U.S. patent application number 11/720014 was filed with the patent office on 2008-02-07 for dimensionally-optimized device for the exchange of heat and method for optimisation of the dimensions of devices for the exchange of heat.
This patent application is currently assigned to BEHR GmbH & Co., KG. Invention is credited to Mourad Ben-Salha, Gottfried Durr, Michael Kranich, Wolfgang Seewald, Karl-Heiz Staffa, Christoph Walter.
Application Number | 20080029242 11/720014 |
Document ID | / |
Family ID | 36046750 |
Filed Date | 2008-02-07 |
United States Patent
Application |
20080029242 |
Kind Code |
A1 |
Ben-Salha; Mourad ; et
al. |
February 7, 2008 |
Dimensionally-Optimized Device For The Exchange Of Heat And Method
For Optimisation Of The Dimensions Of Devices For The Exchange Of
Heat
Abstract
The invention relates to a device for the exchange of heat, in
particular, for a motor vehicle, comprising a number of flow tubes
for the transport of a fluid, whereby the device has a given depth
(T) and, at least in sections, some flow tubes are arranged at a
given separation from each other, whereby the depth and the given
separation are in a ratio (V) to each other.
Inventors: |
Ben-Salha; Mourad;
(Stuttgart, DE) ; Durr; Gottfried; (Stuttgart,
DE) ; Kranich; Michael; (Besigheim, DE) ;
Seewald; Wolfgang; (Stuttgart, DE) ; Staffa;
Karl-Heiz; (Stuttgart, DE) ; Walter; Christoph;
(Stuttgart, DE) |
Correspondence
Address: |
FOLEY AND LARDNER LLP;SUITE 500
3000 K STREET NW
WASHINGTON
DC
20007
US
|
Assignee: |
BEHR GmbH & Co., KG
Stuttgart
DE
70469
|
Family ID: |
36046750 |
Appl. No.: |
11/720014 |
Filed: |
November 17, 2005 |
PCT Filed: |
November 17, 2005 |
PCT NO: |
PCT/EP05/12304 |
371 Date: |
May 23, 2007 |
Current U.S.
Class: |
165/42 |
Current CPC
Class: |
F28D 1/05383 20130101;
F25B 2500/01 20130101; F25B 2309/061 20130101; F25B 9/008 20130101;
F28F 1/126 20130101; F28D 2021/0085 20130101; F25B 39/022
20130101 |
Class at
Publication: |
165/042 |
International
Class: |
F28D 1/053 20060101
F28D001/053; F25B 39/02 20060101 F25B039/02 |
Foreign Application Data
Date |
Code |
Application Number |
Nov 23, 2004 |
DE |
10 2004 056 557.0 |
Claims
1. A heat-exchanging device for an air-conditioning system, in
particular for a motor vehicle, having a plurality of throughflow
tubes for conveying a fluid, with the device having a predefined
depth T and a predefined number of throughflow tubes are arranged
at least in sections with a predefined spacing to one another,
wherein the ratio V between the depth and the predefined spacing is
less than 7 and/or the weighted ratio V' between the depth and the
sum of the predefined spacing and 10 mm is greater than 1.3 and
less than 2.8.
2. The device as claimed in claim 1, wherein the predefined spacing
is less than or equal to 9 mm, preferably less than or equal to 8
mm and preferably less than or equal to 6 mm.
3. The device as claimed in claim 1, wherein the ratio V is less
than 6.8, preferably less than 6.6 and preferably less than
6.3.
4. The device as claimed in claim 1, wherein the ratio V is less
than 6.1, preferably less than 5.9 and preferably less than
5.1.
5. The device as claimed in claim 1, wherein the weighted ratio V'
is at least 1.5, preferably at least 1.85 and preferably at least
2.2.
6. The device as claimed in claim 1, wherein the weighted ratio V'
is at most 2.6, preferably at most 2.4 and preferably at most
2.25.
7. The device as claimed in claim 1, wherein the throughflow tubes
are arranged parallel to one another at least in sections.
8. The device as claimed in claim 1, wherein the throughflow tubes
have a substantially constant predefined first spacing to one
another.
9. The device as claimed in claim 1, wherein the throughflow tubes
have a flat-tube-shaped cross section.
10. The device as claimed in claim 1, wherein the throughflow tubes
are formed in one piece, in particular from a single sheet metal
strip or extruded profile.
11. The device as claimed in claim 1, wherein the throughflow tubes
and preferably the device have a rupture pressure of over 90
bar.
12. The device as claimed in claim 1, wherein the fluid is a
refrigerant and is preferably R 744 (CO.sub.2).
13. The device as claimed in claim 1, wherein a first plurality of
throughflow tubes have a first predefined spacing to one another at
least in sections, a second plurality of throughflow tubes have a
second predefined spacing to one another, and the ratio V between
the depth and at least one of the predefined spacings is less than
7.
14. The device as claimed in claim 1, wherein a first plurality of
throughflow tubes is laterally offset with respect to a second
plurality of throughflow tubes.
15. The device as claimed in claim 1, wherein cooling fins are
arranged between the throughflow tubes.
16. The device as claimed in claim 1, wherein said device has a
depth of between 10 mm and 60 mm, preferably between 20 mm and 50
mm, and particularly preferably between 25 mm and 45 mm.
17. The device as claimed in claim 1, wherein said device the
predefined spacing is between 4 mm and 12 mm, preferably between
4.5 mm and 10 mm.
18. The device as claimed in claim 1, wherein a depth of between 30
mm and 50 mm, preferably a depth of between 35 mm and 45 mm is
assigned a predefined spacing of between 5 mm and 12 mm, preferably
between 5.5 mm and 10 mm.
19. The device as claimed in claim 1, wherein a depth of between 20
mm and 35 mm, preferably a depth of between 25 mm and 30 mm is
assigned a predefined spacing of between 3 mm and 10 mm, preferably
between 4 mm and 8 mm.
20. The device as claimed in claim 1, wherein the throughflow tubes
have a width of between 1 mm and 3 mm, preferably between 1.3 mm
and 2 mm and particularly preferably from approximately 1.4 mm to
1.9 mm.
21. The device as claimed in claim 1, wherein the throughflow tubes
have a wall thickness of between 0.1 mm and 0.6 mm, preferably of
between 0.2 mm and 0.4 mm and particularly preferably of
approximately 0.25 mm to 0.3 mm.
22. The device as claimed in claim 1, wherein the device is an
evaporator.
23. An air-conditioning system, in particular for a motor vehicle,
wherein said air-conditioning system has at least one
heat-exchanging device as claimed in claim 1.
24. A method for dimensioning heat-exchanging devices, having the
following steps: specifying a first dimension of the device;
specifying a second dimension of the device; determining at least
two first target parameters of the device; varying at least one
dimension; determining at least two second target parameters of the
device with the varied dimension; determining the more favorable
target parameters by comparing the first and second target
parameters.
25. The method as claimed in claim 24, wherein the dimensions are
selected from a group of dimensions which contains the depth, the
fin height and the spacing of the throughflow tubes.
26. The method as claimed in claim 24, wherein the parameters are
selected from a group of parameters which contains the installation
space depth, the refrigerating capacity, the volume flow rate, the
air-side pressure drop, the weight and the production costs.
27. The method as claimed in claim 24, wherein the target
parameters are determined multiple times and, from the sets of
target parameters which are determined in this way, the most
favorable sets of parameters are determined.
28. The method as claimed in claim 24, wherein in the determination
of the most favorable target parameter sets, the individual target
parameters are weighted according to predefined criteria.
Description
[0001] The present invention relates to a heat-exchanging device.
The device is described in the context of an air-conditioning
system, in particular for motor vehicles. It is however pointed out
that the device according to the invention can also be used in
other air-conditioning systems or refrigeration circuits.
[0002] From the prior art, air-conditioning systems or
heat-exchanging devices are known which, for cooling, use the
refrigerant R 134a. Also known are air-conditioning systems which,
instead of said refrigerant, use the refrigerant R 744, that is to
say carbon dioxide (CO.sub.2). The advantage of the use of CO.sub.2
over earlier refrigerants is inter alia its better environmental
compatibility, since said refrigerant does not lead to an increase
in the greenhouse effect.
[0003] However, in the prior art, the use of CO.sub.2 for
refrigeration circuits is associated with to some extent
considerable excess costs over conventional refrigerants, since
said refrigerant is under a significantly higher pressure in the
device than R 134a. It is thus for example the case that using the
same geometry or the same dimensions of the refrigeration circuit
as are used for conventional refrigerants results in a very high
weight and also high production costs, which lead to the production
of the devices becoming uneconomical. It is the object of the
invention to adapt individual heat-exchanging devices in terms of
their dimensions to the use of CO.sub.2 as refrigerant in such a
way as to permit more cost-effective and low-weight production.
[0004] Through extensive tests, it was possible to show that a
particularly efficient saving of weight and production costs can be
obtained by means of modifications to the evaporator. Here, it is
also conceivable to accept a moderate reduction in capacity of the
evaporator, since this, as has been shown from comprehensive
analysis, has the smallest effect on the cooling which can be
obtained in the vehicle.
[0005] The object is to adapt the devices, by adapting certain
dimensions, in such a way as to obtain an improvement of the
device, in particular when using the refrigerant CO.sub.2, but in
particular not only in terms of its production costs, capacity,
weight etc.
[0006] A further object is also to improve heat-exchanging devices
which use R 134a as refrigerant.
[0007] The object is achieved according to the invention in that
the evaporator as a component is reduced in terms of its specific
refrigerating capacity by an order of magnitude at which the
repercussions in the refrigeration circuit in the vehicle cabin are
still acceptable. Here, it can be accepted that the capacity level
of a refrigeration circuit using a conventional refrigerant (R
134a) is no longer considerably outperformed as before, but is
rather at a comparable level. More precisely, the evaporator should
be designed so as to be comparable, in terms of its refrigerating
capacity, its weight and its production costs, to evaporators with
conventional refrigerants.
[0008] According to the invention, significant geometric dimensions
of the evaporator are optimized so as to obtain the most favorable
possible cost/benefit ratio within the context of the entire
system.
[0009] The object is achieved in detail by means of the subject
matter of claim 1. Advantageous embodiments and developments are
the subject matter of the subclaims.
[0010] The heat-exchanging device according to the invention has a
plurality of throughflow tubes for conveying a Fluid, with the
device having a predefined depth--also referred to below as the
installation depth--and some throughflow tubes are arranged at
least in sections at a predefined spacing to one another. In this
way, according to the invention, the ratio between the depth and
the predefined spacing is less than 7. The depth of the
heat-exchanging device results substantially from the depth of the
individual throughflow tubes, as will be explained in detail with
reference to the figures.
[0011] The tube spacing of the individual throughflow tubes is to
be understood here, as can be seen more precisely with reference to
the figures, as the spacing by which the sides, which face toward
one another in each case, of the throughflow tubes are spaced apart
from one another. Said tube spacing also determines the height of
the fins which are preferably arranged between the tubes. The tube
spacing is therefore also referred to below as the fin height.
[0012] Here, the spacing is to be understood to mean the shortest
geometric spacing between the throughflow tubes. The spacing which
is predefined at least in sections is to be understood to mean that
the tubes need not strictly have the same spacing from one another
along their entire length.
[0013] It is also possible for a first group of tubes to have a
first spacing to one another and for a second group of tubes to
have a second spacing to one another. This will also be explained
in detail in connection with the figures.
[0014] In a further preferred embodiment, the ratio V is less than
6.5, preferably less than 6.3 and particularly preferably less than
5.9. Through tests and analysis, it was possible to determine that
said ratios lead, when using CO.sub.2 as refrigerant, to a
particularly favorable cost/benefit relationship, with in
particular the specific refrigerating capacity, the air-side and
refrigerant-side pressure drop and the production costs and the
weight being possible criteria for evaluating the cost/benefit
ratio.
[0015] In a further preferred embodiment, the throughflow tubes are
arranged parallel to one another at least in sections. A
substantially constant spacing between the individual throughflow
tubes can be ensured in this way.
[0016] The throughflow tubes are preferably parallel to one another
substantially along their entire length and in this way have a
constant predefined first spacing to one another along
substantially their entire length.
[0017] In a further preferred embodiment, the throughflow tubes
have a flat-tube-shaped cross section. A flat-tube-shaped cross
section is to be understood as a cross section in which one side by
far exceeds a further side in terms of its length, such as for
example an elongated rectangle, an elongated rectangle with rounded
corners or an ellipse in which the first diameter is considerably
greater than the second diameter.
[0018] In a further preferred embodiment, the fluid is a
refrigerant and is preferably R 744 (CO.sub.2).
[0019] In a further preferred embodiment, a first plurality of
throughflow tubes have a first predefined spacing to one another at
least in sections, and a second plurality of throughflow devices
have a second, substantially predefined spacing to one another,
with the ratio V between the depth and at least one of the
predefined spacings being less than 7.
[0020] This means that individual throughflow devices have a
different spacing from one another than other throughflow devices.
Here, the individual spacings can also vary within throughflow
devices which are substantially parallel to one another. In
addition, both the first predefined spacing and the second
predefined spacing can be dimensioned such that the ratio between
the depth and both predefined spacings is in each case less than
7.
[0021] It is preferable, in a heat-exchanging device in which the
fluid is conveyed in a certain direction in a first plurality of
throughflow devices and in a further plurality of throughflow
devices, in which the fluid is conveyed in another direction, for
in each case different predefined tube spacings to be selected. In
this way, it is possible under some circumstances to obtain a more
cost-effective design in terms of the obtained heat transfer
capacity.
[0022] In a further preferred embodiment, a first plurality of
throughflow devices is laterally offset with respect to a second
plurality of throughflow devices. Here, the individual predefined
spacings of the first and of the second plurality can be selected
to be the same or different. The predefined spacings can also vary
within the same plurality of throughflow devices.
[0023] In a further preferred embodiment, cooling fins are arranged
between the throughflow tubes. Said cooling fins serve to improve
the exchange of heat with the surrounding air. Here, as mentioned,
the height of said cooling fins is substantially determined by the
predefined spacing of the respective throughflow tubes which bear
against them.
[0024] In one preferred embodiment, the wall thickness of the
individual cooling fins is between 0.04 and 0.2 mm, preferably
between 0.05 and 0.12 mm and particularly preferably between 0.06
and 0.1 mm. The fin density is between 40 and 90 fins/dm,
preferably between 50 and 80 fins/dm and particularly preferably
between 60 and 70 fins/dm.
[0025] In a further preferred embodiment, the device has a depth of
between 10 mm and 60 mm, preferably between 20 mm and 50 mm, and
particularly preferably between 25 and 45 mm. Said different depths
are determined in particular by the intended application, that is
to say for example by whether the device is to be used in a compact
vehicle, a medium-sized vehicle or a large vehicle.
[0026] In a further preferred embodiment, the predefined spacing
between the throughflow tubes is between 4 mm and 12 mm, preferably
between 4.5 mm and 10 mm. Said spacings are also determined in
particular by the respective applications.
[0027] In a further preferred embodiment, a depth of between 30 mm
and 50 mm, preferably a depth of between 35 mm and 45 mm is
assigned a predefined spacing of between 5 mm and 12 mm, preferably
between 5.5 mm and 10 mm. Said embodiment concerns more
largely-dimensioned heat-exchanging devices which can be used in
particular, but not exclusively, in air-conditioning systems in
medium-sized vehicles or large vehicles. Here, however, the
selected dimensions substantially ensure that the ratio remains
lower than 7.
[0028] In a further preferred embodiment, a depth of between 15 mm
and 40 mm, preferably a depth of between 20 mm and 35 mm is
assigned a predefined spacing of between 3 mm and 10 mm, preferably
between 4 mm and 8.5 mm. Said dimensions or measurements are used
in particular in air-conditioning systems of compact vehicles and
medium-sized vehicles.
[0029] Said dimensions, too, should substantially ensure a ratio of
less than 7. Here, however, a ratio of substantially 7 is also to
be understood as a ratio which slightly exceeds the value 7.
[0030] In a further preferred embodiment, the throughflow tubes
have a width of between 1 mm and 3 mm, preferably between 1.5 mm
and 2 mm and particularly preferably between 1.7 mm and 1.9 mm. The
wall thickness of the throughflow tubes is between 0.1 mm and 0.6
mm, preferably between 0.2 mm and 0.4 mm and particularly
preferably in the region of approximately 0.3 mm. With said
dimensions, it is possible to obtain a particularly advantageous
exchange of heat with the ambient air.
[0031] The device according to the invention is preferably an
evaporator which is a component of a refrigeration circuit of a
motor vehicle air-conditioning system.
[0032] The invention is also aimed at an air-conditioning system,
in particular for a motor vehicle, which has at least one
heat-exchanging device according to the invention.
[0033] The invention is also aimed at a method for dimensioning
heat-exchanging devices, in which method a first dimension of the
device is specified in a first step, a second dimension of the
device is specified in a further step, at least two first target
parameters of the device are determined in a further step, at least
one dimension is varied in a further step, two second target
parameters of the device with the varied dimension are in turn
determined from the varied dimension, and finally the more
favorable target parameters are selected by comparing the first and
second target parameters.
[0034] The first and second dimensions are preferably selected from
a group of dimensions which contains the depth, the fin height of
the cooling fins and the spacing of the throughflow tubes and the
like.
[0035] Dimensions can however also be understood to mean variables
such as the fin density per dm and the like.
[0036] Target parameters are preferably selected from a group of
parameters which contains the installation space depth, the
refrigerating capacity, the air-side pressure drop, the weight and
the production costs. As mentioned in the introduction, said
factors ultimately determine the benefit or value of the
heat-exchanging device for the different refrigerants, in the
present case for R 134a and R 744 (CO.sub.2). Using the method
according to the invention, the significant dimensions of the
heat-exchanging device can be varied and thereby in each case the
related stated output variables determined in order to thereby
arrive at a device which is dimensioned so as to provide
satisfactory, sufficient refrigerating capacity with acceptable
weight, with acceptable production expenditure and acceptable
costs.
[0037] In said method, it is to be taken into consideration that
even small changes to one or the other dimension can lead to
drastic changes in an output variable or a target parameter.
[0038] The target parameters are preferably determined multiple
times in particular for different dimensions, and, from this
plurality of determined sets of target parameters which are
determined in this way, the most favorable sets of parameters are
determined. With said multiple determination of the target
parameters, it is possible to provide very accurate analysis of the
capacities or target parameters of the heat-exchanging device which
can be expected. In the determination of the most favorable target
parameter sets, the individual target parameters are preferably
weighted according to predefined criteria. It is thus for example
possible, where the device is to be used in a large vehicle, for
the weight and production cost target parameters to be weighted
lower than in the case of application in a compact vehicle.
[0039] Further advantages and embodiments of the device according
to the invention and of the method according to the invention can
be gathered from the figures, in which:
[0040] FIG. 1 shows a plan view of a detail of the device according
to the invention;
[0041] FIG. 2 shows a side view of the device according to the
invention from FIG. 1;
[0042] FIG. 3 is a schematic illustration of a further
embodiment;
[0043] FIG. 4 is a schematic illustration of a further
embodiment;
[0044] FIG. 5 is a schematic illustration of a further
embodiment;
[0045] FIG. 6 is a schematic illustration for clarifying the tube
spacings;
[0046] FIG. 7 shows a diagram for clarifying the cooling which is
obtained;
[0047] FIG. 8 is an illustration for analyzing the individual
components;
[0048] FIG. 9a is a graphic illustration of the ratio between
refrigerating capacity and weight of the device according to the
invention;
[0049] FIG. 9b is an illustration of the air-side pressure
drop;
[0050] FIG. 10 is an illustration of the capacity as a function of
the installation depth;
[0051] FIG. 11 is an illustration of the ratio of capacity to
weight as a function of the installation depth;
[0052] FIG. 12 is an illustration of the capacity in relation to
costs as a function of the installation depth;
[0053] FIG. 13 is an illustration of the capacity as a function of
the installation depth to fin height;
[0054] FIG. 14 is an illustration of the capacity in relation to
the weight as a function of the installation depth in relation to
the fin height; and
[0055] FIG. 15 is an illustration of the capacity in relation to
the costs as a function of the ratio of installation depth to fin
height.
[0056] FIG. 1 shows a plan view of a detail of the heat-exchanging
device 1 according to the invention. Said heat-exchanging device 1
has a plurality of first throughflow tubes 3 and a second plurality
of second throughflow tubes 5. In a preferred embodiment, the
refrigerant flows through the plurality of first throughflow tubes
3 in one direction, for example out of the plane of the page, and
in an opposite direction, that is to say into the plane of the
page, in the second plurality of throughflow tubes 5.
[0057] The reference sign 7 denotes a chamber of the throughflow
tube. The throughflow tubes are preferably divided into a plurality
of chambers or ducts.
[0058] Here, the first throughflow tubes 3 and second throughflow
tubes 5 are separated from one another by an intermediate space 8.
Said intermediate space 8 serves for heat insulation, since the
refrigerant in the throughflow tubes 3 and 5 can have a different
temperature, and heat transfer should not take place. Instead of
the intermediate space, it is however also possible for the
throughflow tubes to be arranged continuously along the depth T,
that is to say for only one plurality of flat tubes to be provided.
In this case, one chamber or one duct 7 is preferably designed to
be dead-ended, that is to say no refrigerant flows in said
duct.
[0059] The reference symbol 4 relates to fins which are arranged
between the throughflow tubes 3 and 5 and are shown here in a plan
view from above. The dimension H.sub.Ri denotes the fin height and
is determined substantially by the spacing of the individual
throughflow tubes 3 and 5, more precisely by the spacing of the
sides, which face toward one another in each case, of the
respective throughflow tubes 3 and 5.
[0060] The reference symbol T denotes the installation depth which,
as mentioned above, constitutes a significant geometric variable of
the device. The fins 4 extend substantially along the entire depth
T and are preferably also not interrupted by intermediate spaces.
The ratio V mentioned above determines the ratio of installation
depth T to the fin height H.sub.Ri.
[0061] FIG. 2 shows a side view of the illustration, of which a
detail is shown in FIG. 1, of the heat-exchanging device. Here, b
denotes the tube width of the individual throughflow tubes. In the
case of a heat-exchanging device which utilizes R 134a as
refrigerant, the width of the tubes is between 2 and 4 mm,
preferably between 2.5 and 3 mm.
[0062] In a heat-exchanging device which utilizes CO.sub.2 as
refrigerant, the width of the tubes is preferably--as mentioned
above--in a range from 1.2 to 2 mm. The device has an overall width
of between 120 and 400 mm, preferably between 215 and 350 mm and
particularly preferably between 250 and 315 mm. A further
advantageous width is between 120 and 315 mm. The height of the
device according to the invention is between 140 and 300 mm,
preferably between 200 and 300 mm, particularly preferably between
220 and 250 mm. A further advantageous height is between 140 and
270 mm. In one preferred embodiment, the device is produced
substantially from aluminum or an aluminum-containing material.
[0063] The reference symbol A denotes the so-called transverse
pitch, that is to say the spacing of the respective geometric
centers of the individual throughflow devices to one another. Said
transverse pitch A results in the fin height H.sub.Ri when the
respective tube width b is also taken into consideration, that is
to say the fin height and the transverse pitch are directly
related. The transverse pitch can be incorporated as a measure for
the fin height if, on account of the cross section of the
throughflow tubes 3, 5, there is no geometrically unequivocal and
constant value for the fin height or for the spacing of the
throughflow tubes, for example if the spacing of the throughflow
tubes in FIG. 2 varies in a direction perpendicular to the plane of
the page, which is for example possible in the case of a circular
profile of the throughflow tubes. In this case, the ratio according
to the invention of the depth and spacing of the tubes is to be
replaced by the ratio of depth and transverse pitch.
[0064] FIG. 3 schematically illustrates a further embodiment of the
device according to the invention. Here, the reference symbols 3
and 5 relate in each case to plan views of the individual
throughflow tubes. In contrast to the embodiment shown in FIG. 1,
here, the throughflow tubes 3 and the throughflow tubes 5 are
laterally offset with respect to one another. This means that the
spacing between the throughflow tubes can be determined separately
for the throughflow tubes 3 and for the throughflow tubes 5. In the
exemplary embodiments shown in FIG. 3, the spacing H.sub.Ri the
throughflow tubes 3 is identical to the spacing H.sub.Ri the
throughflow tubes 5.
[0065] FIG. 4 schematically illustrates a further embodiment of the
device according to the invention. In this case, the throughflow
tubes 3 have a greater spacing H.sub.ri from one another than the
throughflow tubes 5 which have a spacing of H.sub.Ri2 from one
another. Here, preferably at least one of the two spacings
H.sub.Ri1 or H.sub.Ri2, in this case at least the spacing
H.sub.Ri1, is selected such that the ratio of the depth T and the
spacing H.sub.Ri1 is less than 7. It is however also possible to
select both spacings such that the corresponding ratio is less than
7.
[0066] FIG. 5 illustrates a further embodiment of the device
according to the invention. In this embodiment, the spacings
between the individual throughflow tubes vary only within the
throughflow tubes 3. It is however also possible for the spacings
to vary only within the tubes 5, or else both within the
throughflow tubes 3 and the throughflow tubes 5. In this
embodiment, too, it must be ensured that at least one of the
spacings H.sub.Ri meets the criterion that the ratio of the depth
and said spacing is less than 7.
[0067] It would also be possible to provide further different
spacings or a plurality of different spacings between the
individual tubes, such as for example spacings H.sub.Ri1,
H.sub.Ri2, H.sub.Ri3, etc. In any case, the above ratio, which is
less than 7, would have to be maintained for one of the spacings
H.sub.Ri.
[0068] FIG. 6 is a schematic illustration for clarifying the
definition of the spacing H.sub.Ri. While the throughflow tubes in
FIGS. 3 to 5 in each case have rectilinear longitudinal sides,
which simultaneously directly determine the spacing, the
throughflow tubes in the embodiment shown in FIG. 6 have an
elliptical cross section. In this case, the spacing between the
throughflow tubes is defined as the spacing of the two tangents T
which are each placed against the throughflow tubes 3.
[0069] It is however also possible, as indicated above, to define
the tube spacing not by means of the spacing of the sides which
face toward one another, but rather by means of the spacing of the
respective geometric center line of the individual throughflow
tubes, as was referred to above as the transverse pitch. This lends
itself to use, as stated, primarily when throughflow tubes have
geometries which deviate from the geometries shown here, such as
for example concave or convex shapes.
[0070] The diagram shown in FIG. 7 shows the simulation of a
cooling curve for a large vehicle. Here, comparable cooling curves
have been illustrated for the coolant R 134a, here by curves 11 and
12, and for R 744, here illustrated by the curves 14 and 15, in
each case in the idle operating point.
[0071] The upper curves 12 and 14 show the temperature profile in
the vehicle interior, and the lower curves 11 and 15 show the
temperature generation at the evaporator itself.
[0072] For the simulation, it was also assumed that the R 744
evaporator has an installation depth which is 25 mm smaller,
specifically an installation depth of 40 mm, whereas the R 134a
evaporator has an installation depth of 65 mm.
[0073] Plotted on the ordinate is the time in minutes, and on the
abscissa the temperature in degrees Celsius. The simulation is
divided into several time sections I to IV, with section I relating
to driving in 3.sup.rd gear at 32 km/h, section II relating to
driving in 4.sup.th gear at 64 km/h, section III relating to idle,
and section IV relating to driving in 2.sup.nd gear at 64 km/h.
[0074] It can be seen that even in 3.sup.rd gear (I) the R 744
evaporator provides more efficient cooling than the R 134a
evaporator. In regions II to IV, the respective evaporators obtain
in each case substantially the same values.
[0075] FIG. 8 illustrates a capacity comparison of different
evaporator designs at a typical operating point. Here, said
operating point is defined so as to permit comparisons which are
independent of the refrigeration circuit.
[0076] It is pointed out that the method and the obtained results
which are described in the following can be incorporated equally
for improving both R 134a evaporators and also R 744 (CO.sub.2)
evaporators.
[0077] In the diagram illustrated here, the assumptions are made of
an air mass flow rate GLV of 8 kg/min, an air inlet temperature
tLVe=40.degree. C., and a relative humidity .phi.LVE of 40%.
[0078] In the diagram, the rhombuses show the values determined for
the refrigerant R 744 (CO.sub.2) and the ellipses show the values
determined for the refrigerant R 134a.
[0079] The fin density is 70 fins/dm for the evaporator with the
refrigerant R 744, and 60 fins/dm for the evaporator with the
refrigerant R 134a.
[0080] Plotted on the ordinate is the installation depth in mm, and
on the abscissa the total capacity in kW. The plotted value pairs
or points 31 to 39 are functions of the temperature T, the fin
height H.sub.Ri, the fin density z.sub.ri and the so-called
transverse pitch s.sub.q. The transverse pitch denotes the spacing
of the respective centers of the individual throughflow tubes from
one another. Here, the individual value pairs or points 31 to 39
span a field which covers the capacity level in refrigeration
circuits of different vehicle classes. Here, the upper curve 22 is
the large vehicle or van segment, and the lower limit curve 23
shows the capacity requirement of compact vehicles.
[0081] The values for the refrigerant R 744 are plotted for
installation depths smaller than 40 mm, that is to say for the
measurement points 31 to 35. The values for the refrigerant R 134a
are plotted for the installation depth region .SIGMA. 40 mm. As
mentioned above, a uniform fin density of 70 fins/dm was selected
for the measurement points 31 to 35, while a uniform fin density of
60 fins/dm was selected for points 36 to 39.
[0082] In the case of the measurement points 31 and 32, a
relatively small transverse pitch was used, and in the case of the
measurement points 33 to 35, a relatively high transverse pitch was
used. The relatively small transverse pitch results in a likewise
small fin height, which is indicated by the line 28. The relatively
high transverse pitch likewise results in a relatively high fin
height, which is shown by the line 27.
[0083] For the measurement points 36 and 37, a relatively small
transverse pitch was selected, which leads to a relatively small
fin height H.sub.Ri, as shown by the line 25. For the measurement
points 38 and 39, a relatively high transverse pitch was selected,
which leads to a relatively high fin height, as shown by the line
26.
[0084] It can be seen from the diagram that the installation depth
for R 744 is considerably reduced for a constant capacity level,
which is plotted on the abscissa. This means that the assignment of
the installation depth T to fin height H.sub.Ri or the ratio is
moved.
[0085] While, in the case of R 134a, a depth of 65 mm is assigned
to a fin height of 7 to 10 mm and a depth of 40 mm is assigned to a
fin height of 4 to 6 mm, when using the refrigerant R 744, a depth
of 40 mm is assigned to a fin height of 7 to 10 mm and a depth of
27 mm is assigned to a fin height of 5 to 8 mm. In earlier designs,
for the refrigerant R 744, the assignment or dimensioning of R 134a
was assumed. This led to considerably higher capacity values than R
134a but also to excess weight and to excess costs, which is caused
inter alia by the considerably higher pressures required with R
744. Said considerably higher capacity values are plotted by way of
example by the points 41 and 42. The capacities at points 41 and 42
are around over 15' above the maximum required capacities.
[0086] It was thereby possible to show that, counter to the
expectations among experts, cost-reducing and weight-reducing
modifications in dimensioning are possible without thereby having
to simultaneously accept reductions in cooling capacity.
[0087] The considerably higher potential in R 744 is based on the
fact that, on account of the high specific delivery rate of the R
744 compressor in the R 744 circuit, a faster pressure reduction is
obtained in the low-pressure part. This leads to improved dynamics
and, at the evaporator, to a higher driving pressure gradient
between the air and the refrigerant.
[0088] The refrigerant-side pressure drop in the evaporator is of a
comparable order of magnitude, with 1 bar pressure drop with R 134a
causing approximately 9K temperature variation, and with R 744,
only 1K. This leads, on average, to a considerably higher driving
pressure gradient between the air and the refrigerant over the flow
length in the evaporator (the R 744 evaporator provides, on
average, a considerably colder surface temperature).
[0089] As mentioned in the introduction, it is sought to provide a
cost/benefit optimum as a function of the variables of installation
space depth, refrigerating capacity, air-side pressure drop, weight
and costs. Here, as mentioned in the introduction, the variables
are the depth T, the fin height H.sub.Ri and the tube spacing, and
variables derived from said variables, like the transverse
pitch.
[0090] According to previous considerations and tests, an
installation depth of 65 mm for the present capacity level is
rather too large; according to estimations, a 55 mm deep design
which reaches the level of the 65 mm depth would be more favorable.
However, an embodiment of said type possibly leads to higher costs
and to a less favorable air-side pressure drop. For the refrigerant
R 134a, a depth of 40 mm has proven to be particularly favorable in
terms of the capacity; in this case, however, disadvantages with
regard to costs and the air-side pressure drop are to be expected.
Said considerations show the extremely complex interrelationships
between different aspects in the assessment and evaluation of the
evaporators to be produced.
[0091] In the case of an evaporator which utilizes the refrigerant
R 744, an installation space depth of between 25 and 45 mm has
proven to be particularly suitable.
[0092] The diagram shown in FIG. 9 displays some of the advantages
of the invention. Here, the sub-diagram denoted by FIG. 9a
illustrates the weight of the evaporator in relation to the
obtainable refrigerating capacity. The physical boundary conditions
such as for example the air mass flow rate GLV are identical to the
conditions which formed the basis in the description of FIG. 8. The
same evaporator dimensions were likewise selected.
[0093] As shown by the measurement points 44 and 45, which relate
to compact vehicles and medium-sized vehicles, it is possible by
adapting the geometric dimensions to obtain comparable
refrigeration capacities, with the measurement point 44 having been
determined for the refrigerant R 744 and the measurement point 45
having been determined for the refrigerant R 134a. In the case of
the measurement point 45, a medium installation depth and a fin
density of 60 fins/dm was used. In the case of the measurement
point 44, a smaller depth, a smaller transverse pitch and a higher
fin density were used than in the case of point 45.
[0094] The two measurement points 46 and 47, which relate to
devices for large vehicles, also show, for the same refrigerating
capacity, a considerably reduced weight of the R 744 evaporator.
For the measurement point 46, a relatively high installation depth
T, a predefined fin density and a relatively high transverse pitch
Sq were selected. In the case of the measurement point 47 for the R
744 evaporator, a smaller depth T than at point 46, the same fin
density as at point 46 and a correspondingly identical transverse
pitch were selected. The result is therefore a considerable weight
reduction, as a result of the smaller installation depth with
otherwise identical transverse pitch, and even weight advantages
over the respective identical-capacity R 134a evaporator. The
smaller installation depth also results in only relatively low
material costs, and therefore a cost reduction.
[0095] In addition, it is possible for evaporators for large
vehicles to obtain an installation depth reduction from 65 to 40
mm, and in the case of compact vehicles, from 40 to 25 mm. This
brings with it the additional advantage that less installation
space is required in the motor vehicle.
[0096] As the diagram in FIG. 9b shows, the air-side pressure drop,
which is illustrated on the abscissa, can also be reduced. The
blocks 51 to 53 relate to the refrigerant R 134a, the blocks 54 to
55 to the refrigerant R 744. It can be seen that, when using R 744,
a considerable reduction in the air-side pressure drop by
approximately 50% is also obtained. This leads to an increased air
quantity for the air-conditioning of the vehicle, to a reduced
power consumption in the fan and also offers the potential for
reducing the noise level in the air-conditioning unit.
[0097] In FIG. 10, the capacity values of individual evaporators
are plotted against the installation depth on the ordinate. Here,
both for the CO.sub.2 and for the R 134a evaporators, the
evaporators with the same fin height are situated in each case on
one line. The reference symbol 63 denotes the line which is
associated with a large fin height, which is referred to below as
the first fin height, the reference symbol 62 denotes the line
which is associated with a second, smaller fin height (referred to
below as the second fin height), and the reference symbol 61
denotes the line which is associated with a fin height (referred to
below as the third fin height) which is smaller still than the
second fin height.
[0098] As can be seen from FIG. 10, the individual lines 61 to 63
have relatively similar gradients, which suggests a proportional
dependency of capacity and installation depth for an otherwise
identical design or fin height. It can also be seen that
evaporators with smaller fin heights but otherwise identical
dimensions have higher capacities on account of the enlargement of
the heat-transmitting surface.
[0099] The hatched regions 60 and 70 delimit the required or
expedient capacity values. The capacity limits have been determined
inter alia by means of the simulation of vehicle cabin cooling.
While, in the upper region 60, a further capacity increase no
longer brings with it any advantages, the cabin cooling below the
lower limit in the region 70 is no longer acceptable. The reference
symbols 65 to 68 show measurement values which lie within the
required capacity region. Said reference symbols 65 to 68 denote
devices of different design.
[0100] The reference symbol 67 relates to an R 134a evaporator with
a large installation depth and the first above-stated fin height.
The reference symbol 65 refers to an R 134a evaporator with the
third above-stated fin height and a relatively small installation
depth.
[0101] The reference symbol 66 relates to an R 134a evaporator with
the second fin height and a medium installation depth.
[0102] The reference symbol 68 refers to an R 134a evaporator with
the first fin height and a medium installation depth. The reference
symbols 71 to 74 show the measurement values of those evaporators
which no longer lie in the tolerable range 75 between the regions
60 and 70. Here, the reference symbol 71 denotes an evaporator with
a small installation depth and the first fin height, the reference
symbol 72 denotes an R 744 evaporator with the third fin height and
a very small installation depth, the reference symbol 73 denotes a
CO.sub.2 evaporator with a small installation depth and the third
fin height, and the reference symbol 74 denotes a CO.sub.2
evaporator with a large installation depth and the first fin
height.
[0103] It can therefore be seen that the CO.sub.2 evaporators have
significantly higher capacity values at the given installation
depths and fin heights than the R 134a evaporators. As can likewise
be seen from FIG. 10, a CO.sub.2 evaporator which has a small
installation depth for example with the second fin height, as shown
by the line 76, could be of interest for the application. The
ellipses 140, 141 indicate regions in which favorable dimensions
are situated.
[0104] FIG. 11 illustrates the relationship of capacity to weight
as a function of the installation depth. Here, the related
variables capacity/weight have again been weighted relative to one
another in order to give consideration to the different
significance of the individual variables. In a further preferred
variant of the method, the capacity and the costs are considered as
variables of equal significance, while the weight and the fin
height are of secondary importance.
[0105] Here, in the diagrams shown in FIGS. 11 to 15, the
weightings have been applied in such a way that the capacity is
weighted in the ratio 50:50 in relation to the costs, the capacity
is weighted in the ratio 80:20 in relation to the weight, and the
installation depth is weighted in the ratio 70:30 in relation to
the fin height. The triangles relate in each case to CO.sub.2
evaporators, and the circles to R 134a evaporators. Since the ratio
capacity/weight is plotted on the abscissa, higher values, that is
to say a ratio displaced more in the direction of capacity, are to
be considered more favorable.
[0106] It can thereby be seen that, for the R 134a evaporators, the
evaporator denoted by the reference symbol 81, with a medium
installation depth and an above-stated second fin height, and the
evaporator denoted by the reference symbol 83, with a small
installation depth and the third fin height, prove to be
particularly favorable.
[0107] Although the evaporator denoted by the reference symbol 84,
with the first fin height, would also lie favorably in relation to
the ratio of capacity to weight, its absolute capacity would
however no longer be acceptable even for the cooling of compact
vehicles. Said evaporator type could conceivably be used for
example in a rear system. An evaporator with the same fin height in
the region of the second fin height could likewise for example be
considered as a further alternative for compact and/or medium-sized
vehicles.
[0108] For the refrigerant CO.sub.2, therefore, the evaporators
denoted by the reference symbols 86 and 87, with a small depth with
a large fin height, and the evaporator denoted by the reference
symbol 88, with a small depth and the third fin height, are most
favorable. Finally, the evaporator denoted by the reference symbol
89, with a relatively small depth, also lies in a relatively
favorable position, but is borderline in terms of its capacity.
[0109] The evaporator 91 is less favorable in a direct comparison
with the first fin height. In addition, said evaporator is already
situated above the presently required upper capacity limit.
[0110] The evaporator denoted by the reference symbol 92 has an
unfavorable ratio of capacity to weight on account of the high
packing density of tubes and fins, with an otherwise excessively
low capacity.
[0111] The reference symbols 95 and 96 relate to trend lines which
have been specified on the basis of the measured values. On the
basis of said trend lines, it is possible to determine or estimate
those dimensions of the evaporator with which favorable designs,
such as here a favorable capacity/weight ratio, are to be
expected.
[0112] In detail, the trend line 95 relates to CO.sub.2 evaporators
and the trend line 96 relates to R 134a evaporators.
[0113] FIG. 12 illustrates the ratio of capacity to production
costs as a function of the installation depth. Here, in the case of
the ratio of capacity to costs, the abovementioned ratio or the
above-specified weighting has again been incorporated.
[0114] It can be seen that, in the case of the R 134a evaporators,
which are denoted by circles, the evaporator 101 with a medium
installation depth and a first fin height has the best
capacity/cost ratio. However, said evaporator has a low output
capacity and is therefore not incorporated in generating the trend
line 115.
[0115] The trend line 115 for the R 134a evaporators and the trend
line 116 for the CO.sub.2 evaporators in each case indicate, as
above, the geometries with which particularly favorable results for
the evaporators are to be expected. Although the evaporator denoted
by the reference symbol 102 with a third fin height performs
considerably less favorably, here, the advantage of the small
installation depth in relation to the evaporators denoted by the
reference symbols 104-106 is to be taken into consideration.
[0116] In the consideration of the CO.sub.2 evaporators which are
denoted by the triangles, the good capacity/cost ratio of the
evaporators denoted by the reference symbols 107 and 108, with the
first fin height or a fin height which is slightly below said first
fin height, is to be maintained, but also that of the evaporator
denoted by the reference symbol 110 with the first fin height.
[0117] The evaporator denoted by the reference symbol 111 with the
third fin height lies, according to expectation, in a slightly less
favorable position on account of the high packing density, which
has an adverse effect on the cost aspect. An evaporator with the
second fin height would logically lie between those with the third
and the first fin heights, and would by all means represent an
alternative of interest.
[0118] Finally, less favorable still is said ratio in the
evaporator denoted by the reference symbol 112 with a large
installation depth with the third fin height, and in the evaporator
denoted by the reference symbol 113 with a small installation
depth.
[0119] In the former, the high costs are prevalent on account of
the small fin height (or the high packing density), while in the
latter, the low capacity at still moderate costs is prevalent. The
evaporator denoted by the reference symbol 114, which corresponds
to the evaporator denoted by the reference symbol 93 in FIG. 11,
again remains unconsidered for the above-mentioned reasons.
[0120] The general result is a lower level of the CO.sub.2
evaporators than the R 134a evaporators. Here, a certain cost
deficit is also to be recognized, which can however be accounted
for by a more stable design for strength and safety reasons
(considerably higher operating pressures when using CO.sub.2 as
refrigerant) and therefore higher weight.
[0121] The illustrations shown in FIGS. 13 to 15 correlate to the
first illustrations shown in FIGS. 10 to 12. However, in the
illustrations shown in FIGS. 13 to 15, the variable "installation
depth" plotted on the ordinate or abscissa has been replaced by the
weighted ratio V' of installation depth and the sum of fin
height+10 mm.
[0122] From the capacity of the individual evaporators over the
weighted ratio V' of installation depth to fin height, as
illustrated in FIG. 13, it can be seen that all evaporators
operated with the same refrigerant (R 134a and CO.sub.2) are now
situated on a constant line substantially independently of their
fin height. This explains the selected weighting of the
installation depth and fin height relative to one another, which
takes the form of the summing of 10 mm to the fin height. Also to
be seen again is the capacity advantage of the CO.sub.2 evaporators
over the R 134a evaporators of the same installation depths. The
individual values are again, as above, measured values or values
determined by simulation which have been confirmed by
measurements.
[0123] The same statements as above also result, in principle, in
the illustration shown in FIG. 14 of the capacity related to the
weight over the installation depth related to the weighted fin
height in relation to the absolute installation depth (cf. FIG.
11). It has moreover been shown that the capacity related to the
weight of the tested R 744 evaporators is a maximum between V'=1.3
and V'=2.8, and appears to drop outside said range. Better values
are shown for evaporators above V'=1.5, and values which are better
still for evaporators above V'=1.85. The evaporators with the
highest capacity related to the weight have a weighted ratio V' of
2.2 and 2.4. The trend line, in contrast, shows a maximum at
approximately V'=2.1.
[0124] The same applies in turn for the illustration, shown in FIG.
15, of the capacity/cost ratio plotted over the related
installation depth in a comparison (cf. FIGS. 12 and 15). Here,
too, the preferences do not change. The capacity, related to the
costs, of the R 744 evaporators exceeds that of the R 134a
evaporators when the weighted ratio V' is less than approximately
2.6.
[0125] It can be seen that, with the method according to the
invention, in which, as a function of the predefined dimensions or
parameters, that is to say the installation depth and the fin
height, different target parameters such as the costs, the power
and the weight can be determined and weighted relative to one
another, in particular by means of different weighting, which
variants ultimately represent the most favorable embodiments. In
this way, it is possible with the method according to the
invention, using different application methods which partially also
incorporate weighting, for the most favorable dimensions for the R
134a and the CO.sub.2 evaporators to be worked out in a
particularly efficient way. It is possible in this way for the
ideal dimensioning for the individual evaporators to be selected
taking into consideration the criteria such as weight, capacity
etc.
[0126] Specially-developed programs are preferably used for the
method, which programs allow the user to specify desired criteria,
to specify the desired target parameters, in order to thereby
satisfy the requirements for example of the air-conditioning of a
motor vehicle. It is necessary in generating a program of said type
to introduce or incorporate the experience gained in each case
through measurement and/or complex thermodynamic
considerations.
[0127] The invention is therefore also aimed at an item of software
which allows the method according to the invention to be carried
out on a computer.
[0128] For the CO.sub.2 evaporators, particularly favorable
installation depths were determined in the range from 20 to 45 mm
with a fin height of 4.0 to 10.0 mm.
[0129] Proven to be particularly advantageous were installation
depths between 35 and 45 mm with fin heights of 5.5 to 10 mm, in
particular for use in large vehicles, and installation depths from
20 to 35 mm with fin heights from 4 to 8.5 mm, in particular for
use in compact and medium-sized vehicles.
* * * * *