U.S. patent application number 10/987960 was filed with the patent office on 2006-05-18 for parallel flow evaporator with variable channel insertion depth.
This patent application is currently assigned to Carrier Corporation. Invention is credited to Robert A. Chopko, Allen C. Kirkwood, Michael F. Taras.
Application Number | 20060101849 10/987960 |
Document ID | / |
Family ID | 36384714 |
Filed Date | 2006-05-18 |
United States Patent
Application |
20060101849 |
Kind Code |
A1 |
Taras; Michael F. ; et
al. |
May 18, 2006 |
Parallel flow evaporator with variable channel insertion depth
Abstract
In a parallel flow heat exchanger having an inlet manifold
connected to a plurality of parallel channels, the degree of
insertion depth of the parallel channels into the inlet manifold is
variable so as to adjust the impedance to the refrigerant flow into
the individual channels. The degree of insertion depth is
progressively reduced toward a downstream end of the manifold for
the individual channels or for the channel sections. The diameter
of the inlet manifold is locally increased or its cross-section
area altered in order to accommodate the flow of refrigerant around
the tube insertions. Similar technique is applied to the outlet
manifold as well to further balance hydraulic resistances.
Inventors: |
Taras; Michael F.;
(Fayetteville, NY) ; Kirkwood; Allen C.;
(Danville, IN) ; Chopko; Robert A.;
(Baldwinsville, NY) |
Correspondence
Address: |
WALL MARJAMA & BILINSKI
101 SOUTH SALINA STREET
SUITE 400
SYRACUSE
NY
13202
US
|
Assignee: |
Carrier Corporation
Syracuse
NY
|
Family ID: |
36384714 |
Appl. No.: |
10/987960 |
Filed: |
November 12, 2004 |
Current U.S.
Class: |
62/515 ; 165/174;
62/525 |
Current CPC
Class: |
F28D 2021/0071 20130101;
F25B 2500/01 20130101; F28F 9/0282 20130101; F25B 39/00 20130101;
F28F 9/02 20130101; F28D 1/05383 20130101 |
Class at
Publication: |
062/515 ;
062/525; 165/174 |
International
Class: |
F28F 9/02 20060101
F28F009/02; F25B 39/02 20060101 F25B039/02 |
Claims
1. A parallel flow heat exchanger comprising: an inlet manifold
having an inlet opening for conducting the flow of fluid into said
inlet manifold and a plurality of outlet openings for conducting
the flow of fluid from said inlet manifold; a plurality of channels
aligned in a substantially parallel relationship and fluidly
connected to said plurality of outlet openings for conducting the
flow of fluid from said inlet manifold; an outlet manifold fluidly
connected to said plurality of channels for receiving the flow of
fluid therefrom; wherein said plurality of channels extend into
said inlet manifold at varying depths.
2. A parallel flow heat exchanger as set forth in claim 1, wherein
the depths of extension into said inlet manifold for said plurality
of channels decrease toward the downstream end of the inlet
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 extension depths and the depths of extension into said
inlet manifold decreasing from section to section toward the
downstream end of the inlet manifold.
4. A parallel flow heat exchanger as set forth in claim 1, wherein
said plurality of channels are substantially flat in planes
transverse to the longitudinal axis of the inlet manifold and
further wherein the cross-section areas of said inlet manifold are
locally enlarged in the vicinities of those areas surrounding said
flat channels to allow for the flow of refrigerant around said
plurality of channels.
5. A parallel flow heat exchanger as set forth in claim 1, wherein
said plurality of channels are substantially flat in planes
transverse to the longitudinal axis of the inlet manifold and
further wherein the cross-section area of said inlet manifold is of
an oval shape.
6. A parallel flow heat exchanger as set forth in claim 1, wherein
said plurality of channels is substantially flat in a direction
transverse to the longitudinal axis of the inlet manifold and
further wherein the cross-section area of said inlet manifold is of
a rectangular shape.
7. A parallel flow heat exchanger as set forth in claim 1 wherein
said plurality of channels extend into said outlet manifold at
varying depths.
8. A parallel flow heat exchanger of the type having an inlet
manifold fluidly interconnected to an outlet manifold by a
plurality of parallel channels for conducting the flow of a fluid
therethrough and adapted for having a second fluid circulated
thereover for purposes of exchange of heat between the two fluids;
wherein said plurality of parallel channels extend into said inlet
manifold at varying depths.
9. A parallel flow heat exchanger as set forth in claim 8, wherein
the depths of extension into said inlet manifold for said plurality
of channels decrease toward the downstream end of the inlet
manifold.
10. A parallel flow heat exchanger as set forth in claim 9, wherein
said parallel channels are divided into sections with each section
having equal extension depths and the depths of extension into said
inlet manifold decreasing from section to section toward the
downstream end of the inlet manifold.
11. A parallel flow heat exchanger as set forth in claim 8, wherein
said plurality of channels are substantially flat in planes
transverse to the longitudinal axis of the inlet manifold and
further wherein the cross-section areas of said inlet manifold are
locally enlarged in the vicinities of those areas surrounding said
flat channels to allow for the flow of refrigerant around said
plurality of channels.
12. A parallel flow heat exchanger as set forth in claim 8, wherein
said plurality of channels is substantially flat in a direction
transverse to the longitudinal axis of the inlet manifold and
further wherein the cross-section area of said inlet manifold is of
an oval shape.
13. A parallel flow heat exchanger as set forth-in claim 8, wherein
said plurality of channels is substantially flat in a direction
transverse to the longitudinal axis of the inlet manifold and
further wherein the cross-section area of said inlet manifold is of
a rectangular shape.
14. A parallel flow heat exchanger as set forth in claim 8 wherein
said plurality of parallel channels extend into said outlet
manifold at varying depths.
Description
BACKGROUND OF THE INVENTION
[0001] This invention relates generally to air conditioning and
refrigeration systems and, more particularly, to parallel flow
evaporators thereof.
[0002] A definition of a so-called parallel flow heat exchanger is
widely used in the air conditioning and refrigeration industry now
and designates a heat exchanger with a plurality of parallel
passages, among which refrigerant is distributed and flown in the
orientation generally substantially perpendicular to the
refrigerant flow direction in the inlet and outlet manifolds. This
definition is well adapted within the technical community and will
be used throughout the text.
[0003] Refrigerant maldistribution in refrigerant system
evaporators is a well-known phenomenon. It causes significant
evaporator and overall system performance degradation over a wide
range of operating conditions. Maldistribution of refrigerant may
occur due to differences in flow impedances within evaporator
channels, non-uniform airflow distribution over external heat
transfer surfaces improper heat exchanger orientation or poor
manifold and distribution system design. Maldistribution is
particularly pronounced in parallel flow evaporators due to their
specific design with respect to refrigerant routing 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 primary reasons for
such failures have generally been related to complexity and
inefficiency of the proposed technique or prohibitively high cost
of the solution.
[0004] In recent years, parallel flow heat exchangers, and brazed
aluminum heat exchangers in particular, have received much
attention and interest, not just in the automotive field but also
in the heating, ventilation, air conditioning and refrigeration
(HVAC&R) industry. The primary reasons for the employment of
the parallel flow technology are related to its superior
performance, high degree of compactness and enhanced resistance to
corrosion. Parallel flow heat exchangers are now utilized in both
condenser and evaporator applications for multiple products and
system designs and configurations. The evaporator applications,
although promising greater benefits and rewards, are more
challenging and problematic. Refrigerant maldistribution is one of
the primary concerns and obstacles for the implementation of this
technology in the evaporator applications.
[0005] As known, refrigerant maldistribution in parallel flow heat
exchangers occurs because of unequal pressure drop inside the
channels and in the inlet and outlet manifolds, as well as poor
manifold and distribution system design. In the manifolds, the
difference in length of refrigerant paths, phase separation,
gravity and turbulence are the primary factors responsible for
maldistribution. Inside the heat exchanger channels, variations in
the heat transfer rate, airflow distribution, manufacturing
tolerances, and gravity are the dominant factors. Furthermore, the
recent trend of the heat exchanger performance enhancement promoted
miniaturization of its channels (so-called minichannels and
microchannels), which in turn negatively impacted refrigerant
distribution. Since it is extremely difficult to control all these
factors, many of the previous attempts to manage refrigerant
distribution, especially in parallel flow evaporators, have
failed.
[0006] In the refrigerant systems utilizing parallel flow heat
exchangers, the inlet and outlet manifolds or headers (these terms
will be used interchangeably throughout the text) usually have a
conventional cylindrical shape. When the two-phase flow enters the
header, the vapor phase is usually separated from the liquid phase.
Since both phases flow independently, refrigerant maldistribution
tends to occur.
[0007] If the two-phase flow enters the inlet manifold at a
relatively high velocity, the liquid phase (droplets of liquid) is
carried by the momentum of the flow further away from the manifold
entrance to the remote portion of the header. Hence, the channels
closest to the manifold entrance receive predominantly the vapor
phase and the channels remote from the manifold entrance receive
mostly the liquid phase. If, on the other hand, the velocity of the
two-phase flow entering the manifold is low, there is not enough
momentum to carry the liquid phase along the header. As a result,
the liquid phase enters the channels closest to the inlet and the
vapor phase proceeds to the most remote ones. Also, the liquid and
vapor phases in the inlet manifold can be separated by the gravity
forces, causing similar maldistribution consequences. In either
case, maldistribution phenomenon quickly surfaces and manifests
itself in evaporator and overall system performance
degradation.
SUMMARY OF THE INVENTION
[0008] Briefly, in accordance with one aspect of the invention, the
insertion depth of the individual parallel channels into the inlet
manifold is varied so as to obtain a more uniform refrigerant
distribution to the individual channels by way of the differential
pressure drop that is created by the variable insertion depth. In
this way, a two-phase refrigerant mixture is more uniformly
distributed among the channels.
[0009] In accordance with another aspect of the invention, the
insertion depth of the individual channels is progressively smaller
toward the downstream end of the inlet manifold such that the
hydraulic resistance to flow is progressively lower toward the
downstream channels.
[0010] In accordance with another aspect of the invention, the
variable insertion depth of the individual channels is accommodated
by appropriately enlarging the diameter of the inlet manifold. The
enlargement can be uniform in a cross-section perpendicular to the
refrigerant flow to result in a cylindrical inlet manifold or it
can be variable such that the portions immediately surrounding the
individual channels are larger and the portions therebetween are
smaller.
[0011] In accordance with yet another aspect of the invention, the
insertion depth of the individual channels into the outlet manifold
is also varied to compensate for variable flow impedance in the
outlet manifold as well.
[0012] In the drawings as hereinafter described, preferred and
alternate embodiments are depicted; however, various other
modifications and alternate constructions can be made thereto
without departing from the true spirit and scope of the
invention.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] FIG. 1 is a schematic illustration of a parallel flow heat
exchanger in accordance with the prior art.
[0014] FIGS. 2 and 3 are schematic illustrations of one embodiment
of the present invention.
[0015] FIGS. 4A, 4B and 4C are schematic illustrations of other
embodiments of the present invention.
[0016] FIG. 5 is a schematic illustration of yet another embodiment
of the present invention.
DESCRIPTION OF THE PREFERRED EMBODIMENT
[0017] Referring now to FIG. 1, a parallel flow heat exchanger is
shown to include an inlet header or manifold 1 1, an outlet header
or manifold 12 and a plurality of parallel disposed channels 13
fluidly interconnecting the inlet manifold 11 to the outlet
manifold 12. Generally, the inlet and outlet manifolds 11 and 12
are cylindrical in shape, and the channels 13 are usually tubes (or
extrusions) of flattened or round shape. Channels 13 normally have
a plurality of internal and external heat transfer enhancement
elements, such as fins. For instance, external fins, disposed
therebetween for the enhancement of the heat exchange process and
structural rigidity, are typically furnace-brazed. Channels 13 may
have internal heat transfer enhancements and structural elements as
well.
[0018] The usual manner of attaching the parallel channels 13 to
the inlet manifold 11 and the outlet manifold 12 is to insert the
channels 13 so that they extend into the internal cavities of the
inlet and outlet manifolds 11 and 12 as shown by the dotted lines.
The usual practice is to have equal insertion depth for each of the
channels 13. They are then fixed in position by way of brazing or
the like.
[0019] In operation, two-phase refrigerant flows into the inlet
opening 14 and into the internal cavity 16 of the inlet header 11.
From the internal cavity 16, the refrigerant, in the form of a
liquid, a vapor or a mixture of liquid and vapor (the most typical
scenario) enters the tube openings 17 to pass through the channels
13 to the internal cavity 18 of the outlet header 12. From there,
the refrigerant, which is now usually in the form of a vapor,
passes out the outlet opening 19 and then to the compressor (not
shown).
[0020] As discussed hereinabove, it is desirable that the two-phase
refrigerant passing from the inlet header 11 to the individual
channels 13 do so in a uniform manner (or in other words, with
equal vapor quality) such that the full heat exchange benefit of
the individual channels can be obtained and flooding conditions are
not created and observed at the compressor suction (this may damage
the compressor). However, because of various phenomena as discussed
hereinabove, a non-uniform flow of refrigerant to the individual
channels 13 (so-called maldistribution) occurs. In order to address
this problem, the applicants have introduced design features that
will create different pressure drop for flow of refrigerant from
the inlet manifold to the individual channels to thereby bring
about a more uniform flow of refrigerant into the channels 13.
Additionally, increased velocity of the refrigerant flow in the
inlet manifold promotes more homogeneous conditions through mixing
and jetting effects.
[0021] Referring now to FIG. 2, the present invention is
illustrated in accordance with one embodiment. Here, instead of the
channels 13 penetrating equally into the internal cavity 16 of the
inlet manifold 11, the penetration thereinto is variable, depending
on the position along the longitudinal axis A. As shown, the
channel 21 closest to the inlet 14 penetrates the furthest into the
internal cavity 16 and those following (i.e. channels 22 and 23)
are so placed and installed with respect to the inlet manifold 11
so as to have progressively smaller insertion depths as shown.
[0022] In operation, the two-phase refrigerant enters the internal
cavity 16 by way of the inlet 14 and, because of the limited
distance between the penetrating end 24 of tube 21 and the opposing
wall 28 of the inlet manifold 11, there would be increased
hydraulic resistance and therefore restricted flow into the 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
downstream channel 23 has its end 27 inserted an even smaller
distance into the cavity, and any subsequent channels are
progressively decreased in their insertion depth. Therefore, the
problem of the more upstream tubes receiving a greater portion of
the refrigerant is overcome by selectively varying the impedance to
the flow at the entrance into each of the channels. Additionally,
increased velocity of the refrigerant flow in the inlet manifold 16
may promote more homogeneous conditions through mixing and jetting
effects.
[0023] It has to be noted that if it becomes difficult to control
the insertion depth of the individual channels during the
manufacturing processes due to a sufficiently large number of
channels, then the insertion depth can be controlled in sections
with each section having equal insertion depth and with the
insertion depth varying from section to section and decreasing in
the downstream direction along the inlet manifold. In such case,
each individual channel shown in FIG. 2 would represent a section
of such channels for a sufficiently large heat exchanger.
[0024] The FIG. 2 illustration is presented in exaggerated form for
demonstrative purposes. Therefore, in order 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 L.sub.1 of
the first tube 21 would preferably be in the range of 7/8''. The
next channel 22 would have an insertion depth of (L.sub.1-L.sub.2)
or (7/8''- 1/16''), and each succeeding tube would have a
diminishing insertion depth by L.sub.2 1/16''. It has to be
understood the insertion depth L.sub.1 of the individual channels
depends on many parameters, including the heat exchanger size,
channel size and number, typical operating range, refrigerant and
oil type circulating through the system, etc.
[0025] As is seen in FIG. 3, because of increased insertion depths
as compared with the prior art, the relatively wide channels 21, 22
and 23, which occupy a large part of the cross-section area of the
inlet manifold I 1, may each introduce undesired impedance to the
refrigerant flow along the longitudinal axis of the inlet manifold
11. This may be accommodated by an increase in the diameter D of
the inlet manifold 11.
[0026] Rather than increasing the diameter D of the inlet manifold
11 along its entire longitudinal axis, an alternative design is
shown in FIG. 4A wherein the cross-section area of a header 31 is
enlarged only in the immediate vicinity of the insertion points of
the channels 21, 22 and 23 into the header 31. In this way, the
restriction to the refrigerant flow around the ends of the channels
is avoided or limited so as to promote favorable uniform conditions
to the refrigerant flow into the channels, as desired. Although the
form and shape of the enlargements may vary, the wavy shape tends
to provide a smoother, less disturbed motion of the refrigerant
passing along the inlet header and would be preferred.
[0027] Alternatively, as shown in FIG. 4B and 4C, an inlet manifold
can be made of an oval or rectangular shape as shown by 37 and 38
respectively, without appreciably increasing its overall
cross-section area. This will prevent refrigerant flow velocity
reduction and potential undesired phase separation.
[0028] Furthermore, as shown in FIG. 5, a similar technique can be
applied to the outlet manifold 41, with the downstream channels
having higher insertion depths. Although the outlet manifold
(typically having a single phase refrigerant vapor) has a less
pronounced influence on the refrigerant distribution among the
channels, such balancing of the flow impedances will further assist
in the maldistribution problem resolution.
[0029] Furthermore, it should be noted that both vertical and
horizontal channel orientations will take advantage from the
teaching of the present invention, although higher benefits will be
obtained for the latter configuration. Also, although the teachings
of this invention are particularly advantageous for the evaporator
applications, refrigerant system condensers may benefit from them
as well.
[0030] While the present invention has been particularly shown and
described with reference to preferred and alternate embodiments as
illustrated in the drawings, it will be understood by one skilled
in the art that various changes in detail may be effected therein
without departing from the true spirit and scope of the invention
as defined by the claims.
* * * * *