ES2657624T3 - Parallel flow evaporator with variable channel insertion depth - Google Patents

Abstract

A parallel flow heat exchanger comprising: an inlet manifold (11) having an inlet opening to conduct the flow of a two-phase fluid within said inlet manifold and a plurality of outlet openings to conduct the flow of fluid from said inlet manifold; a plurality of channels (13) aligned in a substantially parallel relationship and fluidly connected to said plurality of outlet openings to conduct fluid flow from said inlet manifold; an outlet manifold (12) fluidly connected to said plurality of channels to receive fluid flow therefrom, wherein said plurality of channels extend within said inlet manifold at different depths and are substantially flat in transverse planes to the longitudinal axis of the inlet manifold, characterized in that the cross-sectional areas of said inlet manifold are locally enlarged in the vicinity of the areas surrounding said flat channels to allow fluid flow around said plurality of channels.

Description

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DESCRIPTION

Parallel flow evaporator with variable channel insertion depth Background of the invention

This invention relates generally to air conditioning and refrigeration systems and, more particularly, to parallel flow evaporators thereof.

A definition of a so-called parallel flow heat exchanger is widely used in the air conditioning and refrigeration industry today and designates a heat exchanger with a plurality of parallel passages, among which the refrigerant is distributed and flies in the orientation generally substantially perpendicular to the flow direction of the refrigerant in the inlet and outlet manifolds. This definition is well adapted within the technical community and will be used throughout the text.

The poor distribution of the refrigerant in the evaporators of the refrigerant system is a well known phenomenon. It causes a significant degradation of the overall performance of the evaporator and the system in a wide range of operating conditions. Poor refrigerant distribution may occur due to differences in flow impedances within the evaporator channels, uneven airflow distribution over inadequate heat transfer surfaces, poor heat exchanger orientation or poor design of the heat exchanger. manifold and distribution system. The poor distribution is particularly pronounced in parallel flow evaporators due to its specific design with respect to the routing of refrigerant to each refrigerant circuit. Attempts to eliminate or reduce the effects of this phenomenon on the performance of parallel flow evaporators have been made with little or no success. The main reasons for such failures have generally been related to the complexity and inefficiency of the proposed technique or the prohibitively high cost of the solution.

In recent years, parallel flow heat exchangers and welded aluminum heat exchangers in particular have received a lot of attention and interest, not only in the automotive sector but also in the heating, ventilation, air conditioning and cooling industry ( HVAC & R). The main reasons for the use of parallel flow technology are related to its superior performance, high degree of compactness and improved corrosion resistance. Parallel flow heat exchangers are now used in condenser and evaporator applications for multiple products and system designs and configurations. Evaporator applications, while promising greater benefits and rewards, are more challenging and problematic. The poor distribution of the refrigerant is one of the main concerns and obstacles for the implementation of this technology in evaporator applications.

As is known, the poor distribution of the refrigerant in parallel flow heat exchangers occurs due to uneven pressure drop within the channels and in the inlet and outlet manifolds, as well as to the poor design of the manifold and distribution system. . In the collectors, the difference in the length of the refrigerant trajectories, the phase separation, the gravity and the turbulence are the main factors responsible for the poor distribution. Within the heat exchanger channels, variations in heat transfer rate, air flow distribution, manufacturing tolerances and gravity are the key factors. In addition, the recent trend of improved heat exchanger performance promoted the miniaturization of its channels (called mini channels and micro channels), which in turn negatively affected the distribution of refrigerant. Since it is extremely difficult to control all these factors, many of the previous attempts to manage the distribution of refrigerant, especially in parallel flow evaporators, have failed.

In refrigerant systems that use parallel flow heat exchangers, the inlet and outlet manifolds or heads (these terms will be used interchangeably throughout the text) generally have a conventional cylindrical shape. When the two-phase flow enters the head, the vapor phase generally separates from the liquid phase. Since both phases flow independently, poor refrigerant distribution tends to occur.

If the two-phase flow enters the inlet manifold at a relatively high speed, the liquid phase (liquid drops) is transported by the moment of flow farther from the inlet of the collector to the remote part of the head. Therefore, the channels closest to the distributor inlet predominantly receive the vapor phase and the channels away from the distributor inlet mainly receive the liquid phase. If, on the other hand, the velocity of the two-phase flow entering the collector is low, there is not enough time to carry the liquid phase along the head. As a result, the liquid phase enters the channels closest to the inlet and the vapor phase passes to the most remote. In addition, the liquid and vapor phases in the inlet manifold can be separated by gravity forces, which produces similar consequences of poor distribution. In any case, the phenomenon of poor distribution emerges rapidly and manifests itself in the evaporator and in the general degradation of system performance.

JP 2001-304775 describes a single phase parallel flow heat exchanger that includes an input manifold in which the channels extend over a variable distance. However, the input manifold

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It has a constant cross-sectional area, of which a large part is occupied by the channels. This introduces an undesirable impedance to the refrigerant flow along the longitudinal axis of the inlet manifold.

Summary of the Invention

Briefly, according to one aspect of the invention, the insertion depth of the individual parallel channels in the inlet manifold is varied to obtain a more uniform distribution of refrigerant to the individual channels by means of the differential pressure drop created by the variable insertion depth. In this way, a two-phase refrigerant mixture is distributed more evenly between the channels. The variable insertion depth of the individual channels is accommodated by appropriately expanding the diameter of the inlet manifold. The enlargement is variable so that the portions immediately surrounding the individual channels are larger and the portions between them are smaller.

In accordance with another aspect of the invention, the insertion depth of the individual channels is progressively less towards the downstream end of the inlet manifold so that the hydraulic resistance to flow is progressively less towards the rear channels.

According to another aspect of the invention, the insertion depth of the individual channels in the output manifold is also varied to also compensate for the variable flow impedance in the output manifold.

Therefore, the present invention provides the apparatus and method of claims 1 and 5.

In the drawings described below, preferred and alternative embodiments are depicted; however, several other modifications and alternative constructions can be made without departing from the scope of the invention.

Brief description of the drawings

Figure 1 is a schematic illustration of a parallel flow heat exchanger according to the prior art.

Figures 2 and 3 are schematic illustrations of an example that is not part of the present invention.

Figures 4A, 4B and 4C are schematic illustrations of other examples, of which the embodiment shown in Figure 4A is part of the present invention.

Figure 5 is a schematic illustration of yet another embodiment of the present invention.

Description of the preferred embodiment

Referring now to Figure 1, a parallel flow heat exchanger is shown that includes an inlet or manifold head 11, an outlet or manifold head 12 and a plurality of parallel channels 13 arranged to fluidly interconnect the manifold 11 inlet to outlet manifold 12. In general, the inlet and outlet manifolds 11 and 12 are cylindrical in shape, and the channels 13 are usually flattened or round tubes (or extrusions). Channels 13 typically have a plurality of internal and external heat transfer enhancement elements, such as fins. For example, external fins, arranged between them for the improvement of the heat exchange process and structural rigidity, are typically welded to the furnace. Channels 13 may have internal heat transfer improvements and structural elements as well.

The usual way of joining the channels 13 parallel to the input manifold 11 and the output manifold 12 is to insert the channels 13 so that they extend into the internal cavities of the input and output manifolds 11 and 12 as shown by the lines of points. The usual practice is to have the same insertion depth for each of the channels 13. Then they are fixed in position by means of strong welding or the like.

In operation, the two-phase refrigerant flows into the inlet opening 14 and into the internal cavity 16 of the inlet head 11. From the internal cavity 16, the refrigerant, in the form of a liquid, a vapor or a mixture of liquid and steam (the most typical scenario) enters the openings 17 of the tube to pass through the channels 13 to the internal cavity 18 of the output head 12. From there, the refrigerant, which is now usually in the form of steam, exits through the outlet opening 19 and then to the compressor (not shown).

As discussed above, it is desirable that the biphasic refrigerant that passes from the inlet head 11 to the individual channels 13 does so uniformly (or, in other words, with equal steam quality) so that the benefit can be obtained Total heat exchange of the individual channels and flood conditions are not created and observed in the compressor suction (this can damage the compressor). However, due to several phenomena as discussed above, there is a non-uniform flow of refrigerant to the individual channels 13 (referred to as poor distribution). In order to address this problem, applicants have introduced design features that will create different pressure drops for the refrigerant flow from the

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Inlet manifold to individual channels to thereby cause a more uniform flow of refrigerant in channels 13. Additionally, the increased rate of refrigerant flow in the inlet manifold promotes more homogeneous conditions through mixing and jet effects.

With reference now to Figure 2, the present invention is illustrated in accordance with one embodiment. Here, instead of the channels 13 also entering the internal cavity 16 of the input manifold 11, the penetration therein is variable, depending on the position along the longitudinal axis A. As shown, the channel 21 plus close to the entrance 14 penetrates farther into the internal cavity 16 and the following (i.e., channels 22 and 23) are placed in this way and installed with respect to the input manifold 11 to have progressively smaller insertion depths as shown .

In operation, the biphasic refrigerant enters the internal cavity 16 through the inlet 14 and, due to the limited distance between the penetrating end 24 of the tube 21 and the opposite wall 28 of the intake manifold 11, there would be a greater hydraulic resistance and, therefore, a restricted flow in channel 21. The next channel 22, with its reduced insertion depth, provides a greater distance between the end 26 and the wall 28. The next channel 23 downstream has its end 27 inserted an even smaller distance in the cavity, and any subsequent channel progressively decreases in its depth of insertion. Therefore, the problem of the earlier tubes receiving a larger portion of the refrigerant is overcome by selectively varying the flow impedance at the entrance to each of the channels. Additionally, the increased flow rate of refrigerant in the inlet manifold 16 can promote more homogeneous conditions through mixing and jet effects.

It should be taken into account that, if it is difficult to control the insertion depth of the individual channels during manufacturing processes due to a sufficiently large number of channels, then the insertion depth can be controlled in sections with each section having the same depth of insertion and with the depth of insertion varying from section to section and decreasing in the downward direction along the input manifold. In such a case, each individual channel shown in Figure 2 would represent a section of said channels for a sufficiently large heat exchanger.

The illustration in Figure 2 is presented in an exaggerated form for demonstration purposes. Therefore, to understand the magnitudes of the insertion depth for a typical design, exemplary measurements will be provided. Considering an inlet manifold 11 having a typical diameter D of 1 ", the insertion depth Li of the first tube 21 would preferably be in the range of 7/8". The next channel 22 would have an insertion depth of (Li-L2) or (7/8 "- 1/16"), and each subsequent tube would have an insertion depth decreasing by L2 1/16 ". It should be understood that the depth Li insertion of the individual channels depends on many parameters, including the size of the heat exchanger, the size and number of the channel, the typical operating range, the refrigerant and the type of oil circulating in the system, etc.

As seen in Figure 3, due to the increase in insertion depths compared to the prior art, the relatively wide channels 21, 22 and 23, which occupy a large part of the cross-sectional area of the input manifold 11, each can introduce an undesired impedance to the refrigerant flow along the longitudinal axis of the inlet manifold 11. In this example, which is not part of the present invention, this can be accommodated by an increase in the diameter D of the inlet manifold 11.

Instead of increasing the diameter D of the inlet manifold 11 along its entire longitudinal axis, in the present invention, as shown in Figure 4A, the cross-sectional area of a head 31 is enlarged only in the immediate proximity of the insertion points of the channels 21, 22 and 23 in the head 31. In this way, the restriction to the flow of refrigerant around the ends of the channels is avoided or limited to promote uniform favorable conditions for the flow of refrigerant in the channels, as desired. Although the shape and shape of the extensions may vary, the wavy shape tends to provide a smoother and less disturbed movement of the refrigerant that passes along the inlet head and would be preferable.

As shown in Figure 4B and 4C, in examples that are not part of the present invention, an input manifold can be made in an oval or rectangular shape as shown by 37 and 38 respectively, without appreciably increasing its sectional area. Global cross. This will prevent the reduction of the refrigerant flow rate and the possible unwanted phase separation.

In addition, as shown in Figure 5, a similar technique can be applied to the outlet manifold 41, the posterior channels having greater insertion depths. Although the outlet manifold (which typically has a single phase refrigerant vapor) has a less pronounced influence on the distribution of refrigerant between the channels, such equilibrium of the flow impedances will further assist in solving the problem of poor distribution.

In addition, it should be noted that both vertical and horizontal channel orientations will take advantage of the teachings of the present invention, although greater benefits will be obtained for the latter

setting. In addition, although the teachings of this invention are particularly advantageous for evaporator applications, refrigerant system condensers can also benefit from them.

While the present invention has been shown and described particularly with reference to preferred embodiments and alternatives as illustrated in the drawings, one skilled in the art will understand that various changes can be made in detail without departing from the scope of the invention as defined. in the claims.

Claims (7)

5 10 fifteen twenty 25 30 35 40 Four. Five fifty 55 1. A parallel flow heat exchanger comprising: an inlet manifold (11) having an inlet opening to drive the flow of a two-phase fluid within said inlet manifold and a plurality of outlet openings to drive the fluid flow from said inlet manifold; a plurality of channels (13) aligned in a substantially parallel relationship and fluidly connected to said plurality of outlet openings to conduct fluid flow from said inlet manifold; an outlet manifold (12) fluidly connected to said plurality of channels to receive fluid flow therefrom, wherein said plurality of channels extend within said input manifold at different depths and are substantially flat in planes transverse to the longitudinal axis of the input manifold, characterized in that the cross-sectional areas of said input manifold are locally enlarged in the vicinity of the areas surrounding said flat channels to allow fluid flow around said plurality of channels. 2. A parallel flow heat exchanger as set forth in claim 1, wherein the depths of extension in said input manifold for said plurality of channels decrease toward the downstream end of the input manifold. 3. A parallel flow heat exchanger as set forth in claim 2, wherein said parallel channels are divided into sections with each section having equal depths of extension and extension depths in said inlet manifold decreasing from section to section toward the downstream end of the inlet manifold. 4. A parallel flow heat exchanger according to any one of the preceding claims, wherein said plurality of channels extends into said outlet manifold at different depths. 5. A method of operating a parallel flow heat exchanger comprising: providing an inlet manifold (11) having an inlet opening and a plurality of outlet openings; providing a plurality of channels (13) aligned in a substantially parallel relationship, fluidly connected to said plurality of outlet openings, and extended within said input manifold at different depths; wherein said plurality of channels is substantially flat in places transverse to the longitudinal axis of the input manifold and providing an output manifold (12) fluidly connected to said plurality of channels; conduct a flow of a two-phase fluid into the inlet manifold through the inlet opening; wherein a pressure drop is varied in each of said plurality of channels extended in said input manifold at varying depths, and wherein the channels extended in said input manifold at variable depths mix and eject the two-phase fluid to facilitate homogeneous conditions, characterized by providing cross-sectional areas of said inlet manifold that are enlarged locally in the vicinity of the areas surrounding said flat channels to allow fluid flow around said plurality of channels. 6. The method of claim 5, wherein the depths of extension in said input manifold for said plurality of channels decrease toward the downstream end of the input manifold. Method according to claim 5, wherein said parallel channels are divided into sections with each section having equal depths of extension and depths of extension in said input manifold decreasing from section to section towards the downstream end of the input manifold.