U.S. patent application number 10/812474 was filed with the patent office on 2005-10-06 for integral primary and secondary heat exchanger.
Invention is credited to Papapanu, Steven James.
Application Number | 20050217839 10/812474 |
Document ID | / |
Family ID | 35053008 |
Filed Date | 2005-10-06 |
United States Patent
Application |
20050217839 |
Kind Code |
A1 |
Papapanu, Steven James |
October 6, 2005 |
Integral primary and secondary heat exchanger
Abstract
An integral automotive refrigerant condenser and oil cooler with
a cross flow, one pass configuration has condenser and oil cooler
tubes with a common total tube thickness, in spite of the
substantially smaller refrigerant flow passage size. The condenser
tubes have the flow passages arranged inside the tube in a novel
configuration that accommodates the smaller passage size within the
common thickness tube.
Inventors: |
Papapanu, Steven James;
(Lockport, NY) |
Correspondence
Address: |
PATRICK M. GRIFFIN
DELPHI TECHNOLOGIES, INC.
Legal Staff, Mail Code: 480-410-202
P.O. Box 5052
Troy
MI
48007-5052
US
|
Family ID: |
35053008 |
Appl. No.: |
10/812474 |
Filed: |
March 30, 2004 |
Current U.S.
Class: |
165/185 |
Current CPC
Class: |
F28F 1/022 20130101;
F28D 2021/0084 20130101; F28D 2021/0089 20130101; F28D 1/0443
20130101; F28F 2009/0287 20130101 |
Class at
Publication: |
165/185 |
International
Class: |
F28F 019/00 |
Claims
1. An integral primary and secondary heat exchanger of the cross
flow type having a regularly spaced, parallel series of flow tubes
with distinct fluids flowing through the primary and secondary
tubes, and a common external fluid flowing over the exterior
surface of all tubes, each primary and secondary tube having
adjacent internal flow passages with a respective, pre determined
internal web thickness between adjacent passages and a pre
determined optimal internal passage flow height measured in the
direction perpendicular to the flow of the common exterior fluid
that is determined by the flow characteristics of its respective
internal fluid, each tube also having a respective pre determined
minimum perimeter wall thickness, and in which the secondary tube
has a pre determined overall tube thickness, measured perpendicular
to the flow direction of the external fluid, that is determined by
its optimal internal passage flow height and minimum perimeter wall
thickness, and in which the optimal internal passage flow height of
the primary tube is substantially less than that of the secondary
tube, but in which it is desired to maintain the overall tube
thickness of all tubes substantially equal, characterized in that
said primary tube includes a series of flow passages extending
around and inboard of the perimeter thereof, each of which is
spaced from the perimeter of the tube by its respective minimum
perimeter wall thickness and is also spaced from each adjacent flow
passage by at least said pre determined web thickness, so as to
create an efficiently packaged array of flow passages within said
primary tube, while maintaining the constant overall tube thickness
across all tubes.
2. An integral heat exchanger according to claim 1 in which the
secondary flow tubes are oil cooler tubes and the primary flow
tubes are refrigerant condenser tubes.
3. A heat exchanger according to claim 2, in which the internal
flow passage height of the condenser tube is less than half that of
the oil cooler tube, whereby the condenser tube is arrayed with a
double row of flow passages about the tube perimeter.
4. A heat exchanger according to claim 3, in which said condenser
tubes are arranged in a single pass flow pattern between the
manifold tanks.
Description
TECHNICAL FIELD
[0001] This invention relates to structurally integrated condenser
and secondary heat exchanger, and specifically to a structurally
integrated condenser and oil cooler.
BACKGROUND OF THE INVENTION
[0002] An automobile requires several heat exchangers to dump waste
heat to the ambient, including radiators to cool the engine,
condensers to cool the air conditioning system refrigerant, and one
or more secondary heat exchangers to cool secondary fluids such as
engine oil and/or transmission fluid, and charge air coolers for
super charging systems. Since ambient air flow is available through
the vehicle grill, behind the grill has been the typical location
for all such heat exchangers. In a continuing effort at cost
reduction and simplification, and package size reduction, efforts
have been made to structurally integrate as many components of
these various front end heat exchangers as possible. For example,
condensers and radiators have been combined into a front-back
module, sharing a common tube header plate, as disclosed in U.S.
Pat. No. 5,509,199. Another and even older approach has been to
package one or more heat exchangers in an over-under configuration.
For example, U.S. Pat. No. 2,037,845 shows an over/under charge air
cooler and radiator, in which a series of parallel tubes extend
between common manifold tanks, with the interior of the common
manifold tanks divided into discrete sections for the two internal
fluids (air and engine coolant) by double separators. While there
is no particular discussion of tube thickness as measured in the
direction perpendicular to external air flow (sometimes referred to
as tube height), the tube thickness/height appears to be shown as
common for the tubes in both the radiator and the charge air cooler
sections.
[0003] Other patents disclosing over/under configurations for
discrete heat exchangers recognize that a common tube
thickness/height for the tubes in both heat exchangers is not
necessarily optimal from a heat transfer standpoint. Oil,
especially, is much thicker than engine coolant, and would require
a larger flow passage. The internal height of the flow passage,
together with tube wall thickness (determined by factors such as
internal pressure resistance), determines total tube
thickness/height. That is, overall tube thickness would be
approximately the internal flow passage height plus twice the tube
wall thickness. For fluids that are substantially less viscous than
oil, the optimal internal flow passage height, and consequent
overall tube thickness/height, would be substantially less. So, for
example, U.S. Pat. No. 4,923,001, which discloses an over/under
engine cooling radiator and oil cooler, shows substantially thicker
tubes for the oil cooler portion. Consequently, the oil cooler tube
slots in the common slotted header plate must be thicker. This
disparity in optimal tube thickness is even more pronounced when
the primary heat exchanger (the one largest in face area) is a
refrigerant condenser, and the secondary heat exchanger is an oil
cooler. It is well known that the optimal internal flow passage
height for a refrigerant condenser tube is on the order of 0.3-0.6
mm, see for, example, FIG. 3 of U.S. Pat. No. 4,492,268, where the
internal flow passage height is denoted at "d". This range of
optimal internal flow passage height, even at the upper end, is
substantially smaller than for oil, which may be two or three times
more. Thus, in U.S. Pat. No. 6,321,832, which shows an over/under
configuration of condenser and oil cooler, it is recognized that
the oil cooler tubes may have to be made thicker, and that the
corresponding header slots that accommodate them would be thicker
as well.
[0004] While thermal efficiency concerns dictate different
thickness tubes when there is a disparity in optimal fluid passage
internal height, manufacturing concerns dictate the opposite, since
any lack of uniformity in component size and spacing detracts from
manufacturing efficiency. Maintaining a common tube external
envelope (common tube depth and thickness) lends itself well to a
common header slot size and spacing, as well as a common air fin
height. Thus, in U.S. Pat. No. 6,394,176, an over/under combined
condenser and oil cooler is proposed which, just as U.S. Pat. No.
2,037,845 discloses, shows a common header tank divided by a double
separator, as well as a common tube outer envelope and air center
height. The common tube outer envelope (called a common tube cross
section) requires a compromise between the optimal flow passage
size of the two tubes, a compromise that is quantified as an
upper/lower range of the product of the hydraulic diameters of the
two different flow passages. Since hydraulic diameter is a function
of the flow passage cross-section and perimeter, it is essentially
a function of the flow passage internal height, as well. In fact,
for the typical square or round tube flow passage shape, the
hydraulic diameter and internal flow passage height (one side of
the square, or the diameter of the circle) are one and the same. In
effect, it is the accepted wisdom that the maintenance of a common
tube perimeter envelope or cross section will enforce a compromise
in thermal performance of the respective tubes in the two
operationally distinct (but structural integrated) heat exchangers.
In the aforementioned patent, this compromise in condenser
performance is manifested by the fact that the refrigerant flow
pattern is a three pass pattern, indicative of a condenser flow
passage size that is larger than the optimum, and optimum that
would allow for a one pass design.
SUMMARY OF THE INVENTION
[0005] The integrated primary and secondary heat exchanger of the
invention, which comprises a condenser and oil cooler, provides for
a common tube exterior envelope, and specifically for a common tube
thickness/height, but without compromising on the optimal interior
flow height of the refrigerant flow passage.
[0006] In the embodiment disclosed, the oil cooler tube passage
flow height is sized for its optimal performance, with a standard,
single row, side by side array of flow passages along the width of
the tube, thereby establishing the common tube thickness to be
maintained. The condenser tube has its substantially smaller
optimal internal height flow passages arrayed within the cross
section of the condenser tube in a novel fashion that allows the
use of and optimally sized flow passage in both tubes. Rather than
the standard, side by side array of flow passages, as in the oil
cooler tube, the smaller sized refrigerant flow passages are
arrayed along and around the perimeter of the tube, spaced from the
exterior surface thereof by a minimal tube wall thickness that is
established by conditions of tube material conductivity, strength,
manufacturing capabilities, and other concerns. Depending on how
much smaller the optimal refrigerant flow passages size is,
compared to the oil cooler tube, a staggered array of passages may
be created, or, ultimately, a double row of passages. The web of
tube material between adjacent flow passages is maintained at at
least the minimal thickness needed for burst pressure resistance.
This more efficient packaging of the optimally sized refrigerant
flow passages within a common size tube gives maximum assembly
efficiency for the heat exchanger as a whole, without undue
sacrifice in the thermal performance of the condenser.
BRIEF DESCRIPTION OF THE DRAWINGS
[0007] FIG. 1 is a front face view of an integral condenser and oil
cooler according to the invention.
[0008] FIG. 2 is a cross section through an oil cooler tube.
[0009] FIG. 3 is a cross section through one embodiment of a
condenser tube made according to the invention.
[0010] FIG. 4 is a cross section through another embodiment of a
condenser tube made according to the invention.
DESCRIPTION OF THE PREFERRED EMBODIMENT
[0011] FIG. 1 shows a simplified plan view of the face of an
integral condenser and oil cooler made according to the invention,
indicated generally at 10. One section of 10, the primary heat
exchanger, is a condenser, and is indicated generally at C. The
other, significantly smaller, is beneath the condenser C, and
indicated generally at O. That over/under configuration could be
reversed, and depends only on where the refrigerant and oil lines
run in any particular vehicle. The two sections of the heat
exchanger maintain a high level of part continuity and regularity.
Thus, two common header tanks 12 and 14 are each with a regularly
slotted header plate 16 and 18. These are numbered differently only
to indicate opposite sides of the core, but would in fact be nearly
identical parts. The condenser C has a series of condenser tubes
20, running in parallel from tank 12 to tank 14, and oil cooler O
has a smaller number of oil cooler tubes 22, all with the same
length (measured horizontally on the page, height/thickness
(measured vertically on the page) and width/depth (measured
perpendicular to the plane of the page). Heat exchanger 10 is the
so called cross flow type, meaning that one fluid flow internally
in one direction (right to left or left to right on the page),
while air or some other external fluid flows perpendicular to that.
Here, the external fluid is ambient air, but could be any common
fluid. Brazed or otherwise conductively joined between the facing
outer surfaces of the tube are a series of same height corrugated
air fins or air centers 24. Ambient air flows over the air centers
24, perpendicular to the plane of the page, and over the exterior
of all of the tubes 20 and 22, and centers 24 concurrently. In the
process, the oil flowing inside of tubes 22, and the refrigerant
inside of tubes 20, is cooled, and since the condenser is typically
mounted in front of the engine cooling radiator, the cooling air is
the "first in," unwarmed by any other heat exchanger. When the oil
cooler is incorporated with the radiator, as is more common, then
the air has typically already run through, and been warmed by, a
conventional condenser. As disclosed, the interior of each header
tank 12 and 14 is divided into discrete volumes to serve the
condenser C and oil cooler O by double walled, or otherwise
insulated, separators 26 and 28. The separators 26 and 28 allow the
condenser C and oil cooler O to operate independently, and either
tank 12 or 14, so divided, can serve as the inlet tank, or the
outlet tank, for either one. As noted above, the insulated
separators 26 and 28 are a well known means for dividing the common
header tank. In summary, the high level of common parts, including
the tanks 12 and 14, regularly slotted header plates 16 and 18,
same exterior size tubes 20 and 22, and common height centers 24,
maximizes assembly efficiency. As described next, this high degree
of uniformity does not negatively affect the thermal efficiency of
the condenser C.
[0012] Referring next to FIG. 2, a cross section through an oil
cooler tube 22 is shown. Oil cooler tube 22 is not novel in and of
itself, but is combined with the novel condenser tube 20 as
described below. Oil cooler tube 22 would typically be a unitary
aluminum extrusion, with a width or depth D in the direction of
airflow that is generally determined and limited just by the
packaging space available at the front of the vehicle. As
disclosed, D is approximately 18 mm. The overall shape is generally
flat, but occasionally, and somewhat misleadingly, referred to as
oval, because of the rounded leading and trailing edges. "Stadium
shaped" is a more accurate term to describe the overall shape that
is mostly flat, but with the edges rounded off. Inside of tube 22,
a series of side by side flow passages, seven internal passages 30
and two edge tube edge passages 32, are surrounded on the outside
by a perimeter wall 34 of thickness Two, and separated from one
another by webs 36 with a thickness Wo. The actual values of Two
and Wo are determined by several factors and limitations, including
the necessary tube strength, (internal burst pressure within the
passages 30 and 32), corrosion resistance of the wall 34, and basic
manufacturing limitations on how thinly webs can be extruded. Here,
Two is approximately 0.25 mm, and Wo is approximately 0.2 mm as
well. The internal dimensions of the passages 30 and 32 are
determined more by considerations of thermal performance, than by
manufacturing limitations. This was not always the case, especially
with the less advanced extrusion technology of a decade or two
ago.
[0013] To understand the relationship between flow passage size and
the refinement of extrusion technology over the last couple of
decades, some brief explication of flow passage size in general is
in order, as it relates to thermal efficiency. It has been a text
book truism for many decades that the internal flow passages in a
cross flow heat exchanger operate more efficiently as they
progressively decrease in cross sectional area, all other things
being equal. Quite simply, as the cross sectional area of the
passage decreases, so the internally flowing fluid presents
relatively more perimeter surface area to the externally flowing
air, compared to the internal volume enclosed by the perimeter, and
thus cools (or heats) more readily. This is intuitively obvious, in
the sense that a thin heated object will cool far more quickly in
an air stream than will a thick heated object. In order to relate
passages of differing cross sectional shapes (triangle, star
shaped, etc) to the simplest shape, which is circular, the concept
of "hydraulic diameter" was developed, which is a function of the
ratio of cross sectional area of the flow passage to its perimeter,
the units being length squared divided by length, or just length.
As tube passage cross sections grow smaller, that ratio decreases,
since relatively less area is enclosed by the perimeter. This is
generally stated in text books as heat transfer increasing as the
hydraulic diameter decreases. For simple passage shapes such as
square or circular, the hydraulic diameter is simply stated as just
the side of square or the diameter of the circle. For any design
case, an optimal hydraulic diameter can be calculated, taking into
account the fluid characteristics, the mass flow rate, tube length,
tolerable pressure drops, and similar factors.
[0014] In decades past, extrusion technology had not advanced to
the point where tube passages could be formed with an optimally
small size. Here, as indicated in FIG. 2, the optimal internal
passage flow height Ho (equal to the hydraulic diameter in the case
of a square passage) is approximately 1.5 mm, based on the flow
characteristics of the oil, the mass flow rate, tube length, etc.
That size passage can be produced by current, and even past,
extrusion technology, and yields an overall tube thickness Tto, of
approximately 2 mm (two times Two of 0.25 mm plus Ho of 1.5 mm).
That will vary for other particular cases, but it establishes an
overall tube thickness that it is desired to commonize over the
entire core 10.
[0015] As noted above, empirical tests had long ago determined
that, for condensing refrigerant, the optimal internal flow height
(for a simple shape passage) was in the range of 0.3-0.6 mm. This
optimally small size could only be practically provided in a
fabricated tube 20 or 30 years ago (such as separate webs brazed
between upper and lower plates), but now extrusion technology has
advanced to the point where extruded tubes can provide the same
small passage size and short internal flow height. Other
constraints, as noted above, still dictate a minimal perimeter wall
thickness, which, in this case, is about 0.31 mm, and that value,
plus the optimal internal flow height, together pre determine a
minimum overall tube thickness or height for a condenser tube. A
minimum web thickness between flow passages that will withstand
typical burst pressures in an aluminum tube is 0.2 mm, which does
not affect overall tube thickness, but is still a design
constraint. In the typical cross flow heat exchanger for automotive
application, the designer will want to use the minimum overall tube
thickness, and no more, so as to reduce air pressure drop and
resistance across the core, as well as tube weight and cost. But,
since the optimal internal flow height for refrigerant will
typically be so much smaller for refrigerant than for oil, a design
decision to standardize the total tube thickness for both the oil
cooler and condenser tube requires a condenser tube that is thicker
than it would ideally be in a condenser alone. Stated differently,
the common total tube thickness, minus the smaller internal flow
height of the refrigerant passage, yields a tube perimeter wall
thickness far greater than the minimum needed.
[0016] Referring next to FIG. 3, the design according to the
invention of a tube for a primary heat exchanger tube in general is
illustrated. The terms primary and secondary are arbitrary, and the
specific case of a condenser tube as the primary tube is consider
below. A primary tube 50 is illustrated in which the optimum
internal passage flow height H is substantially less than for the
oil cooler tube 22 (or other secondary tube), though not as small
as it typically would be for a condenser tube. Tube 50 has the same
depth D and total thickness T equal to that of oil cooler tube 22.
Stated differently, its exterior or outer envelope is uniform. A
minimum perimeter wall thickness Tw and and web thickness W as
illustrated are comparable to oil cooler tube 22. Round internal
flow passages 38 are shown which, as noted above, have a diameter
equal by definition to the hydraulic diameter and also equal to the
internal passage flow height H. Since the secondary tube, such as
the oil cooler tube 22, determines the total overall tube thickness
T, placing the substantially smaller flow passages 38 in a
standard, side by side array would give a perimeter wall thickness
that was substantially greater than the minimum needed. Instead,
according to the invention, the flow passages 38 are arrayed around
the perimeter of the tube 50, always within the minimum wall
thickness Tw from the perimeter, and always no closer to an
adjacent flow passage 38 than the minimum web thickness W. For the
case shown in FIG. 3, this effectively creates a perimeter wall 40
which, while it does not have a constant wall thickness, is no
thinner than the minimum thickness Tw at any point. Likewise, the
flow passages 38 are separated from adjacent flow passages 38 by
webs 42 that are no thinner than the minimum web thickness W at any
point. So, all the requirements of tube strength are met. While the
flow passage array is odd looking by conventional standards, more
flow passages 38 are packaged within the available envelope than
would be the case with a conventional, side by side array, and with
a perimeter wall 40 which, on average, is thinner than the overly
thick perimeter wall that would be created by such a conventional
array, so that weight and cost are reduced.
[0017] FIG. 4 illustrates the case when the primary tube is the
condenser tube 20, with its even smaller optimum internal flow
height Hc of approximately 0.5 mm, which is the same as both the
passage geometric diameter and its hydraulic diameter. This optimum
Hc is less than half that of the oil cooler tube 22, and small
enough to allow for a one pass flow pattern through all condenser
tubes 20. This indicates that no compromise in the thermal
performance of the condenser tube 20 or its value of Hc was made in
order to keep its total thickness equal to the oil cooler tube 22.
A minimum wall thickness Twc of 0.31 mm is maintained, and a
minimum web thickness Wc of 0.2 mm. Far more flow passages can be
packaged, creating an even thinner (on average) perimeter wall 46
that with tube 50, and webs 48 of adequate thickness. In effect, by
following the design constraint described, a double row of such
flow passages 44 is created, not just a staggered row, two rows of
23 in this specific case, with an extra end passage placed near
each rounded tube edge, for a total of 48. Only about half that
number of flow passages could be accommodated with a conventional,
side-by-side array. While the flow passages 44 are separated from
each other, in the direction normal to air flow, by significantly
more than the minimum web thickness needed (approximately 0.4 mm in
this case), there is not sufficient room for the addition of any
extra flow passages between the two rows. If, for some reason, the
optimum Hc were significantly less than illustrated (a different
refrigerant, a far lower mass flow rate, a shorter length tube,
etc), then there might be sufficient room between the rows. Such
extra flow passages would not be especially efficient, thermally
speaking, being barricaded from the perimeter of the tube 20 by the
perimeter passages 44, but might make sense in some cases, to
reduce tube weight, if nothing else. In any case, there will be at
least the array of flow passages around the perimeter of the
tube.
[0018] In conclusion, a condenser tube designed for use in a
typical, refrigerant only condenser core would, with flow passages
of the size shown in FIGS. 3 and 4, have much thinner tubes. For
example, in FIG. 4, with Hc equal to 0.5 mm, and Twc equal to 0.31
mm, total tube thickness would be only about 1.12 mm, almost half
as thick, with the flow passages arrayed side by side, if the tube
were designed for a condenser alone. To array the flow passages any
other way, and thereby increase the total tube thickness Ttc, would
be counter intuitive and counter productive, because of the added
air pressure loss. But, with the added manufacturing design
constraint of Ttc being equal to the thicker Two, the odd appearing
flow passage packaging scheme of the invention provides an
advantage, by packaging much more refrigerant flow area within the
same envelope, and reducing added metal, weight and cost to the
tube. Furthermore, the thermal performance of the two tubes 20 and
22 is not compromised to any great extent, as it is with
conventional, side by side flow passage arrangements, as are found
in other integrated, primary and secondary heat exchangers. That
is, the hydraulic diameter/internal flow passage height is
optimized for both the oil and condenser tubes, independently of
each other, and not limited to a compromise product of the two
variables that inevitably requires one variable to be too large as
the other approaches its optimum, and vice versa.
[0019] Variations in the disclosed embodiments could be made. As
noted, optimal values for the internal flow passage height will
vary with the fluid characteristics, flow rates, total tube lengths
and tolerable pressure losses. In general, for a combination of oil
cooling and refrigerant condensing, and with a typical tube length
found in a vehicle front end heat exchanger 500 to 900 mm), a range
of 1 to 3 mm for the internal passage flow height in the oil cooler
tubes 22, coupled with a range of 0.3 to 1.0 mm for the condenser
tubes 20 will allow for independent, thermally efficient operation,
with acceptable pressure drops, while maintaining a common
structural foot print. Again, this is tantamount to establishing a
range of hydraulic diameter for the flow passages, for the case of
the simple flow passage shapes involved. Flow passage shapes other
than simple round or square could be used, such as a square shape
enhanced by internal fins, a common way to increase effective flow
passage perimeter surface (and thereby decrease hydraulic diameter)
in the past, before extrusion technology had advanced to the point
where simple shapes could be extruded with a small enough internal
flow height to inherently have an optimally small hydraulic
diameter. Such a simple shape (round, square, rectangular), with a
simple internal flow height, is preferable with today's technology.
Regardless, even a finned or otherwise complex flow passage shape
will have an average or de facto internal flow height, so that
optimizing a flow passage hydraulic diameter is essentially the
same as optimizing the internal flow passage height, which is a
simpler quantity to deal with conceptually, especially in terms of
total tube thickness. The header tanks 12 and 14 need not be common
tanks, as shown, and might not be made as common tanks, if cross
contamination or thermal cross conduction between the two fluids
were a great concern. In other words, four separate header thanks
could be used, with no internal separators, but a common header
plate regularly slotted and with standard tube and center sizes,
could still be used, even if the tanks were separate for each of
the two heat exchanger sections. One pair of opposed slots could be
left empty, and one tube removed, so as to the adjacent ends of
such separate tanks. Such a design would still maintain almost all
of the manufacturing advantages of common component size.
* * * * *