increase in surface area obtained through the characteristics also increases performance. The bulges can be designed and distributed in such a way that there is practically the same heat flux density in concentric circles around the connection sleeves, that is to say is distributed invariably with respect to the connection sleeves. Of particular note here is the splitting of the flow of the first fluid into flow threads and their conduction into the areas of current dead spaces.

The advantages gained from the above-mentioned specific combination of corrugation of the slats and their provision with the bulges is all the more surprising since this creates a new way of creating performance in comparison with other types of designs of highly efficient slats, the creation of which by corrugation or bulge alone has so far always been unsuccessful . Compared to the most effective lamella of the applicant of a completely different type, namely according to DE-A 3 336 985, an increase in the external heat transfer coefficient of between approximately 8 and 20% and thus the heat output transferred under the same operating condition, same material and the same can be achieved Achieve dimensions by approx. 5 to 10%, ie apart from the different slat types, the conditions are completely the same.

It is preferred to provide completely closed bulges, while the desired effects, in particular with regard to increasing the heat transfer, are already achieved if the bulges are predominantly closed.

Due to the low height of the closed bulges of 15-80%, which is sufficient for the increase in performance, the condensate drainage is only insignificantly disturbed compared to a lamella without bulges. Furthermore, if the external heat transfer coefficient is at least the same, there is no need for exhibitor slits or openings in the lamella, which act comparatively like a sponge in relation to condensed water.

Since the lamella according to the invention thus does not require exhibitor slots and openings, the cutting punches required for this purpose, which are required, for example, to produce the openings in the lamella according to DE-C 3 336 985, can be dispensed with. Since one punch is required per opening and up to 36 rows of tubes are punched at the same time due to the industrial large-scale production in the lamellar tool, cutting punches can be saved compared to a lamella, for example according to DE-C 3 336 985 144, whereby the manufacturing costs of the lamellar tool can be significantly reduced .

Furthermore, in contrast to cutting punches, the punches required for the lamella according to the invention for producing the corrugation or the bulges are maintenance-free.

An embodiment of the heat exchanger lamella according to the invention is therefore particularly recommended when it is used in the motor vehicle as an evaporator or air cooler, in which condensation water forms on the lamella due to the temperature falling below the dew point. Due to the practical elimination of all exhibitors, slits or openings in the lamella, the condensed water can flow off undisturbed, so that the water retention capacity is always lower. Due to the smaller amount of condensed water held between the heat exchanger fins by adhesion in the case of the invention On the one hand, the heat exchanger lamella further improves the external heat transfer coefficient, since the heat resistance is reduced by the condensed water, and on the other hand, the lamella surface dries faster, which reduces the activity of odor-producing bacteria.

Another advantage of the lower water retention capacity of the heat exchanger lamella according to the invention is the better suitability for reheat operation (in vehicle air conditioning systems), since the amount of water evaporated and thus the fogging of the windshield after the compressor has been switched off is less.

Due to the strong deformation of the lamella practically without openings, the stability is greatly increased, in particular in comparison to slotted lamellae, so that the lamella thickness can be significantly reduced with the same strength.

In connection with the above-mentioned increase in performance of the heat exchanger fins according to the invention compared to other high-performance fins (e.g. DE-C 3 336 985), the fin thickness can be significantly reduced without reducing the performance of the above-mentioned high-performance fins according to DE-C 3 336 985. Since the heat exchanger fins are manufactured in large numbers for the automotive industry, the reduction in material costs as well as the weight reduction and the associated improvement in driving properties and the reduction in gasoline consumption are of decisive advantage. Another advantage in the large-scale production of the heat exchanger lamella is the elimination of the cutting punches, which are required to produce the openings in the lamella according to DE-C 3 336 985 and require a high level of maintenance, while the tool inserts for producing the profile of the lamella according to the invention are almost maintenance-free are.

It is understood that the average specialist will select the size, shape, number and distribution of the bulges in a targeted manner. So he will choose not too little bulges, but also not too many small bulges, since otherwise the air of the first fluid cannot follow the lamella. Bulges that are too large, on the other hand, would already create the danger of creating your own current dead spaces. They can also be used less for optimal division into current threads. The wavelength of the corrugation is also related to the size of the bulges, since each bulge is only assigned to one flank of the corrugation and is therefore set back in its respective foot area with respect to the next wave crest. The expressions preferably extend only over the part of the flank length of the corrugation which is in turn optimally selected in accordance with claims 4 and 16, respectively.

The arrangement of the knob-like bulges depending on the pipe distribution is also important for increasing the heat transfer coefficient. While within the scope of claim 8 e.g. with small spacings of the connecting sleeves transversely to the air direction, two bulges arranged in alignment in the air direction may be sufficient to achieve a predetermined heat transfer coefficient with low pressure loss (cf. also FIG. 1), the arrangement of the bulge can be offset to the air direction within the scope of claim 9 with higher performance requirements be (see also Fig. 2).

A further increase in the heat transfer coefficient is achieved in the two cases mentioned, the arrangement of more than two bulges, the center distance of the bulges depending on the lamella spacing b should be reduced so that just no thickening of the boundary layer occurs (see claim 14). It is preferred according to claim 6 that each flank of the corrugation is provided with bulges.

The optimal geometry of the zigzag curl is defined in claim 4. In the case of large fin spacings, a small number of wave crests between two adjacent connection points of a row of pipes, preferably only one wave crest, is expedient; 1.5 or 2 wave crests should preferably be selected in the range of the desired small lamella spacings (cf. also claim 16). Given the geometry of the connection points and the maximum corrugation angle e of 15 to 20 _°, the number of corrugation peaks results in an effective corrugation height, measured at right angles to the base surface or main plane of the slat. The height of the bulges measured as a projection perpendicular to the fin flank is preferably 30 to 50% of the fin spacing in the tubular fin heat exchanger in the sense of claim 19.

Claims 17 and 18 relate to the already mentioned lamellar spacing via the connecting sleeves and underline the necessity that the entire lamellar surface should have a profile that is as uniform as possible in order to achieve a maximum heat transfer coefficient with a relatively low air-side pressure loss. According to claim 18, surfaces with an excessively large angle with respect to the base area of the fins are to be avoided, since in the fin package of the tube fin heat exchanger the fin spacing is reduced by the factor cos e at these points and the water retention capacity increases as the fin spacing decreases.

In claims 11 to 13 preferred forms of bulges are given. Although currently tapered shapes are considered to be particularly useful, other shapes can also be permitted for the purpose of increasing turbulence and increasing the surface area according to claims 12 and 13, which shapes can be embossed without tearing the lamella. Axially symmetrical bulges are preferred; Elongated bulges, for example, are also possible, e.g. with oval cross section.

Claim 10 deals with a special arrangement of the bulges which, in order to obtain a constant heat flow density in the lamella, are arranged in such a way that in the region of low flow velocities, ie large ones Pipe distances, more bulges between two adjacent connection sleeves and less or in the area of high flow velocities, ie small pipe distances between two adjacent connection sleeves, or in the borderline case there are no features.

The effect that can be achieved by such measures, that there is approximately constant heat flow density in the lamella on concentric circles around the connecting sleeves, in conjunction with the higher dimensional stability of the lamella also permits a reduction in the lamella thickness without a deterioration in the efficiency of the ribs, since the entire lamella cross-section available for heat conduction "Concentrated heat" is evenly distributed on each concentric circle around the connection point and the conductivity of the entire lamella is thus used uniformly. Here is the great advantage of the heat exchanger fins according to the invention compared to slotted fins or fins with exhibitors, openings and guide bars, which until now have only been used in the area of high-performance fins, that no areas with extremely high heat transfer coefficients occur, but the heat flow density evenly while reducing the distance to Pipe increases.

Preferably, to make the lamella easier to manufacture, all of the bulges and all connection points protrude from the same side of the lamella.

In the bridge-shaped strips described in DE-A 3 131 737, which are formed by cutting and lifting the material, on the other hand a locally extremely high heat transfer coefficient is achieved by the roof shape of the strips on the one hand and by the necessary reconstruction of the laminar boundary layer on the other. The resulting extremely high heat flow density when the heat flow lines enter the strips issued from the lamella plane, however, requires a locally considerably higher temperature gradient since, for manufacturing reasons, the lamella material cannot be locally varied according to the heat flow density. The increased temperature gradient in the direction of the heat flow lines from the connection point to the center of the bridge-like strip then leads to a reduced temperature difference between the bridge-like strip and the gaseous first fluid flowing on the outside of the heat exchanger lamella, as a result of which the amount of heat transferred is reduced and the locally very high heat transfer numbers do not can be converted accordingly into transferred heat output.

The invention is explained in more detail below with reference to schematic drawings of exemplary embodiments.

• Fig. 1 eine Draufsicht auf einen Abschnitt einer ersten Ausführungsform einer erfindungsgemäßen Lamelle;
• Fig. 2 eine Draufsicht auf eine in bezug auf die Ausbuchtungsanordnung variierte zweite Ausführungsform einer erfindungsgemäßen Lamelle;
• Fig. 3 und 4 zur weiteren Optimierung eine dritte und eine vierte Ausführungsform einer erfindungsgemäßen Lamelle in Draufsicht;
• Fig. 5 einen Schnitt nach der Linie I-I in Fig. 4 mit Darstellung zweier benachbarter Lamellen im Rohrlamellenwärmetauscher;
• Fig. 6 zum Vergleich einen der Fig. 5 entsprechenden Querschnitt durch die bekannte Wärmetauscherlamelle nach der DE-A 2 530 064;
• Fig. 7 ein Diagramm zur Erläuterung des Einflusses des Wellungswinkels e auf die Wärmeübergangszahl und den Druckverlust von ausschließlich zickzackgewellten Lamellen nach dem Stand der Technik;
• Fig. 8 einen Ausschnitt eines Lamellenpaketes mit Schnittführung durch die beiden Achsen zweier benachbarter Wärmetauscherrohre des Rohrlamellenwärmetauschers; sowie
• Fig. 9 schematische Seiten- und Stirnansichten eines solchen Rohrlamellenwärmetauschers.

Show it:

• Figure 1 is a plan view of a portion of a first embodiment of a lamella according to the invention.
• 2 shows a plan view of a second embodiment of a lamella according to the invention which is varied with respect to the bulge arrangement;
• 3 and 4 for further optimization, a third and a fourth embodiment of a lamella according to the invention in plan view;
• 5 shows a section along the line II in FIG. 4, showing two adjacent fins in the tube finned heat exchanger;
• 6 shows a cross section corresponding to FIG. 5 through the known heat exchanger lamella according to DE-A 2 530 064;
• 7 shows a diagram for explaining the influence of the corrugation angle e on the heat transfer coefficient and the pressure loss of exclusively zigzag-corrugated fins according to the prior art;
• 8 shows a section of a plate pack with a cut through the two axes of two adjacent heat exchanger tubes of the tube plate heat exchanger; such as
• Fig. 9 schematic side and end views of such a tube fin heat exchanger.

1 to 4, fins 1 of a tube fin heat exchanger are shown in various embodiments, in which by shaping finned sheet with the preferred thickness of 0.07 to 0.5 mm, preferably 0.07 to 0.15 mm, from Al or an Al alloy thereof, the stamping, drawing or embossing processes produce the surface profiling described below.

Several rows of connecting sleeves 4 for receiving heat exchanger tubes 14 extend in each heat exchanger fin 1 (cf. FIGS. 8 and 9). The connecting sleeves 4 are each designed to receive a heat exchanger tube 14 carrying the second fluid as a cylindrical, elliptical or differently shaped sleeve so that in the direction of the first gaseous fluid which acts on the heat exchanger lamella 1 itself, an outer diameter which is defined except for slight deviations is produced. The connecting sleeves 4 are bent on their outer free edge, in the manner of an outer ring flange, to form a collar 13, in order to thereby fix the mutual spacing of the fins 1 in a fin package of a tube fin heat exchanger 22 (see FIGS. 8 and 9) . The connecting sleeves 4 in turn protrude from an annular receiving trough 3 formed in the lamella 1, in which the corresponding collar 13 of a connecting sleeve 4 of an adjacent lamella 1 engages on the side facing away from the connecting sleeve. In applications in motor vehicle heat exchangers, the fins 1 are acted upon by ambient air as a gaseous first fluid which, via the fin 1 in the tube fin heat exchanger 22, enters into heat exchange with the second fluid carried in the heat exchanger tubes. In general, the flow direction 2 of the first fluid runs transversely to the flow direction of the second fluid following the axial direction of the heat exchanger tubes 14 or the inlet sleeve 4. The flow direction 2 of the first fluid is indicated by directional arrows in the plan views of the heat exchanger lamella 1 shown.

The connection sleeves 4 are arranged in rows transverse to the flow direction 2 of the first fluid. Embodiments are possible in which successive rows of connecting sleeves 4 are set to a gap (FIGS. 1 to 4), as well as those in which adjacent connecting sleeves 4 successive rows are aligned in flow direction 2 (not shown in the drawing). Both embodiments of the slat 1 are possible. The connection sleeves 4 are preferably designed identically. In each row, adjacent connection sleeves have 4 equal distances. The distances are generally the same in different rows. Likewise, the distance between two successive rows in the flow direction is also the same. In the embodiment according to FIG. 1, two conical bulges 6 are arranged on a zigzag-shaped corrugation between two adjacent connection sleeves 4 of the same row in such a way that the bulges 6 are symmetrical on the one hand between the two adjacent connection sleeves 4 of a row and on the other hand approximately in the middle between wave crest 5 and trough 11 are arranged.

• Statt von Wellenberg 5 und Wellental 11 kann man verallgemeinernd auch von Wellungsberg bzw. -tal sprechen. Zwischen einem Wellenberg 5 und einem Wellental 11 liegt jeweils eine Flanke 20 der Wellung. Die Form der Ausbuchtungen 6 kann man auch als Noppen bezeichnen, die je nach Verformungsfähigkeit und Werkzeugaufwand in einem gewissen Rahmen frei gewählt werden. So können außer den dargestellten kegelförmigen Ausbuchtungen 6 auch prismenförmige, zylinderförmige, kugelabschnittförmige Ausbuchtungen 6 oder solche in Form einer drehparabolischen Ausbuchtung bzw. eines Pyramiden- oder Kegelstumpfes oder anderer erhabener Ausprägungen verwendet werden. Die Ausbuchtungen 6 sind alle in derselben Richtung aus der Ebene der Lamelle 1 herausgedrückt. Zusätzlich zu der zickzackförmigen Grundwellung kann noch eine Randwellung 12 an der Ein- bzw. Austrittskante des ersten gasförmigen Fluids beim Trennen der Wärmetauscherlamellen eingeprägt werden, die eine zusätzliche Versteifung der Lamellenkante ergibt und bei Taupunktunterschreitung ein Austreten von Spritzwasser aus der Lamelle 1 reduziert.

The following generally applies:

• Instead of Wellenberg 5 and Wellental 11, one can generally speak of Wellungsberg or -tal. A flank 20 of the corrugation lies between a wave crest 5 and a wave trough 11. The shape of the bulges 6 can also be referred to as knobs, which are freely selected within a certain range, depending on the deformability and tool complexity. In addition to the conical bulges 6 shown, prism-shaped, cylindrical, spherical segment-shaped bulges 6 or those in the form of a parabolic bulge or a truncated pyramid or truncated cone or other raised features can also be used. The bulges 6 are all pushed out of the plane of the lamella 1 in the same direction. In addition to the zigzag-shaped basic corrugation, an edge corrugation 12 can also be embossed on the inlet or outlet edge of the first gaseous fluid when separating the heat exchanger fins, which results in additional stiffening of the fin edge and reduces spray water leakage from the lamella 1 if the temperature falls below the dew point.

All bulges 6 are closed and each have a smaller height than the fin spacing b (see FIG. 5) in the tubular fin heat exchanger 22 (see FIG. 8.9). The bulges 6 are at a distance from both ends of the same flank 20 of the corrugation, so that an expression 6 is only formed on a single flank 20 of the corrugation. In the embodiments of FIGS. 1 to 4, each flank 20 of the corrugation has bulges. gen 6, which each occupy between 50 and 80% of the flank length measured in flow direction 2.

FIG. 2 shows an advantageous modification of the aligned arrangement of the bulges 6 according to FIG. 1, so that they are arranged one behind the other in relation to the flow direction 2 of the first fluid according to FIG. 2. The offset arrangement of the bulges 6 results in a higher pressure loss, but a further increase in the heat transfer coefficient can also be achieved in parallel.

1 and 2, only one bulge 6 per flank 20 is arranged between adjacent connecting sleeves 4, in FIG. 1 in the center, in FIG. 2 on flanks 20 adjoining in flow direction 2, laterally offset symmetrically to the center of the respective flank 20 with mirror symmetry to the imaginary center of the adjacent sleeves in the row.

A further optimization of the basic idea of FIG. 2 is shown in FIG. 3, in which several bulges 6, in the special case of FIGS. 3 and 5, are arranged offset to the flow direction 2 between two connecting sleeves 4 of the same row of tubes. In this case, three and two bulges 6 are alternately arranged on the flat lamella surface between the crest 5 and the trough 11 between two adjacent connecting sleeves 4 of a row of heat exchanger tubes 14 or connecting sleeves 4 in the flow direction 2 of the first fluid. The distribution of the bulges 6 takes place symmetrically to the imaginary center line between adjacent connection sleeves 4, as in the case of FIG. 4 discussed below, with an equidistant distribution of the bulges 6 of a group of bulges 6 lying between two adjacent connection sleeves 4.

Due to the more uniform distribution of the bulges 6, which are also arranged more densely in the region of the low flow velocities between the adjacent connecting sleeves 4 of a row of pipes, a heat flow density of the same size is achieved on concentric circles around the connecting sleeves 4.

With larger spacings of the connecting sleeves 4, a further increase in the number of bulges 6 with the same maximum bulge diameter can have an advantageous effect, while with very small lamella spacings an increase in the number of bulges 6 with a simultaneous reduction in the maximum diameter d of the bulges 6 is recommended.

The height f of the bulges 6 is to be designed in all the embodiments of FIGS. 1 to 4 such that, depending on the permissible pressure loss, it is preferably 30 to 50% of the existing lamella spacing b. The distance a between the centers of adjacent bulges 6 is expediently 1 to 3 times, preferably 1.3 to 2 times the diameter of the base area of the individual bulges 6. A further step in the direction of a concentric circle around the connecting sleeve 4 uniform heat flow density or homogeneous external heat transfer coefficient distributed over the entire fin surface is shown in FIG. 4, in which the number of flanks 20 between two adjacent connecting sleeves 4 of a pipe row he was increased from FIG. 3 from 2 to 3.

As a result, eight bulges 6 can be positioned between two adjacent connection sleeves 4 of a row of heat exchanger tubes 14 or connection sleeves 4 in an arrangement offset from the air direction 2, specifically in the flow direction 2 in the sequence 3 - 2 - 3. A further increase in the number of wave crests 5 is conceivable for smaller lamella spacings b. The limit to the increase in the number of waves and bulges is given by the air flow, which reacts in the event of a too fine corrugation and the extremely high number of bulges associated with the formation of a greater thickness of the boundary layer 9 (see FIG. 6), since the flow the gaseous first fluid can no longer follow a very fine corrugation or bulge. The tool cost also increases with an increasing number of bulges, since the bulges 6 are stamped due to the required interchangeability of the tool profiles when worn with stamps which are embedded in a corrugated base plate, so that the tool costs increase as the number of stamps increases.

The length of the individual flanks 20 of the corrugation is expediently at least twice and at most five times the fin spacing b in the tubular fin heat exchanger.

In Fig. 5 a lamella 1 is shown in section. This figure shows the shape of all bulges 6 in one direction and the preferred relation of the heights f of the bulges 6 in relation to the lamella spacing b. There is also the effect of locally very large corrugation angles a, which lead to a local reduction of the lamella spacing from dimension b to g and thus to increased adhesive forces between the lamella 1 and condensation water droplets.

In FIG. 6, on a lamella 1 (without features 6) designed according to the state of the art according to DE-A 2 530 645 with exclusive corrugation by drawn flow lines, the increase in the boundary layer 9 that occurs when the corrugation angle 8 is too large is illustrated. Since the air cannot nearly follow the lamella 1, stationary vortices 10 of low intensity are formed in the troughs 11, which have only a slight boundary-layer-reducing effect and also adjust in temperature to the lamella 1, since they are essentially stationary are and are not transported in the same way as the turbulence bales in the main flow direction 2.

The resulting powers or pressure losses are plotted against the corrugation angle 8 in FIG. 7. It shows that from corrugation angles e of 20 _° no significant increase in performance is achieved and that there is only a steep increase in air-side pressure losses at corrugation angles e of more than 20 _° , since the resistance coefficient increases with an increased corrugation angle e Redirection also increases the flow path and flow rate.

In Fig. 8 it is shown how in a tube fin heat exchanger 22 (see FIG. 9) staggered heat exchanger tubes 14 are firmly connected to the connecting sleeves 4 of the individual fins 1 in a thermally conductive manner. The attachment is carried out by the usual methods in the manufacture of tube fin heat exchangers, e.g. by expanding the heat exchanger tubes 2 and / or brazing. In the plate pack, the connecting sleeves 4 cause the spacing of adjacent plates 1 by a collar 13 at the free end of the respective connecting sleeve 4 engaging in an annular recess 3 on the rear of the foot zone of the next plate 1. This makes it possible that no additional bulges of the lamella have to take over spacer functions.

The two views of FIG. 9 schematically show an entire tube fin heat exchanger 22, the fins of which are designed according to FIGS. 1, 2, 3 or 4 and, according to FIG. 8, are combined to form a fin stack carried by heat exchanger tubes 14. The individual heat exchanger tubes 14 are combined in terms of flow by means of reversing bends 24, optionally also using header boxes (not shown) or header tubes indicated in FIG. 9, in such a way that they partly cross-flow, partly cross-counter- and cross-countercurrent to the first fluid from a common inlet 26 can flow through a common outlet 28 of the second fluid. The direction of flow of the second fluid is indicated by the arrows 30 at the inlet 26 and 31 at the outlet 28. Furthermore, the flow direction 2 of the first fluid can be seen in the end view of the tube fin heat exchanger 22. In a motor vehicle, the tube fin heat exchanger 22 can be mounted by means of the fastening tabs 32.

Claims (20)

1. Fin (1) made of aluminium or an aluminium alloy for the joint ribbing of several heat-exchanger tubes (14) of a finned tubular heat exchanger (22) in motor vehicles, in which ambient air as a first heat-exchange fluid is guided along the surface of the fin (1) and a second heat-exchange fluid is guided into the heat-exchanger tubes (14), with a design of the fin (1) undulated in the direction of flow (2) of the first fluid, with joining sleeves (4) formed on in the fin (1) for connection to the heat-exchanger tubes (14), at least one wave crest (5) extending between two joining sleeves (4) mutually adjacent transversely relative to the direction of flow (2) of the first fluid, and with local air-guide profiles formed on from the undulated surface of the fin (1) in interspaces between the joining sleeves (4), characterized in that the air-guide profiles are bosses (6) in the finned tubular heat exchanger (22), in that the bosses (6) are arranged respectively on one flank (20) of the undulation, in that the particular boss (6) is at a distance from the two ends of the flank (20), and in that the heigt (f) of the bosses (6) is at least 15% and at most 80% of the fin spacing (b) in the finned tubular heat exchanger (22).

2. Fin according to Claim 1, characterized in that the bosses (6) are of continuous form.

3. Fin according to Claim 2, characterized in that the maximum diameter of the boss (6) is between 50% and 80% of the flank length.

4. Fin according to one of Claims 1 to 3, characterized in that the undulation has a periodic rectilinear zig-zag form, the pitch angle (e) of which is in the range of 10_° to 30_°.

5. Fin according to one of Claims 1 to 4, characterized in that all the bosses (6) and all the joining sleeves (4) project on the same side of the fin (1).

6. Fin according to one of Claims 1 to 5, characterized in that each flank (20) is equipped with bosses (6).

7. Fin according to one of claims 1 to 6, characterized in that a group of two or more bosses (6) is arranged on the same flank (20), transversely relative to the direction of flow (2) of the first fluid, between joining sleeves (4) mutually adjacent transversely relative to the direction of flow of the first fluid.

8. Fin according to one of Claims 1 to 7, characterized in that bosses (6) located between two joining sleeves (4) mutually adjacent transversely relative to the direction of flow (2) of the first fluid are aligned with one another individually or in groups in the direction of flow (2) of the first fluid.

9. Fin according to one of Claims 1 to 7, characterized in that bosses (6) located between two joining sleeves (4) mutually adjacent transversely relative to the direction of flow (2) of the first fluid are arranged offset individually or in groups.

10. Fin according to one of Claims 1 to 9, characterized in that more bosses (6) are arranged in regions of wide flow cross-sections for the first fluid between mutually adjoining sleeves (4) and fewer or no bosses (6) are arranged in the region of small flow cross-sections for the first fluid between mutually adjacent joining sleeves (4).

11. Fin according to one of Claims 1 to 10, characterized in that the bosses (6) have a conical form.

12. Fin according to one of Claims 1 to 10, characterized in that the bosses (6) have the form of a dome.

13. Fin according to one of Claims 1 to 10, characterized in that the bosses (6) have a pyramidal, prismatic or cylindrical form.

14. Fin according to one of Claims 1 to 13, characterized in that the centre distance (a) of mutually adjacent bosses (6) is 1 to 3 times, preferably 1.3 times to double the diameter of the base area of the individual bosses (6).

15. Fin according to one of Claims 1 to 4, characterized in that the length of the individual flanks (20) of the undulation is at least twice and at most five times the fin spacing (b) in the finned tubular heat exchanger (22).

16. Fin according to one of Claims 1 to 15, characterized in that two or three flanks (20) of the undulation come up to a respective row of joining sleeves - (6) located next to one another transversely relative to the direction of flow (2) of the first fluid.

17. Fin according to one of Claims 1 to 16, characterized in that the joining sleeves (4) are equipped with collars (13) at their free ends, in that the fin (1) is equipped with a complementary annular receiving recess (3) on its side facing away from the adjoining sleeves (4), and in that the recess width

is smaller than half the flank length.

18. Fin according to Claim 17, characterized in that the gradient from the receiving recess (3) to the wave crest (5) is no more than 20_° steeper than the pitch angle (e) of the undulation.

19. Fin according to one of Claims 1 to 18, characterized in that the heigt (f) of the bosses (6) is 30 to 50% of the fin spacing (b) in the finned tubular heat-exchanger (22).

20. Fin according to one of Claims 1 to 19, characterized by a superposed marginal undulation (12) of the fin in the region of its edges extending transversely relatie to the direction of flow (2) of the first fluid.

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