EP0401752B1 - Refrigerant condensor for a vehicle air conditioner - Google Patents

Description

The invention relates generally to a condenser for a refrigerant of a vehicle air conditioning system with finned heat exchange tubes through which the refrigerant is passed in cross-flow to the ambient air flowing to it.

Specifically, the invention relates to such a condenser with the further features of the preamble of claim 1, as e.g. is known from GB-A-2 023 798, which will be discussed in more detail below.

In the case of a condenser with the general features mentioned, the heat exchange tubes are generally arranged in a plurality of rows of tubes arranged one behind the other in the flow direction of the ambient air, the respective heat exchange tubes being connected in cross-counterflow. The ribbing preferably, but not exclusively, consists of foils made of Al, Cu or alloys of these materials with a thickness of less than 0.15 mm.

Such condensers for vehicle air conditioning systems are commercially available. So far, all heat exchange tubes have been provided with a common ribbing with fins, which may have been provided with exhibitor-like interruptions for the purpose of improving the heat transfer. Such slat-like slits were oriented in such a way that an optimal heat flow from the tube into the slits was achieved. Such blind-shaped slots accordingly ran along the connecting line of pipes of the same row of pipes or along the connecting line of directly adjacent pipes of adjacent rows of pipes. However, the heat flow between adjacent tubes of the same row of tubes or directly adjacent rows of tubes is not reduced. In addition, the pattern of such is increasing the heat transfer efficiency exhibitor-like slots evenly distributed over the entire ribbing.

In the case of these known condensers, an average temperature level is established in the ribs between adjacent rows of tubes through which flow flows in opposite directions, because of their good heat-conducting connection, which has a performance-reducing effect. This reduction in performance is so pronounced that a cross-countercurrent, which can theoretically produce a significantly higher effective temperature difference, practically brings little improvement in performance compared to a simple cross-current. This effect is increased in the case of condensers for a refrigerant in a vehicle air conditioning system in that the tubes of adjacent rows of tubes (calculated in each case in the direction of flow of the ambient air) are very small and the heat flow transmitted via the ribbing between the tubes of adjacent rows of tubes is particularly large. In the present context, only the particularly serious heat losses via heat conduction are considered, while the heat losses via radiation, which are approximately one order of magnitude smaller, should be disregarded.

A known condenser (cf. DE-U-1 685 651) for the refrigerant of a refrigeration machine - i.e. not for the application according to the invention in a vehicle air conditioning system - consists of one assembly or several identical assemblies, depending on the performance requirement, which are then in accordance with the features of Preamble of claim 1 arranged and connected in cross-countercurrent. All assemblies each contain only one row of tubes and are physically and therefore also thermally completely separate from each other.

From JP-A-58 138986 a heat exchanger is also known in which two separate assemblies for cooling, on the one hand, saturated gas and, on the other hand, saturated gas, in the latter case in one row, are insulated from one another in terms of heat.

JP-A-58 108394 already shows a heat exchanger in which two different assemblies have a common tube fin ribbing and are thermally coupled to one another by the fin bridges between at least one row of slots. You already know overlapping parallel slots, which can also be provided with exhibitors (Fig. 11 and 13). It is also already considered to obtain a partial thermal decoupling by deforming the material of the slats. This known heat exchanger is concerned with the special case of a condenser, in which the internal heat exchange fluid is conducted in the gaseous state in pipes of relatively large diameter and in the already liquefied state in pipes of relatively small diameter.

US-A-2 963 277 already shows a heat exchanger in which a plurality of rows of pipes have a common fin ribbing and a certain division into two assemblies is provided by a row of slots between adjacent rows of pipes which reduces the heat flow in the common fin ribbing (FIG. 4 ).

If, as in the last two known cases mentioned, heat-decoupling slits are arranged along at least one straight line between adjacent assemblies separately from the other design of the slat, the slats and thus also the adjacent assemblies along this line (s) are significantly sensitive to buckling. In addition, there is a larger dimensioning.

With complete mechanical and thus also thermally conductive decoupling of adjacent assemblies, there are problems with mechanical strength of the entire condenser and considerably higher production costs, since practically at least two separate condensers have to be produced and connected in terms of flow in a space that is as constant as possible. These problems are greatly exacerbated by condensers for a refrigerant in a vehicle air conditioning system due to the small dimensions which are adapted to only a small amount of space in motor vehicles.

From the aforementioned GB-A-2 023 798 it is already known to decouple individual rows of pipes through openings in a heat-conducting manner through openings in a condenser for a refrigerant of a vehicle air-conditioning system transversely to the direction of flow, in order to increase the efficiency of the entire condenser increase. In addition are from the same Pre-publication also known openings-forming exhibitors from the slat level, which also create turbulence to increase the efficiency. In these two cases, however, there is no effect of the flaps exhibited as an additional separation on the side of the incoming air in order to functionally divide the heat exchanger into partial heat exchangers transversely to the direction of flow of the incoming air.

The invention has for its object to make the advantages of operating in cross-countercurrent usable for a condenser of a refrigerant intended for use in a vehicle air-conditioning system and thereby to promote the mechanical strength of the entire condenser with little effort and to minimize the dimensions of the entire condenser enable.

This object is achieved in a condenser with the features of the preamble of claim 1 by the characterizing features.

In the case of the condenser according to the invention, too, several assemblies, preferably all of them, are physically combined via a common ribbing. The thermal decoupling of neighboring assemblies results from the inclusion of the already existing turbulence-generating slots in the heat-insulating connection zone without requiring additional space for their creation (small dimensions). A greater strength results from the arrangement of the overall heat-insulating slots as waves or polygon (in the simplest form according to claim 11). As a result, the mechanical strength of the entire condenser, which can even be produced in one piece, can be increased, especially in the small dimensions of condensers for vehicle air conditioning systems, but at least by combining several assemblies or several rows of pipes. A largely thermal decoupling can even be carried out by designing the ribbing between the modules. Only by combining the assemblies is it possible to manufacture and handle the small-sized condensers for vehicle air conditioning systems, or at least by summarized parts of the same, practical and possible.

It has thus been shown that it is not necessary to recreate all interruptions within the connection zone between adjacent assemblies, but that the previously mentioned known exhibitor-type interruptions, which were previously only intended to promote heat transfer, are included in the thermally conductive decoupling of the two adjacent assemblies can involve.

The inventive design can be easily achieved because the pipes are offset from one another in the direction of flow of the ambient air and then, as mentioned, the known exhibitor-like recesses for increasing the heat transfer are included in the sequence of the recesses intended for thermal decoupling between adjacent assemblies are.

Compared to the also conceivable possibilities of thermally conductive decoupling of adjacent assemblies on continuous ribbing, e.g. the installation of insulation material, a cross-sectional weakening, a change in resistance due to doping or the like, the solution according to the invention is less complex.

In the solution according to the invention, the material of the ribbing of the heat exchange tubes of adjacent assemblies can also be the same as in the known condensers for motor vehicle air conditioning systems. By suitably arranging interruptions along the connection zone between adjacent assemblies, however, the heat flow there is significantly reduced by heat conduction. It has been shown that even when the ribbing is formed as foils with a thickness of less than 0.15 mm by the interaction of these foils as a tight package, there is still sufficient mechanical strength of the entire condenser with mechanical assembly of the assemblies, in the limit case without any additional Solidification measure can be achieved. In addition, there is the advantage that the heat exchange tubes of different assemblies can be finned in one operation, as with a conventional condenser and so maintain the manufacturing advantages of the known condensers. Dimensions according to claim 4 are preferred, but, for example, according to claim 5, even heat-conductive decouplings that are lower than the values according to claim 4 can still result in a significant increase in the temperature difference between the refrigerant and ambient air.

In the practically preferred development of the invention according to claim 3, the fin area of each row of tubes assumes the temperature of the refrigerant of the respective row of tubes practically immediately and practically without interaction with other rows of tubes. It has been shown that surprisingly unusually high efficiency improvements can be achieved in comparison with conventional comparable condensers. With the same use of material or the same depth and the same air-side pressure loss, efficiency improvements of around 25% can be achieved, which can be used, for example, in a correspondingly smaller depth with the same cooling capacity.

In all condensers according to the invention for vehicle air conditioning systems, deliberately a thermally conductive design of the ribbing of all heat exchange tubes is deliberately avoided and instead a thermally conductive decoupling of at least two assemblies is selected, which are flowed through in opposite directions during operation of the cross-counterflow. It can remain open in detail how the heat exchange pipes are connected in each individual assembly, for example in each assembly in cross flow or even in cross counterflow. It is also possible to combine known interconnection elements of this type in each assembly. In the borderline case, one could even assign an assembly to each row of pipes and flow through each row of pipes in an opposite direction. However, it has been shown that for practical applications it is usually possible to make do with only two thermally decoupled assemblies, even if these assemblies individually or both contain more than one row of pipes. Three or four rows of pipes are preferred, one row of pipes being arranged in one assembly and the other two rows of pipes being arranged in a second assembly in the former case are, while in the second case two rows of pipes are arranged in each of the two assemblies.

In a condenser according to the invention for vehicle air conditioning systems, an average temperature can no longer be established in a common ribbing of adjacent heat exchange tubes from different assemblies, but a more or less pronounced temperature jump takes place between the two assemblies.

The effective temperature difference between the refrigerant on the one hand and the ambient air on the other hand can be increased significantly again when the condenser is designed. In this case, dimensions according to claims 14 and 15 are preferably used for the two modules mentioned. The meaning of these measures will be explained in more detail later on using functional diagrams of the essential parameters (FIGS. 4 to 6b). From DE-AS 1 072 257 it is known per se to change the number of tubes through which the flow passes in parallel along the flow path of the refrigerant such that the pressure gradient is essentially constant over the entire flow path.

According to claims 5 and 6, the material of the ribbing can be removed, in particular punched out, in the interruptions in the connection zone between adjacent assemblies. In this case, narrow slots are preferably used in order to lose as little ribbing material as possible. But you can also use according to claims 7 to 9, the material of the ribbing in the area of the interruptions to exhibitors who also promote the heat transfer between the refrigerant and ambient air.

Claim 10 preferably provides that the interruptions known per se are formed on blinds, while the other interruptions, which are additionally provided for the thermal separation of the rows of pipes, can be designed as simple thermal interruptions without the formation of blinds. In this regard, reference is made in particular to the alternative possibilities in FIG. 3.

• Fig. 1 eine schematische Darstellung der Verschaltung der Wärmetauschrohre eines vierreihigen Verflüssigers mit vier Baugruppen; eine gemeinsame Lamellenverrippung mit wärmeleitmäßiger Entkopplung im Sinne der Erfindung ist dabei ergänzt zu denken;
• Fig. 2 in Draufsicht auf eine gemeinsame Lamelle eine Anordnung von Unterbrechungen in der Verbindungszone zwischen benachbarten Baugruppen unter Einbeziehung von an sich bekannten ausstellerartigen Unterbrechungen für die Erhöhung des Wärmeübergangs;
• Fig. 3 mögliche Bauformen solcher Unterbrechungen, welche im Rahmen der Erfindung zusätzlich zur wärmeleitmäßigen Entkopplung vorgesehen sind, in drei Varianten a), b) und c) als ausstellerartige Unterbrechungen, wie sie insbesondere in der Fig. 2 dargestellt sind, oder in der Variante d) als einfacher Schlitz; Ausbildungen mit schlitzförmigen Unterbrechungen wären bei der Ausführungsform nach Fig. 2 jedoch ebenfalls möglich;
• die Fig. 4 bis 6b Funktionsdiagramme;
  dabei
• Fig. 6b ein Kältemittelzustandsdiagramm, in welchem Kältemittelkreisläufe eingetragen sind, welche den anhand der Fig. 5 und 6a diskutierten verschiedenen Auslegungen des Verflüssigers in bezug auf den kältemittelseitigen Druckverlust entsprechen;
• Fig. 7 in Anlehnung an Fig. 2 eine Draufsicht auf eine Lamelle eines erfindungsgemäßen Verflüssigers;
• Fig. 8 einen Schnitt nach der Linie B-B in Fig. 7; und
• die Fig. 9 und 10 schematisierte Verschaltungen der Kältemittel führenden Rohre von Verflüssigern des Standes der Technik, und zwar nach Fig. 9 im Kreuzstrom und nach Fig. 10 im Kreuzgegenstrom.

The invention is explained in more detail below with the aid of schematic drawings using several exemplary embodiments. Show it:

• Figure 1 is a schematic representation of the connection of the heat exchange tubes of a four-row condenser with four modules. a common fin ribbing with thermally conductive decoupling in the sense of the invention is to be thought of as a supplement;
• 2 is a plan view of a common lamella, showing an arrangement of interruptions in the connection zone between adjacent assemblies, with the inclusion of exhibitor-like interruptions known per se for increasing the heat transfer;
• Fig. 3 possible designs of such interruptions, which are provided in the context of the invention in addition to the thermal decoupling, in three variants a), b) and c) as exhibitor-like interruptions, as shown in particular in FIG. 2, or in the variant d) as a simple slot; Formations with slot-shaped interruptions would also be possible in the embodiment according to FIG. 2;
• 4 to 6b functional diagrams;
  there
• 6b is a refrigerant state diagram in which refrigerant circuits are entered which correspond to the various designs of the condenser discussed with reference to FIGS. 5 and 6a with regard to the refrigerant-side pressure loss;
• 7, based on FIG. 2, a top view of a lamella of a condenser according to the invention;
• Fig. 8 is a section along the line BB in Fig. 7; and
• 9 and 10 are schematic connections of the refrigerant tubes of condensers of the prior art, namely in FIG. 9 in cross flow and in FIG. 10 in cross counterflow.

In the illustrative known liquefier 9 and 10 provided, the direction of flow of the ambient air is illustrated by the arrows A. In both exemplary embodiments, four rows of pipes are arranged transversely to the direction of flow.

9, the refrigerant is introduced through a connection 2 into a collector 4, to which the four rows of ribbed heat exchange tubes 6 are connected on the input side. All heat exchange tubes 6 have a common, uniform ribbing. On the output side, the four rows of heat exchange tubes 6 are connected to a further collector 8, which is provided with an outlet 10 of the refrigerant. It can be seen that the refrigerant flows in parallel in the four rows from the collector 4 to the collector 8 and crosses the incoming ambient air.

In Fig. 10 the same configuration of ribbed heat exchange tubes 6 is connected in cross-counterflow with respect to the incoming ambient air. Here, between the two inlet and outlet collectors 4 and 8, four opposite directions are shown, in which the refrigerant on the one hand crosses the incoming ambient air and, on the other hand, is guided in counterflow to it from the inlet-side collector 4 to the outlet-side collector 8.

In the embodiment shown, each counter-turn connects only two adjacent pipes in a row. It is also known to increase the pressure loss in each flow-through branch between the collectors 4 and 8 to increase the number of pipes per row up to the limit case that only a single coil or counter-turn is arranged between the inlet-side connection 2 and the outlet 10 is.

The common ribbing of all heat exchange tubes by foils, in particular made of aluminum or an aluminum alloy with a thickness of less than 0.15 mm, usually up to about 0.1 mm, is shown at 12.

The known embodiments of FIGS. 9 and 10 relate specifically to round tube heat exchangers.

Exemplary embodiments of condensers according to the invention are now illustrated, which also Round tube heat exchangers can be.

The condenser is divided into at least two assemblies, each of which e.g. can contain two rows of pipes without restricting generality. In particular, one assembly can be arranged on the inlet side of the refrigerant and the other assembly on the outlet side of the refrigerant, with both assemblies e.g. are switched as the opposite direction.

A ribbing common to the assemblies can have foils made of Al, Cu or alloys of these materials with a thickness of less than 0.15 mm up to a minimum of 0.08 mm according to current rolling technology.

In the case of training as a flat tube heat exchanger, fin fins with foils are expediently provided, which then expediently have thicknesses between 0.15 and 0.25 mm.

As in the description of the prior art according to FIGS. 9 and 10, the direction of flow of the refrigerant is indicated by arrows B, while arrows A show the direction of flow of the ambient air.

The input-side connection 2 to an input-side header 4 and the output-side terminal 10 to an output-side header 8 are also used and the heat exchange tubes are designated by 6.

In the embodiment according to FIG. 2, two assemblies 14 and 16 are shown, while the embodiment according to FIG. 1 shows four assemblies 54, 56, 58 and 60.

1 shows, in a special connection, a preferred circuit diagram of the individual assemblies 54 to 60, specifically on a four-row condenser with common fin fins.

On the input side, four refrigeration circuits are connected in parallel in the assemblies 54 and 56, as shown overall in FIG. 9 for the known condenser.

The two assemblies 58 and 60 are formed by only two circuits connected in parallel, so that thereby with a constant internal cross section of the heat exchange tubes 6 in the assemblies 58 and 60 relative to the assemblies 54 and 56 of the Pressure loss is increased significantly.

In a manner not shown, parallel connections of the type of the modules 54 and 56 could also be continued in the entrance area of the condenser, or circuit measures of the type shown for the modules 58 and 60 could already begin in the modules 54 and 56.

In FIG. 1, intermediate collectors are also dispensed with, in that the individual circuits of the input-side modules 54 and 56 are transferred in pairs by so-called tripods 26 into the two further circuits of the modules 58 and 60.

It goes without saying that the circuit measures described can also be implemented analogously with different numbers of circuits in the individual assemblies. However, the numbers and configurations shown here are preferred.

The embodiment according to FIG. 1 assumes that the individual assemblies 54 to 60 are decoupled in terms of heat conduction in the area of the common fin ribs, as will be explained in more detail below with reference to FIGS. 2, 7 and 8, for example.

The four rows of pipes shown are all decoupled in terms of thermal conductivity into the individual assemblies 54, 56, 58 and 60.

In addition, during the transition from the assemblies 54, 56 to 58, 60, the pressure drop on the refrigerant side is increased by interconnecting parallel circuits 62 to one circuit using a tripod 26.

With a condenser connected in this way, the short-circuit heat flow between the heat exchange tubes in the fin is minimal.

In the embodiment of FIG. 1, a common lamellar ribbing with a largely decoupling in terms of thermal conductivity should be added, as is described in detail with reference to the following FIGS. 2 or 7 and 8.

FIG. 2 shows a plan view of a single heat exchange lamella for a four-row arrangement of heat exchange tubes 6, not shown here. One heat exchange tube each 6 of a tube bundle heat exchanger is arranged in the usual way in a receiving opening 28 of the lamella 30, which is part of the ribbing 12 (analogous to FIGS. 9 and 10). The openings can be formed in the usual way, for example with connecting sleeves for connection to the respective heat exchange tube. One can imagine the individual receiving openings 28 as representative of the arrangement of the collective exchange tubes.

The individual lamellae 30 are held in the usual way at a mutual distance by spacers 32 worked out of the lamella, for example lobes of the lamella material that are exposed.

From the arrangement of the receiving openings 28, one can first recognize the assignment to those condensers in which the heat exchange tubes 6 are each half-spaced in the flow direction of the ambient air.

Known exhibitor-like sole strands 34 are initially arranged in the lamella 30 in order to increase the heat transfer, which extend between adjacent receiving openings 28 each along a row of tubes and thus also lie transversely to those connection openings which are adjacent in the row after next. It can be seen in the embodiment according to FIG. 2 that such slots 34 are not able to decouple neighboring pipes from neighboring rows of pipes from one another in a heat-conducting manner.

For the purpose of this thermally conductive decoupling, additional interruptions 36 are provided which, in the embodiment according to FIG. 2, describe a polygon course together with the slots 34 or are arranged at 45 ° to extend the rows of receiving openings 28.

If necessary, the thermally conductive decoupling can be further increased in that the slots 34 and the interruptions 36 are arranged to overlap one another. However, a good effect can also be achieved without this overlap, although the overlap is preferred because of the increase in the thermal conductivity.

The sequence of slots 34 and interruptions 36 describes the direction of extension of a connecting zone 38 between the two assemblies 14 and 16 and the regions 40 and 42 of the lamella 30 respectively assigned to them.

Without restricting the generality, the interruptions 36 can be designed as simple slots 44 in the manner of variant d) of FIG. 3.

The variants a), b) and c) according to FIG. 3, however, represent preferred configurations of the exhibitor-like additional interruptions 36 shown in FIG. 2, which, however, are also known per se for the slots 34.

In variant a), the material exhibitors are webs 46 which are bent out of the lamella 30 on one side and are preferably arranged together in the shape of a blind.

In the case of variants b) and c), on the other hand, the material exhibitors are cut out of the ribbing on both sides via interfaces 48, so that highlighted roof-like parts 50 are formed, which are each only integrally connected to the lamella 30 on the end side. The variant b) describes a flat roof and the variant c) a gable roof, various forms being possible and also common in connection with the interruptions 34. Accordingly, the interruptions 34 can also have all the shapes selected in FIG. 3, variants a) to c). In the borderline case, one could also provide simple slots according to variant d) deviating from the usual at these points, so that both the interruptions 34 and the interruptions 36 then serve only for thermally conductive decoupling.

Analogously, the arrangement can also be transferred to three-row slats or those with a different number of rows.

The interruptions 36 and the slots 34 known per se are each separated from one another along the connecting zone 38 by relatively narrow connecting webs 52, so that the heat flow takes place solely through these narrow connecting webs and thereby the average thermal conductivity along the connecting zone 38 in accordance with the ratio between interruption and connecting web is reduced.

In Fig. 4, the temperature profile is that of the condenser flowing ambient air and the refrigerant led to the ambient air in cross-counterflow with three counter-turns. The refrigerant is guided in the tubes within a module in cross flow to the air and from module to module in opposite directions, ie in counter flow to the air. Within an assembly, the refrigerant can also be conducted in cross-counterflow with one or two counter-turns if the assembly consists of more than one row of pipes. However, due to the small distance between the neighboring pipes of different rows of pipes, the different temperature is averaged by the lamella, so that the temperature difference, which is increased in contrast to the pure cross flow of the pipes, is not effective in the case of cross counterflow.

FIG. 4 therefore shows the solution optimized for the effective temperature difference, in which each row of tubes one to four according to FIG. 2 is each assigned to an assembly 54, 56, 58, 60.

With such a division of e.g. 4-row condenser in likewise four assemblies 54, 56, 58 and 60, the refrigerant temperature decreasing in the flow direction of the refrigerant according to FIG. 4 cannot be compensated for by short-circuit heat flows in the ribbing, but the curve drawn in FIG. 4 is established as the ribbing temperature, which is below the refrigerant temperature curve also shown.

10, provided that the same outlet temperature is to be reached, the ribbing temperature is considerably lower on average, since the heat in the fin from the heat exchange tubes at the higher temperature at the condenser inlet to the Heat exchange pipes of lower temperature flows at the condenser outlet.

The effective temperature difference can be clearly illustrated by the area between the ribbing and the air temperature curve.

4 shows the increase in the effective temperature difference of a condenser connected according to claim 1 a condenser according to the prior art, also connected in cross-countercurrent, shown as a hatched area (A1).

In contrast to the effective temperature difference of a condenser connected according to the prior art, which is represented by the hatched area (A2), more than a doubling of the effective temperature difference is achieved by the condenser according to the invention. Since the temperature curve shown corresponds to an average operating state of a vehicle air conditioning system, at lower air speeds, i.e. a stronger air heating, an even greater increase in effective temperature difference possible by the condenser according to the invention.

5 and 6 optimization criteria for the refrigerant-side pressure loss are shown. The temperature curve in the refrigerant circuit which arises with different refrigerant-side pressure losses is shown in the refrigerant state diagram in FIG. 6b.

The pressure drop on the refrigerant side must be selected in each individual assembly so that the outlet temperature of the liquefied refrigerant t _{KA is} in the range from its minimum t _KA1 to the minimum of the saturation temperature t _KE1 of the refrigerant entering the condenser.

5, 6a and 6b are explained below using examples.

If you choose a design with a very small pressure drop on the refrigerant side, e.g. 0.05 bar, the internal heat transfer coefficient α, which is plotted qualitatively in FIG. 5 over the pressure drop on the refrigerant side, is minimal.

The minimum effective pressure loss Δ _PK on the refrigerant side results in a maximum effective temperature difference, designated Δt _log in FIG. 5, between the refrigerant on the one hand and the ambient air on the other hand, since the saturation temperature does not decrease in the course of the refrigerant's flow path. On the other hand, the heat transfer coefficient (denoted by K in FIG. 5) is small due to the minimum internal heat transfer coefficient.

The product of heat transfer coefficient with the effective temperature difference (designated in Fig. 5 with K · Δt _log ), which is decisive for the condenser performance, therefore does not reach the maximum value at a pressure loss of 0.05 bar on the refrigerant side.

For this reason, the minimum condensing temperature at the inlet (denoted by t _KE in Fig. 6a) is not reached in a given refrigerant circuit of a vehicle air conditioning system, because due to the smaller heat transfer coefficient K under otherwise constant conditions (such as outer surface, ambient temperature etc. ) The saturation temperature of the refrigerant t _KE and the saturation pressure p _KE must be higher than with a design with a higher heat transfer coefficient. Due to the low pressure drop on the refrigerant side, a lowering of the refrigerant outlet temperature (which is denoted by tKA in FIG. 6a) for the interior cooling of the motor vehicle is additionally prevented.

The refrigerant cycle process, which is used in a condenser with small refrigerant-side pressure drops, e.g. of 0.05 bar, is shown in the refrigerant state diagram in FIG. 6b.

6b shows the limit curve for the liquid state and the limit curve for the gaseous state, which meet at the critical point and can also be referred to as "saturation lines".

• Punkt A: Eintritt in den Verdampfer;
• Punkt B: Austritt aus dem Verdampfer bzw. Eintritt in den Verdichter;
• Punkt C: Austritt aus dem Verdichter bzw. Eintritt in den Verflüssiger;
• Punkt D: Austritt aus dem Verflüssiger bzw. Eintritt in das Drosselorgan des Kältemittelkreislaufes.

The state of the refrigerant is primarily described by the refrigerant pressure P and the enthalpy h, which are plotted as the ordinate or abscissa in FIG. 6b. They represent:

• Point A: entry into the evaporator;
• Point B: exit from the evaporator or entry into the compressor;
• Point C: Leaving the compressor or entering the condenser;
• Point D: Leaving the condenser or entering the throttling element of the refrigerant circuit.

The cycle that occurs in condensers with a pressure loss of 0.05 bar on the refrigerant side is shown in FIG. B, C and D, the direction of the refrigerant circuit is indicated by an arrow. Of the three refrigeration cycles shown, an average inlet pressure p _{KE is reached} at point C, while the outlet pressure P _KA and thus also the saturation temperature assigned by the vapor pressure curve is by far the highest at point D. Since the supercooling of the liquid refrigerant to values below the saturation temperature corresponding to the pressure assumes comparable values for all condenser designs, the liquid refrigerant of which can flow freely from the condenser, the refrigerant outlet temperature measured thermometrically at the outlet of the condenser is also comparatively high. Since the enthalpy h increases with the temperature of the liquid refrigerant, the enthalpy of entry of the refrigerant into the evaporator is also highest at point A.

For this reason, a comparatively small enthalpy difference Δh _{o is available} for heat absorption in the evaporator with constant overheating of the refrigerant emerging from the evaporator (point B), so that less heat can be absorbed per kg of refrigerant circulated by the compressor than with the other two with ' or "designated refrigerant cycle processes. This, in turn, leads to a comparatively high evaporation pressure (points A and B) with otherwise constant conditions, with the resultant higher air outlet temperature from the evaporator and finally a comparatively high interior temperature.

If you increase the pressure drop on the _{refrigerant side} to the value of approx. 0.7 bar that is optimal for the condenser and is shown in _FIGS . 5 and 6a with t _KE1 , the effective temperature difference in FIG. 5 drops, on the other hand, the internal heat transfer coefficient _{α 1} decreases and with it the heat transfer coefficient K too. 5, the increase in the heat transfer coefficient is greater than the decrease in the effective temperature difference from 0.05 bar on the refrigerant-side pressure loss to the pressure loss 0.7 bar, so the product of effective temperature difference with the heat transfer coefficient K · Δt _log at which is decisive for the condenser performance refrigerant side Pressure loss _KE1 t in Fig. 5 at its maximum, which as already explained is equivalent to the minimum of the saturation temperature at the inlet of the condenser t _KE according to Fig. 6a. The pressure loss on the _{refrigerant side, which} is 0.65 bar higher at t _KE1, further lowers the saturation temperature at the condenser outlet tKA.

If one looks at the refrigerant condenser described last in the entire refrigeration cycle according to FIG. 6b, one can see the minimum refrigerant _{inlet pressure} P _KE , which is synonymous with the minimally saturated refrigerant _{inlet temperature} t _KE1 in point C ', and the pressure loss Δp _K des represented by the gradient to the left Condenser with the consequence that the outlet pressure p _KA and the refrigerant outlet temperature are lower, whereby the enthalpy difference h _o 'available to the evaporator is greater than that of a condenser with a pressure loss of 0.05 bar on the refrigerant side.

As already explained, this results in a comparatively lower evaporation, air outlet and vehicle interior temperature.

A further reduction in the condenser outlet temperature tKA can be achieved by a further increase in the pressure _drop on the refrigerant side from t _KE1 to t _KE2 .

With this dimensioning, however, the condenser output determined by KΔt _{log is} no longer maximum, since the effective temperature difference decreases more than the heat transfer coefficient increases, so that the saturation temperature at the condenser inlet also increases (see point C '' in Fig. 6b).

However, if the compressor with "steep characteristic curve," almost ie delivery pressure independent volume flow, is used, in accordance with the vapor pressure curve is reduced with the saturation temperature t _KE rising refrigerant inlet pressure is not p _KE the refrigerant mass flow, so that the (from the refrigerant outlet temperature from the condenser point D '' in 6b) resulting maximum enthalpy difference Δh _o ^{,, of} the refrigerant in the evaporator for a further reduction in the evaporation pressure in points A ″ and B ″ and thus to the minimum possible air outlet temperature from the evaporator and the maximum possible interior cooling.

In the case of the condenser mentioned in FIGS. 7 and 8, three assemblies 14, 15 and 16 are provided without any restriction of generality, each of which is assigned to a single row of tubes. Only one fin of the fin package forming the ribs of the corresponding heat exchange tubes is shown. Each lamella has 30 receiving openings 28, into each of which a heat exchange tube is fitted mechanically firmly and in a heat-conducting manner. It can be seen in FIG. 8 that the corresponding receiving openings 28 protrude from the lamellar plane in the form of a sleeve.

From the distribution of the receiving openings 28 it also follows that the heat exchange tubes are regularly offset from one another in the flow direction A of the ambient air.

In the sequence of interruptions 36 provided between the individual assemblies, known interruptions 34 are included, which are each arranged transversely between pairs of heat exchange tubes (or receiving openings 28), which belong to different rows of tubes of separate assemblies 14, 15 and 16.

The slots 34 and interruptions 36 thus form a series of interruptions in the lamella 30 along the respective connection zone 38 between the modules 14 and 15 or 15 and 16, between which connecting webs 52 remain and which are each arranged between pairs of heat exchange tubes or receiving openings 28 are the directly adjacent rows of pipes belonging to the adjacent assemblies, here rows of pipes.

The interruptions 36 are designed here in accordance with the uppermost variant a) of FIG. 3 as elongated slots with a one-sided exhibitor. In contrast, the slots 34 known per se are designed as blinds, the special shape of which is clear from FIG. 8. These are two middle full webs and two outer half webs, which are exhibited parallel to each other and have an angle of attack of preferably 15 to 30 ° to the air.

The slits 34, which are designed as blinds, run in the offset pipe arrangement in each case in the same row of pipes with longitudinal extension between adjacent pipes of the same row of pipes or, in other words, with transverse extension, that is to say separating, between adjacent pipes of pipe pairs lying one behind the other in the flow direction A, each with an intermediate pipe row staggered pipes are separated from each other.

Spacers 64 can also be seen, which are shown at a greater height from the lamella plane on the same side as the sleeves of the receiving openings 28 in order to distance the individual lamellae in the compressed lamella package. Possible shapes and dimensions of such exhibitors are known per se. Figures 7 and 8 show two different preferred possible shapes, which differ in the one or two-sided web display. 8 are expediently tapered so as not to fit into the opposite opening of the next spacer of the adjacent lamella.

The slats 30 are also expediently foils made of Al, Cu or alloys of these materials with a thickness of less than 0.15 mm.

7 or 8, condensers with three or four rows of pipes are preferably formed with the construction, but condensers with only two rows of pipes are also possible in the sense of the preceding description.

The lamella 30 is common to the individual rows of pipes; the cohesion takes place via the connecting webs 52 which remain between the interruptions.

Claims (16)

1. Liquefier for a refrigerant of a vehicle air-conditioning installation with ribbed or finned heat-exchange tubes (6) through which the refrigerant is conducted in crosscurrent to the incident ambient air, wherein the heat-exchange tubes are arranged in a plurality of tube rows, with a common fin system, arranged one behind the other in the incident-flow direction of the ambient air, and the tubes (6) of tube rows situated in succession to one another in the ambient air flow direction are offset relatively to one another, and wherein slots (34) are formed in the fin system which improve the transfer of heat and which extend in the direction in which the tube row extends, between neighbouring tubes (6) of a tube row, characterised in that
   the tube rows form a plurality of subassemblies or units (14, 16) which are arranged one behind the other in the incident-flow direction of the ambient air and which at the refrigerant side are connected as to flow in series in countercurrent to the incident-flow direction and are connected mechanically by their fin system (12), and that between neighbouring units (14, 16) there are arranged interruptions (36) which reduce the heat flow between the -units and which together with the slots (34) form a connection zone (38) of adjacent units which is in the form of a polygonal or wave-train formation and in which the mean thermal conductivity λ_m is below 20% of the thermal conductivity λ of the material of the fin system (12).
2. Liquefier according to claim 1, characterised in that in the connection zone (38) the mean thermal conductivity λ_m is below 10% of the thermal conductivity λ of the material of the fin system (12) of the two neighbouring units (14, 16).
3. Liquefier according to claim 1 or 2, characterised in that each row of heat-exchange tubes forms a unit (14, 15, 16).
4. Liquefier according to one of claims 1 to 3, characterised in that, measured in the direction of the extent of the connection zone (38), the mean length of the connecting webs (52) amounts to less than 50%, preferably less than 20 %, most preferably less than 10%, of the mean length of the interruptions (36).
5. Liquefier according to one of claims 1 to 4, characterised in that interruptions (36) are formed as material gaps (44), preferably punched holes.
6. Liquefier according to claim 5, characterised in that the material gaps are slots (44) extending along the connection zone.
7. Liquefier according to one of claims 1 to 6, characterised in that interruptions (36) are formed as displaced-material elements (46; 50).
8. Liquefier according to claim 7, characterised in that displaced-material elements are strips or webs (46) bent out unilaterally from the fin system (12) and arranged preferably jointly in louvre formation.
9. Liquefier according to claim 7 or 8, characterised in that displaced-material elements (50) are displaced bilaterally of the fin system (12).
10. Liquefier according to one of claims 1 to 9, characterised in that the slots (34) are formed as louvre elements.
11. Liquefier according to one of claims 1 to 10, characterised in that the slots (34) and the interruptions describe a polygonal formation.
12. Liquefier according to one of claims 1 to 11, characterised in that only two units (14, 16) are provided.
13. Liquefier according to one of claims 1 to 12, characterised in that a (first) unit (54, 56) through which the refrigerant first flows is designed with a relatively small cold-side pressure loss, and a (second) unit (58, 60) through which the refrigerant thereafter flows is designed with a relatively high cold-side pressure loss.
14. Liquefier according to claim 13, characterised in that the pressure loss of the first unit (54, 46) is made such that the product of effective temperature difference (Δt_log) between ambient air and refrigerant, on the one hand, and the heat transition coefficient k, on the other hand, is maximal.
15. Liquefier according to claim 13 or 14, characterised in that the pressure loss of the second unit (58, 60) is made such that the exit temperature (t_KA) of the liquefied refrigerant is in the range from the minimum thereof to the minimum of the saturation temperature (t_KE) of the refrigerant entering the liquefier.
16. Liquefier according to one of claims 1 to 15, characterised in that the fin system consists of sheets (30) made of Al, Cu, or alloys of these materials, with a thickness of less than 0.15 mm.

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