U.S. patent application number 15/509427 was filed with the patent office on 2018-08-16 for compact indirect evaporative cooler.
The applicant listed for this patent is FF SEELEY NOMINEES PTY LTD. Invention is credited to Nan Chen, Robert William Gilbert, Shaun Mahoney, David Mark Swindon.
Application Number | 20180231262 15/509427 |
Document ID | / |
Family ID | 55458187 |
Filed Date | 2018-08-16 |
United States Patent
Application |
20180231262 |
Kind Code |
A1 |
Gilbert; Robert William ; et
al. |
August 16, 2018 |
COMPACT INDIRECT EVAPORATIVE COOLER
Abstract
An indirect evaporative cooling system with a core greatly
reduced in size compared to conventional evaporative cooling
systems The system has a heat exchanger core having heat exchange
plates defining a plurality of wet air flow passages and a
plurality of dry air flow passages. At least one fan drives air
through the passages. The dry air passages have a small height and
a short length, configured so that a substantially laminar airflow
having a raised shear rate arises in the dry air passages, and so
that a back pressure across a length of the dry air passages
remains low.
Inventors: |
Gilbert; Robert William;
(Lonsdale, AU) ; Swindon; David Mark; (Lonsdale,
AU) ; Chen; Nan; (Lonsdale, AU) ; Mahoney;
Shaun; (Lonsdale, AU) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
FF SEELEY NOMINEES PTY LTD |
Lonsdale |
|
AU |
|
|
Family ID: |
55458187 |
Appl. No.: |
15/509427 |
Filed: |
September 8, 2015 |
PCT Filed: |
September 8, 2015 |
PCT NO: |
PCT/AU15/50528 |
371 Date: |
March 7, 2017 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
62047160 |
Sep 8, 2014 |
|
|
|
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
Y02B 30/563 20130101;
F28F 3/08 20130101; Y02B 30/54 20130101; F28F 2275/205 20130101;
F24F 1/0007 20130101; F24F 12/006 20130101; F28D 9/00 20130101;
Y02B 30/56 20130101; Y02B 30/545 20130101; F24F 5/0035 20130101;
F28F 13/02 20130101; F28D 5/00 20130101 |
International
Class: |
F24F 5/00 20060101
F24F005/00; F28D 5/00 20060101 F28D005/00; F28D 9/00 20060101
F28D009/00; F28F 3/08 20060101 F28F003/08 |
Claims
1. An indirect evaporative cooler comprising: a heat exchanger core
having heat exchange plates defining alternating wet and dry air
flow passages, wherein the relationship between the passage height,
measured as space between plates defining at least the dry air flow
passages, and the length of the air flow passages falls within an
area on a graph defined by the following points, wherein passage
length is plotted on one axis and passage height, measured as
distance between plates, is plotted on another axis: TABLE-US-00004
Passage Passage Length (mm) Height (mm) 500 1.9-7.5 400 1.7-5.5 300
1.55-4.5 200 1.3-3.5 100 0.9-2.7 75 0.8-2.3 50 0.7-1.7 25
0.6-1.1.
2. The indirect evaporative cooler as claimed in claim 1, wherein
the relationship between the passage height of at least the dry air
flow passages and the length of the passages falls within an area
on a graph defined by the following points, wherein passage length
is plotted on one axis and passage height is plotted on another
axis: TABLE-US-00005 Passage Passage Length (mm) Height (mm) 500
2.8-4.5 400 2.5-3.7 300 2.2-3.3 200 1.8-2.5 100 1.5-2.2 75 1.1-1.7
50 0.9-1.1 25 0.65-0.7.
3. The indirect evaporative cooler as claimed in claim 1, wherein
passage length is from 80 to 200 mm.
4. The indirect evaporative cooler as claimed in claim 1, wherein
the space between plates defining dry passage height is from 0.6 mm
to 2.5 mm.
5. The indirect evaporative cooler as claimed in claim 4, wherein
the space between plates defining dry passage height is from 0.7 mm
to 1.4 mm.
6. The indirect evaporative cooler as claimed in claim 5, wherein
the space between plates defining dry passage height is from 0.8 mm
to 1.2 mm.
7. The indirect evaporative cooler as claimed in claim 1, wherein
the space between plates defining wet passage height is from 0.6 mm
to 2.5 mm.
8. The indirect evaporative cooler as claimed in claim 7, wherein
the space between plates defining wet passage height is from 0.7 mm
to 1.4 mm.
9. The indirect evaporative cooler as claimed in claim 8, wherein
the space between plates defining wet passage height is from 0.8 mm
to 1.2 mm.
10. A method of indirect evaporative cooling, the method comprising
directing a laminar flow of air at a flow rate of 2.5-7.0 m/s
through at least dry air passages of a heat exchanger having heat
exchange plates defining alternating wet and dry air flow passages,
the plates having a separation of from 0.6 mm to 2.0 mm and
defining air passages having a length of from 25 to 300 mm,
dividing the air after passing through the dry passages into first
and second air streams, directing the first air stream into the wet
air flow passages in counter-current flow to the airflow in the dry
passages, and directing the second air stream to a space to be
cooled.
11. The method of indirect evaporative cooling as claimed in claim
10, wherein air is flowed through the dry air flow passages at
3.0-7.0 m/s.
12. The method of indirect evaporative cooling as claimed in claim
10, wherein air is flowed through the wet air flow passages at
2.5-4.0 m/s.
13. An indirect evaporative cooler comprising: a heat exchanger
core having heat exchange plates defining a plurality of wet air
flow passages and a plurality of dry air flow passages; and at
least one fan configured to drive air through the dry air passages,
the dry air passages being configured so that a substantially
laminar airflow having a raised shear rate arises in the dry air
passages, and so that a back pressure across a length of the dry
air passages remains acceptably low.
14. The indirect evaporative cooler as claimed in claim 2, wherein
passage length is from 80 to 200 mm.
15. The indirect evaporative cooler as claimed in claim 14, wherein
the space between plates defining dry passage height is from 0.6 mm
to 2.5 mm.
16. The indirect evaporative cooler as claimed in claim 5, wherein
the space between plates defining wet passage height is from 0.6 mm
to 2.5 mm.
17. The indirect evaporative cooler as claimed in claim 15, wherein
the space between plates defining wet passage height is from 0.6 mm
to 2.5 mm.
18. The indirect evaporative cooler as claimed in claim 17, wherein
the space between plates defining wet passage height is from 0.7 mm
to 1.4 mm.
19. The indirect evaporative cooler as claimed in claim 18, wherein
the space between plates defining wet passage height is from 0.8 mm
to 1.2 mm.
20. The method of indirect evaporative cooling as claimed in claim
11, wherein air is flowed through the wet air flow passages at
2.5-4.0 m/s.
Description
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] This application claims the benefit of U.S. Provisional
Patent Application No. 62/047,160 filed 8 Sep. 2014, which is
incorporated herein by reference.
TECHNICAL FIELD
[0002] The invention relates to an indirect evaporative cooler, and
in particular, an indirect evaporative cooler with a compact core.
The compact core can also be used in a heat recovery heat
exchanger.
BACKGROUND OF THE INVENTION
[0003] At its core, a modern indirect evaporative cooler typically
comprises a stack of thin parallel spaced-apart plates. The plates
define alternating wet and dry air-flow passages. In the wet
passages a "working" airstream passes over wetted surfaces,
accepting and carrying away sensible heat as well as latent heat of
evaporation, leaving evaporatively cooled wet surfaces. In the dry
passages an initially warm airstream is cooled as heat is
transferred by convective transfer from the airstream to the cooled
plate surfaces and by conductive heat transfer through the plates.
The temperature gradient between the airstreams on either side of
the thin plates drives the flow of heat from dry passage side to
wet passage side.
[0004] A two-stage evaporative cooler improves over a single-stage
cooler in that an airstream pre-cooled in a first or "indirect"
cooling stage is used as the working air stream in the wet passages
in a second or "direct" stage. Alternatively, the cooled airstream
is divided into two streams, one of which is used to cool the
living space, the other of which is returned as "return air" to the
wet passages. A pre-cooled working air stream can carry away more
heat from the wetted surfaces, further lowering the temperature of
the evaporatively cooled wet surfaces, thus lowering the
temperature of the airstream being conditioned in the dry
passages.
[0005] In a heat-scavenging heat exchanger, on the other hand, the
thermodynamics are worked in reverse. For example, when heating a
living space in winter, cold outside air is drawn in through dry
air channels. Warm humid air leaving the living space passes
through wet channels in counter-current flow. As the outgoing warm
humid air passes adjacent to the cold incoming air, the water vapor
in the outgoing air condenses, evolving heat of condensation (equal
to the heat of vaporization consumed in cooling), and transferring
sensible heat across the metal plates to warm the incoming air in
the dry passages. As a result, a good portion of the heat leaving
the living space can be recaptured and returned to the living
space.
[0006] Indirect evaporative cooling systems offer many advantages
over mechanically driven refrigerant-compression-type cooing
systems, such as lower electrical consumption, higher reliability,
and freedom from environmentally detrimental refrigerants such as
R-134.
[0007] However, although many improvements have been made over the
years, a major disadvantage in evaporative cooling systems is the
considerable bulkiness of the heat exchanger, and particularly the
core. The cooling capacity of the evaporative cooler is generally
considered to be a function of the total evaporative surface housed
in the system. A commercial evaporative cooler can be several times
the size of a comparable refrigerant-compression based cooling
system.
[0008] Attempts have been made to improve the efficiency of heat
exchangers. As explained in U.S. Pat. No. 5,718,848, the boundary
layer of gas adjacent the plates constitutes an obstruction to
transfer of energy. Gases are notorious insulators against
conduction, and air is very resistant to heat transfer. According
to U.S. Pat. No. 5,718,848, evaporation of water into the working
air stream is improved by using multiple wicks positioned so that
sufficient turbulence is developed to effect periodic restart of
the process of evaporation of moisture from the wicks.
[0009] A more recent proposal for improving heat exchanger
efficiency is described in U.S. Pat. No. 8,636,269. The evaporative
heat exchanger is formed of corrugated sheets of material having
one wettable surface and an opposed dry vapour-resistant surface.
Sheets are stacked with wettable surfaces facing each other to form
wet passages and dry surfaces facing each other to form dry
passages between the sheets. The general direction of air flow is
at an angle to the corrugations. Working air enters the labyrinthic
matrix and encounters numerous intersections of the adjacent
corrugated sheets. At each of these intersections there is intense
interaction between air and water, resulting in rapid evaporation
of water from the wetted surfaces, thereby humidifying the working
air and cooling the water on the wetted surfaces. In the dry
passages on the dry side of the sheets, hot incoming air to be
conditioned loses thermal energy by interaction with the
evaporatively cooled sheet. According to U.S. Pat. No. 8,636,269,
heat exchange between the wet and dry passages and evaporation
within the wet passages can readily take place due to the intensity
of mixing promoted by the corrugated construction.
[0010] The conventional wisdom is thus to design the wet and dry
air passages for increased turbulence in order to promote heat
transfer and evaporative cooling. However, as turbulence increases,
so does resistance to flow, which can be measured as increased
pressure drop over distance. This pressure drop must be overcome,
for example, by provision of additional or higher power air movers,
e.g., fans, which undesirably increases power consumption of the
system. Adding fans also increases construction and operating
costs. It is known that pressure drops can be mitigated by
increasing the height of the airflow passages. However, considering
that air is an insulator, as the height of air passages is
increased, the effectiveness of heat transfer is lowered. Further,
an increase in plate spacing will increase the total height, and
thus bulk, of the heat exchanger core.
[0011] It is disclosed in U.S. Pat. No. 8,636,269 that airflow at a
shallower angle to the direction of the corrugations results in
smoother airflow and lower resistance, but at a penalty of reduced
heat transfer efficiency. The patent teaches that this loss of heat
transfer efficiency can be regained by extending the overall length
of the core. Increasing passage length increases the bulk of the
core. And since pressure drop increases as the length of the air
passages is increased, additional air movers will also be required.
Thus, bulkiness appears to be an inherent characteristic of heat
exchangers.
[0012] US Patent Application No. 2004/0061245 (Maisotsenko et al)
teaches an evaporative air cooling system with an improved heat
exchange surface and an improved system of distributing evaporating
fluids. The cooling system is comprised of plates having a water
conducting layer on one side and a low permeability layer on the
other side. The low permeable layer exhibits low conductivity to
heat except when very thin and is preferably plastic. Since heat
transfer is good perpendicularly through a plastic layer but poor
along the surface of a plastic layer, the "result of this
differential heat transferability is that heat will transfer from
one side of the plate to the other along the interface of the
plastic while at the same time heat will not readily transfer along
the surface. The result is that discrete temperatures and a
temperature differential can occur at different points in the plate
and it will not be averaged due to the heat transfer by the plate."
(paras. [0026]-[0027]). Also mentioned is that when plate spacing
is within certain bracketed values (be 1.57 mm to 1.83 mm, 2.17 mm
to 2.33 mm, 2.16 to 2.87 mm, or 3.13 mm to 3.39 mm), the pressure
drop across plates is reduced from 1% to 15%. Further, deposit
build-up is reduced along the plate surfaces due to the transverse
quarter wave, i.e., turbulence, increasing the dynamic energy of
the flow in the direction of the flow at the boundary layer. Thus,
Maisotsenko et at adhere to the accepted wisdom that dynamic
energy, i.e., turbulence, is needed in the air flow to break up the
boundary layer for efficient heat transfer.
[0013] There remains a need for a more compact one or two stage
evaporative cooler, as well as a compact heat exchanger for
recapture of heat.
[0014] Any discussion of documents, acts, materials, devices,
articles or the like which has been included in the present
specification is solely for the purpose of providing a context for
the present invention. It is not to be taken as an admission that
any or all of these matters form part of the prior art base or were
common general knowledge in the field relevant to the present
invention as it existed before the priority date of each claim of
this application.
[0015] Throughout this specification the word "comprise", or
variations such as "comprises" or "comprising", will be understood
to imply the inclusion of a stated element, integer or step, or
group of elements, integers or steps, but not the exclusion of any
other element, integer or step, or group of elements, integers or
steps.
[0016] In this specification, a statement that an element may be
"at least one of" a list of options is to be understood that the
element may be any one of the listed options, or may be any
combination of two or more of the listed options.
SUMMARY OF THE INVENTION
[0017] According to a first aspect the present invention provides
an indirect evaporative cooler comprising:
[0018] a heat exchanger core having heat exchange plates defining
alternating wet and dry air flow passages, wherein the relationship
between the passage height, measured as space between plates
defining at least the dry air flow passages, and the length of the
air flow passages falls within an area on a graph defined by the
following points, wherein passage length is plotted on one axis and
passage height, measured as distance between plates, is plotted on
another axis:
TABLE-US-00001 Passage Passage Length (mm) Height (mm) 400 1.7-5.5
300 1.55-4.5 200 1.3-3.5 100 0.9-2.7 75 0.8-2.3 50 0.7-1.7 25
0.6-1.1
[0019] According to a second aspect the present invention provides
a method of indirect evaporative cooling, the method comprising:
[0020] directing a laminar flow of air at a flow rate of 2.5-7.0
m/s through at least dry air passages of a heat exchanger having
heat exchange plates defining alternating wet and dry air flow
passages, the plates having a separation of from 0.6 mm to 2.0 mm
and defining air passages having a length of from 25 to 300 mm,
[0021] dividing the air after passing through the dry passages into
first and second air streams, [0022] directing the first air stream
into the wet air flow passages in counter-current flow to the
airflow in the dry passages, and [0023] directing the second air
stream to a space to be cooled.
[0024] According to a third aspect the present invention provides
an indirect evaporative cooler comprising: [0025] a heat exchanger
core having heat exchange plates defining a plurality of wet air
flow passages and a plurality of dry air flow passages; and [0026]
a fan configured to drive air through the dry air passages, [0027]
the dry air passages being configured so that a substantially
laminar airflow having a raised shear rate arises in the dry air
passages, and so that a back pressure across a length of the dry
air passages remains acceptably low.
[0028] In some embodiments of the invention, the relationship
between the passage height of at least the dry air flow passages
and the length of the passages falls within an area on a graph
defined by the following points, wherein passage length is plotted
on one axis and passage height is plotted on another axis:
TABLE-US-00002 Passage Passage Length (mm) Height (mm) 500 2.8-4.5
400 2.5-3.7 300 2.2-3.3 200 1.8-2.5 100 1.5-2.2 75 1.1-1.7 50
0.9-1.1 25 0.65-0.7
[0029] In some embodiments of the invention the passage length is
in the range from 80 to 200 mm.
[0030] In some embodiments of the invention the space between
plates defining dry passage height is from 0.6 mm to 2.5 mm, more
preferably from 0.7 mm to 1.4 mm, and more preferably from 0.8 mm
to 1.2 mm.
[0031] In some embodiments of the invention, the space between
plates defining wet passage height is from 0.6 mm to 2.5 mm, more
preferably from 0.7 mm to 1.4 mm, more preferably from 0.8 mm to
1.2 mm.
[0032] In some embodiments of the invention, air is flowed through
the dry air flow passages at 3.0-7.0 m/s, more preferably at
2.5-4.0 m/s.
[0033] In some embodiments of the invention, in use the back
pressure along the length of the dry air passages is less than 100
pascal (Pa), preferably less than 60 Pa, more preferably less than
50 Pa, most preferably less than 45 Pa.
[0034] Desiring a more compact heat exchanger core, and realizing
that the "conventional wisdoms" of core design inevitably produce a
large size indirect evaporative cooler core (i.e., using turbulence
to break down the insulating boundary layer and thus promote heat
transfer; spacing plates far enough apart to prevent constriction
of flow, etc.), the present inventors developed a design for a new
type of core with surprising compactness and efficiency.
[0035] The present invention recognises that when the spacing
between the plates is greatly reduced and, at the same time, care
is taken to ensure a substantially laminar flow rather than
turbulent flow, good evaporative cooling and good heat exchange
occurs, allowing the length of the passages to be reduced. And, as
the passage length is reduced, the pressure drop along this shorter
passage length remains acceptable.
[0036] That is, by pushing air at high shear rates through passages
having narrower spacing between plates--and shorter passages--the
evaporative cooling system becomes not only compact in size, it
also becomes efficient.
[0037] The advantages of the inventive design include: [0038] good
heat transfer since thinner boundary layers in the narrower
passages minimize the insulating effect of air, [0039] more plate
area for a given stack height since plate spacing is reduced,
[0040] shorter passage lengths--as little as only 20% of the
conventional passage length--a major advantage in an installed
system, and [0041] no flow impedance attributable to
turbulence-induced flow resistance, since flow is laminar.
[0042] The plates may be formed of any suitable material or
materials, including metal and plastic, for example, PVC. The
plates preferably have an absorbent flock on one side. The plates
are preferably formed with long arched sections between standoff
spacers (see FIG. 6), with a radius of e.g. 10 mm, whereby static
air in corners is minimized.
[0043] The invention has applications not only for indirect
evaporative coolers, but also to air-to-air heat exchangers (heat
scavengers, heat recovery systems) and/or water-to-air heat
exchangers.
BRIEF DESCRIPTION OF THE DRAWINGS
[0044] An example of the invention will now be described with
reference to the accompanying drawings, in which like reference
numbers indicate similar parts, and in which:
[0045] FIG. 1 shows a simplified core comprised of a stack of
plates with alternating wet and dry passages;
[0046] FIG. 2 shows the CFD model, with variables being plate
spacing or air passage height (H) and passage length (L);
[0047] FIG. 3 shows the results of CFD in the range 0.6 to 1.5 mm
passage height and 25 mm to 100 mm passage length;
[0048] FIG. 4 shows the results of CFD in the range 0.6 to 5 mm
passage height and 25 to 500 mm passage length;
[0049] FIG. 5 is a graph showing inventive and preferred passage
heights and lengths,
[0050] FIG. 6 is a close-up of a practical embodiment of a core
structure showing long arched sections between standoff
spacers;
[0051] FIG. 7 schematically represents an evaporative cooler;
and
[0052] FIG. 8 shows the design of an evaporative cooler according
to one embodiment of the invention.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0053] FIG. 1 shows a known airflow configuration for a single core
evaporative cooler system comprising parallel heat exchange plates
1. Each plate has a wettable surface on one side and a dry
vapour-resistant surface on the opposite side. Plates are stacked
with wettable surfaces facing each other to form wet air passages 2
and dry surfaces facing each other to form dry air passages 3
between the plates. Incoming source air 4 to be cooled is directed
through the dry passages 3 of heat exchanger 10. Upon exiting the
dry passages, the dry cooled air stream is divided approximately
evenly into a "supply air" stream 5 for cooling a living space and
"return" or "working" air stream 6 which is directed into the wet
passages 2. The wet passages 2 have a hydrophilic surface provided
with a wicking material 7 which is capable of being kept
continuously wet by being intermittently or continuously charged
with water. Water evaporates from the wicking surfaces and is
carried away by the cool dry air (which has been pre-cooled by the
initial passage through the cool dry channel), extracting latent
heat, cooling the plates. After passing through the wet passage 2,
the now warmed moisture-laden return air stream 8 is vented to the
atmosphere. In turn, since the pre-cooled return air is more
effective in lowering the temperature of the hydrophilic surface 7,
the incoming air stream 4 in the dry passages 3 is more effectively
cooled by contact with the dry-side surfaces of the plates 1. The
temperature gradient between the air streams on either side of the
thin thermally conductive plate 1 drives the heat flow from dry
passage side to wet passage side.
[0054] The spacing between the plates 1 is approximately the same
whether the plates 1 are defining a wet passage 2 or a dry passage
3. In the illustrated example, the wicking material extends about
one quarter of the way into the wet passages 2, leaving about 50%
of the passage unobstructed. Since the volume of return air is only
about 50% of the incoming air, it follows that the effective height
of the wet air passage 2 for the return air need only be about 50%
of that of the height of the dry air passage 3 for the incoming
air.
[0055] The incoming air 4 may be either fresh air from the
environment outside the living space, or may be recirculation air
drawn from inside the living space.
[0056] The present invention is based on the recognition that
practical evaporative cooling can be accomplished in a smaller
system when three conditions are met: [0057] passage height is
reduced, [0058] passage length is reduced, and [0059] air flow is
laminar and at raised shear rates.
[0060] While not wanting to be bound by any particular theory of
the invention, it is believed that when the plate spacing defining
the dry or wet passages is reduced to approximately 0.6-2.5 mm,
preferably 0.6-2.0 mm, more preferably 0.7-1.4 mm, most preferably
approximately 0.8-1.2 mm, air can be moved in the passage with
laminar flow and high shear rate. With smaller spacing according to
the present invention the speed profile is steep, air flow remains
laminar, the boundary layer becomes thinner, and hence heat
transfer is dramatically increased. It is known that the longer the
passage, the greater the pressure drop between inlet and outlet.
Since a smaller passage length is required in the present invention
to accomplish the system-effective heat transfer, pressure drop is
kept to a minimum. Compared to the prior art heat exchangers
relying on turbulent air flow to break up the boundary layer, with
turbulent air taking energy and increasing back pressure, the
inventive design actually provides efficiencies in heat transfer as
well as air flow in both the wet and the dry passages.
[0061] In the case of plate spacings greater than 2.5 mm,
preferably 2.0 mm, even if flow is kept laminar, the efficiency of
conductive and radiative heat transfer is lost. With spacings
narrower than 0.6 mm air flow becomes constricted, pressure drop
becomes high, and the benefits of the invention are lost.
[0062] By designing a heat exchanger on the basis of this new
insight, it becomes possible to reduce the length of a heat
exchanger passages to about 20% of that of a conventional 400 to
800 mm heat exchanger, i.e., to about 80 to 200 mm. Thus, in
accordance with the invention, it has become possible for the first
time to provide a truly compact indirect evaporative cooler.
[0063] To graphically illustrate the relationship between various
passage heights (H) and passage lengths (L), CFD was carried out
using the model illustrated in FIG. 2. Inlet air was set at
35.degree. C. in all cases. Air was flowed at 3 m/s between two dry
plates gradually cooled to 15.degree. C., thus the target
temperature for exit air is 15.degree. C.
[0064] The results are graphically represented in FIG. 3 (showing
the results for 0.6 mm to 1.5 mm plate spacing vs. 25 to 100 mm
passage length) and FIG. 4 (showing the results for 0.6 mm to 5.0
mm plate spacing vs. 25 to 500 mm passage length).
[0065] The criteria for selecting feasible combinations of passage
height and passage length are pressure drop and temperature drop,
i.e., the air passage must exhibit (a) "manageable" pressure drop
and (b) sufficient cooling. If pressure drop is too large,
additional or more powerful fans may be required to move air,
requiring more energy, reducing system efficiency. Thus, pressure
drop should be less than 100 pascal (Pa), preferably less than 60
Pa, more preferably less than 50 Pa, most preferably less than 45
Pa. Regarding cooling, air temperature at the passage outlet should
be within 4.degree. C. of target, preferably within 1.degree. C. of
target, most preferably within 0.5.degree. C. of target.
[0066] In general, it is readily apparent that when passage height
is small, pressure drop increases rapidly as passage length is
increased. In FIG. 3, the pressure drop (Pd) varies from an
unacceptable 200 Pa at 0.6 mm passage height and 100 mm passage
length to a negligible 10 Pa at 1.5 mm passage height and 25 cm
passage length.
[0067] Considering in greater detail the specific case of the 25 mm
passage length, it can be seen in FIG. 3 that when the passage
height is 0.6 mm and source air introduced into the dry passage
inlet at 35.degree. C., the air is about 15.3.degree. C. at the
passage outlet, thus, heat transfer, measured as outlet
temperature, is quite good. However, as the passage height is
increased, heat transfer efficiency is reduced and the temperature
at the outlet rapidly increases. At a passage height of 0.8 mm the
outlet temperature is 16.5.degree. C.; at a passage height of 1.0
mm the outlet temperature is 18.3.degree. C.; at a passage height
of 1.2 mm the outlet temperature is 20.degree. C.; at a passage
height of 1.5 mm the outlet temperature is 22.5.degree. C.; at a
passage height of 2.0 mm (FIG. 4) the outlet temperature is
25.degree. C.; and at a passage height of 2.5 mm the outlet
temperature is 26.8.degree. C. Evaluating these results for
acceptable heat transfer efficiency, according to which air
temperature at the passage outlet should be within 4.degree. C. of
target, preferably within 1.degree. C. of target, most preferably
within 0.5.degree. C. of target, it follows that when the passage
length is 25 mm, the passage height should be no more than 1.1 mm,
preferably no more than 0.7 mm, most preferably not more than 0.65
mm in height.
[0068] Evaluating the same 25 mm passage length at various heights
on the basis of pressure drop, it can be seen from FIG. 3 that when
the passage is 0.6 mm, the pressure drop is an acceptable 50 Pa,
and as passage height increases, pressure drop decreases further.
Thus, for a 25 mm passage length, pressure drop is acceptable to
excellent in all cases.
[0069] The remaining passage lengths and heights are evaluated for
heat transfer efficiency and pressure drop, and the useful and
preferred values are collected in Table 1.
TABLE-US-00003 TABLE 1 Passage Heat Transfer Pd Length Preferred H
Acceptable H Preferred H Acceptable H 500 0.6-4.5 0.6-7.5
.gtoreq.2.8 .gtoreq.1.9 400 0.6-3.7 0.6-5.5 .gtoreq.2.5 .gtoreq.1.7
300 0.6-3.3 0.6-4.5 .gtoreq.2.2 .gtoreq.1.55 200 0.6-2.5 0.6-3.5
.gtoreq.1.8 .gtoreq.1.3 100 0.6-2.2 0.6-2.7 .gtoreq.1.5 .gtoreq.0.9
75 0.6-1.7 0.6-2.3 .gtoreq.1.1 .gtoreq.0.8 50 0.6-1.1 0.6-1.7
.gtoreq.0.9 .gtoreq.0.7 25 0.6-0.7 0.6-1.1 .gtoreq.0.65
.gtoreq.0.5
[0070] The data in Table 1 is plotted in FIG. 5, wherein dots
represent the bounds of heat transfer, Xs represent the bounds of
pressure drop, dash lines represent acceptable values, and solid
lines represent preferred values. The scale of the graph changes
above 2.0 mm passage height. For heat transfer, all passage heights
to the left of the lines are included. For pressure drop, all
passage heights to the right of the lines are included. Since it is
necessary for the heat exchanger core to exhibit both good heat
transfer and acceptable pressure drop, it follows that everything
between the left and right dash lines is considered within the
bounds of the present invention. Everything between the solid lines
is particularly preferred core design parameters within the scope
of the present invention.
[0071] While the CFD has been carried out only for dry plate
spacing and passage lengths, the addition of wet side to the
simulation would not have been expected to alter the results. While
air flowed was a constant 3 m/s in the CFD model, and while this
flow rate is believed to most closely approximate practical
conditions, the invention is obviously not limited to this flow
rate.
[0072] The heat exchanger of the present invention, with
micro-passages as defined herein, can be adapted to virtually any
design, and be used in virtually any system, and offer the benefits
of reduced bulk. General systems and materials will be described in
the following, without the invention being limited to the exemplary
materials.
[0073] The air streams may be in counter-current flowing in
opposite directions, or one stream may be perpendicular to the
other stream. The compact heat exchanger of the present invention
can be used not only for evaporative cooling, but also for heat
recovery as discussed in greater detail below.
[0074] The major difference between the present invention and the
prior art is in the reliance on laminar air flow with high shear in
narrow air passages. Thus, prior art structures and methods
intended to introduce turbulence into the air flow are to be
avoided. For example, in a preferred embodiment of the invention,
the heat exchanger plates are formed of a hydrophobic material. One
side of the plate remains hydrophobic, while the other side of the
plate is rendered hydrophilic by suitable treatment discussed
below, and wherein the hydrophilic side is surface treated, either
before or after being rendered hydrophilic, in order to provide
sufficient microcapillary or wicking surface for supply of water
for evaporation along the length of the passage. Such a "wet"
surface provides very little interference to laminar air flow.
Device for Evaporative Cooling
[0075] FIG. 7 shows a sectional view of a practical arrangement for
a device exploiting the advantages of indirect evaporative cooling.
Air enters from external ambient through fan 9, which supplies high
pressure air to chamber 10. Heat exchanger 11, also referred to as
a core, is manifolded such that high pressure air from chamber 10
can only flow through the dry channels of the heat exchanger, and
air which flows through the dry channels must flow all the way
through the dry channels, emerging into chamber 12. A portion of
the air emerging from the dry channels into chamber 12 is required
to be turned around to flow back to through the wet channels spaced
between the dry channels of the heat exchanger 11. This requires a
pressure in chamber 10 to be sufficient to overcome the flow
resistance of the wet channels to leave exhaust 13 at the required
flow rate. This pressure may be achieved by providing a baffle or
restriction in chamber 12 via air duct 14, the pressure
differential across the baffle at the required flow rate resulting
in a static pressure in chamber 10.
[0076] Fan 9 is required to pressurize air to overcome the pressure
loss associated with passing all of the air supplied through the
dry channels, plus the static pressure in chamber 12. The static
pressure in chamber 12 is sufficient to overcome the flow
resistance of the proportion of air flowing through the wet
channels to exhaust 13. The static pressure in chamber 12 is
regulated by adjusting baffle thereby producing a static pressure
differential across the baffle. The air flow through the baffle at
such a differential pressure represents a loss of power equal to
the product of the air flow and pressure differential. This loss is
an additional power load on fan 9 which provides no additional
cooling or otherwise useful energy to the air flow. Although fan 9
is shown schematically as an axial flow fan, in practice a
centrifugal or combined flow fan is generally used due to the high
pressures required.
[0077] Fan 15 in the exhaust duct 13 of the heat exchanger may
optionally be provided to produce a negative pressure relative to
the pressure in chamber 12 sufficient to produce the required air
flow through the wet passages of the heat exchanger. Thus the
static pressure immediately before fan 15 will be the sum of the
static pressure in chamber 12 and the pressure differential
required for the air flow through the wet passages of the heat
exchanger 11. The operation of the fans 9 and 15 can be controlled
through electronic speed controllers or other means to produce a
desired ratio of air flow between the supply air and exhaust air.
Furthermore, the magnitude and/or ratio of these air flows can be
readily adjusted by varying the speeds of the two fans 9, 15
thereby enabling optimization of the performance of the indirect
evaporative cooler. This allows the indirect cooler to operate
under a wide range of conditions through direct control of the fans
without the need to intervene and adjust mechanical baffles as in
prior art designs, and also allows for automatic control of the
operation of the indirect cooler, for example, under the control of
a programmable electronic controller.
[0078] In a typical indirect evaporative cooler, with a supply to
exhaust ratio of 1:1, the fan is required to deliver air at around
600 Pa. If the supply air required is, say, n units, the power
required will be 600.times.2n=1,200n power units. This typically
produces a static pressure of around 150 Pa in chamber 48 and thus
a pressure differential of 150 Pa across the wet passages to
exhaust. The pressure differential across the dry passages of the
heat exchanger is 600 Pa-150 Pa=450 Pa.
The Plates
[0079] The plates 16 separate dry product air passages and wet
working air passages. Each plate is made of a thin material to
allow easy heat transfer across this plate and thus to readily
allow heat to transfer from the dry product passage to the wet
working passage. The plate is preferably formed of a plastic film
such as a thin-wall dense film or sheet of polyethylene,
polypropylene, polystyrene, polyvinyl chloride, polyethylene
terephthalate, or the like material having good vapor barrier
properties. Although a low permeability material such as plastic
does not readily transfer heat, heat transfer perpendicularly
through the plastic layer will be good since the plates according
to the present invention are very thin.
[0080] Air streams generally flow between two plates rather than
across one plate. If two such plates are aligned with dry sides
facing, then the air streams flow between the two plates on the dry
sides; if the wet sides are facing, the air streams flow between
the plates on the wet sides. In embodiments having more than two
plates, air streams may first flow between the dry sides of two
plates, then flow through one or both plates and enter wet
passages, in which they will flow across one of the two previous
plates (on the reverse side) and the wet side of a third plate.
[0081] Plates of various materials and configurations are described
in U.S. Pat. Nos. 6,581,402; 6,705,096; 7,228,699; and 8,468,846,
the disclosures of which being incorporated herein by
reference.
The Wicking Material/Surface Treatment
[0082] While the plate 16 may be formed of a hydrophobic polymer
such as extruded polypropylene, one surface of the polymer may be
rendered substantially hydrophilic by subjecting it to Corona
treatment, plasma discharge, plasma jet, flame treatment, acid
etching and nano-surfacing or nano-coating.
[0083] The hydrophilic side of the plate may be provided with an
additional liquid retaining layer formed from a fibrous non-woven
material 17. Although reference may be made to a liquid retaining
surface, it is clearly understood that the surface is in fact a
liquid retaining and releasing surface. The evaporation rate off of
a hydrophobic woven or spun bond material where water has been
impregnated in between the fibers is higher than from a hydrophilic
material where water has been absorbed into the material and
between the fibers. This means that a much smaller temperature
difference across the plate is required to achieve the same
evaporation rate, which therefore increases the heat transfer rate.
See for example WO 2010/011687 (Gillan) teaching a hydrophobic
fiber sheet formed to wick evaporative fluid.
[0084] In the present invention the wicking material or wicking
layer/treatment is relatively thin, and thus contains only a small
amount of water. Accordingly, the system is efficient in that air
rather than water is cooled.
[0085] The wicking material may be hydrophilic or hydrophobic, and
suitable materials include cellulose, fiberglass, organic fibers,
organic-based fibers, porous plastics, carbon-based fibers,
polyesters, polypropylene, silicon-based fibers and combinations of
these substances. The wick layer material may be in a number of
forms: films, weaves, braids, fibers, beds of particles such as
beads and combinations thereof.
[0086] A substantially compliant nonwoven wicking material is
disposed on and fixedly attached at a number of locations to the
hydrophilic surface of the first polymer substrate. Similarly,
substantially compliant nonwoven material is also disposed on and
fixedly attached at a number of locations to the hydrophilic
surface of the second polymer substrate. The substantially
compliant nonwoven material can be a spunbonded material, a melt
blown material, hydroentangled (spunlaced) material or made through
any other processes such as co-forming, airlaying, wetlaying,
carding webs, thermal bonding, needle punching, chemically bonding
or combinations thereof. Embodiments of spunbonded material include
polyolefin, Polyethylene terephthalate (PET) and nylon. Embodiments
of melt blown material include polyolefin, Polyethylene
terephthalate (PET) and nylon. Embodiments of hydroentangled
material include cotton, rayon or viscose staple fiber, lyocell
staple fiber, polyolefin staple fiber, polyester staple fiber and
nylon staple fiber.
[0087] Nonwoven webs can be formed from fibers and filaments based
on hydrophobic or hydrophilic polymers. Representative, but not
complete, examples of polymers that are hydrophobic for making
nonwoven webs are polyolefins and polyethylene terephthalate.
Representative, but not complete, examples of hydrophilic polymers
for making nonwoven webs include cellulosic materials like cotton,
rayon or viscose etc. The application of the fact that under
suitable conditions of porosity, fiber/filament diameter, density
(GSM) etc, significant capillary action and wicking of water can
occur in a web has been innovatively applied in the invention. The
invention innovatively utilizes the porosity of certain porous
nonwoven webs that can often be sufficient to enable the easy
transport of water and other fluids because of wicking caused by
capillary action.
[0088] The prior art teaches that hydrophilic materials can better
hold water. However, in relation to the cooling apparatus, the
application of this quality has a disadvantage that in the case of
a nonwoven web made from hydrophilic polymers, some of the water
will swell the fibers and the rest will go around and over the
fibers. This would lose the rigidity in the heat exchanger pads.
Further, in relation to the cooling apparatus, hydrophilic
non-woven would swell, while one of our objectives is to retain the
thinnest film of the water to facilitate better heat transfer and
evaporation. Porous low density nonwoven webs made from hydrophobic
fibers or filaments can transfer water through wicking action.
Water can flow along, around and over but not through the
hydrophobic polymer fibers. The porosity and associated wicking
action by a porous nonwoven web can render the nonwoven web
effectively hydrophilic in terms of its capability to be wet and
easily spread water even if the fibers or filaments constituting
the nonwoven web are made from hydrophobic polymers. The invention
thus innovatively employs the materials known to be hydrophobic for
the retention of water as required. The invention overcomes the
problem in maintaining rigidity of heat exchanger pads due to the
use of hydrophilic material, as evidenced by relevant prior art, by
employing hydrophobic material.
[0089] Examples of fibers that are hydrophobic are polyolefins and
polyethylene terephthalate. Porous low density nonwoven webs made
from these hydrophobic fibers or filaments can be hydrophilic
through wicking action.
Spacers
[0090] The plates may be separated by any conceivable spacing means
18. The plates can be deformed with a punch or a roller to
introduce raised points or walls. Adhesive or plastic can be
printed to provide the desired structures at the desired
height.
[0091] The spacers may be spaced apart a distance of 30-50 mm.
Keeping in mind the thinness and minimal structural integrity of
the plates, the spacers should be spaced as far apart as possible
while still ensuring proper spacing of the sheets.
Heat Scavenging Heat Exchanger
[0092] The device may be operated in winter months to scavenge heat
from exhaust gases of a space and thus pre-heat fresh air, while
simultaneously humidifying the fresh air.
[0093] Such an indirect evaporative cooler would have cycle
selection means, so that during summer months, it may be used to
provide cooled, non-humidified air, and during winter months, it
may be used to scavenge heat from gases exiting a space while
simultaneously humidifying the space.
[0094] In winter months, it is advantageous to exchange heat
between exhaust air leaving a warmed space and cold fresh air being
brought in from the atmosphere, i.e. the outdoor air or other
source of environmental air. This reduces the heat required to warm
the fresh air. The present invention also allows the addition of
humidity to the fresh air, thus addressing another winter problem:
cold outside air that has condensed moisture out and therefore has
a low absolute humidity or extremely dry air that results in dry
inside air as the moisture on the inside reduces with fresh air
changes with the outside. The "cycle selection" as to which stream
of air is exhausted to the atmosphere, and which goes to the space
to be conditioned, is a feature of embodiments having this
arrangement.
Modular Design/Scaling Up
[0095] Heat exchangers can be manufactured in standard sizes as
modules, and a number of these modules can be assembled to provide
an air handler with a capacity required for a given living or
working space.
System Operation
[0096] During operation, an evaporative liquid is distributed to
the substantially compliant nonwoven material of each of the
hydrophilic surfaces of the unit and a fluid flows within a space
separating each of the units with heat being exchanged between the
evaporative liquid and the fluid, and another fluid flowing through
at least some passages from the number of passages. The other fluid
exchanges heat with the hydrophilic surface. (Heat is transferred
from the nonwoven material to the hydrophilic surface/wet side and
through the plate to the hydrophobic surface/dry side). In one
instance, the fluids are air, and the heat exchanger has moist air
from the evaporated cooling of the substantially compliant nonwoven
material having the evaporative liquid distributed over and dry air
flowing through the passages and being cooled. In that instance to
a heat exchanger is referred to as a Dry Air, Moist Air (DAMA) heat
exchanger.
[0097] As disclosed hereinabove, in the indirect evaporative
cooling heat exchanger of this invention described herein above,
during operation, an evaporative liquid is distributed to the
substantially compliant nonwoven material of each of the
hydrophilic surfaces of a unit and a secondary fluid flows within a
space separating each of the units with heat being exchanged
between the evaporative liquid and the fluid. A primary fluid flows
through at least some passages from a number of passages. The
primary fluid exchanges heat with the hydrophilic surface. (Heat is
transferred from the nonwoven material to the hydrophilic surface
and through the substrate to the hydrophobic surface). Another
evaporative liquid is distributed to the direct evaporative
component. The direct evaporative component is positioned
downstream from the indirect evaporative cooling component and
receives at least a portion of the primary fluid from the indirect
evaporative cooling component. In the direct adiabatic evaporative
component, the primary fluid is cooled by addition of the other
evaporative liquid. During the cooling process in the adiabatic
direct evaporative component, there is no change in the total
energy or enthalpy, but, a portion of sensible heat of the primary
fluid is converted into latent heat.
[0098] In one instance, the two-stage evaporative cooling apparatus
of this invention includes a primary fluid supply component,
located upstream from the indirect evaporative cooling component of
this invention and supplying the primary fluid to the indirect
evaporative cooling component of this invention. During operation
the primary fluid supply component draws ambient fluid through a
filtering component and supplies filtered ambient fluid as the
primary fluid. The filtering component can include one of a variety
of filters (including but not limited to conventional filters,
carbon filters, electrostatic filters, etc.). A first evaporative
liquid supply system supplies the evaporative liquid to the
indirect evaporative cooling component of this invention. A second
evaporative liquid supply system supplies the other evaporative
liquid to the direct adiabatic evaporative cooling component. In
one instance, a liquid holding component (such as a tank) provides
a supply of the evaporative liquid and the other evaporative
liquid. The first evaporative liquid supply system (a pump in one
instance) and the second evaporative liquid supply system (another
pump in one instance) are disposed inside the liquid holding
component. The first evaporative liquid supply system, the second
evaporative liquid supply system and a liquid holding component are
comprised of aseptic material. A liquid disinfection system can be
disposed to receive the evaporative liquid and the other
evaporative liquid and render both of them disinfected. In one
embodiment the liquid disinfection system includes a system
utilizing ultraviolet (UV) radiation in order to disinfect the
evaporative liquid on the other evaporative liquid. It should be
noted that other liquid disinfecting systems, such as, but not
limited to, system utilizing ozone and other liquid disinfecting
systems are within the scope of this invention.
[0099] Note that the indirect evaporative cooler of the invention
may also accomplish both direct and indirect evaporative cooling of
the product air stream. A portion of the dry sides may be wetted,
in a manner similar to the wick materials used on wet sides or in a
different manner, so as to cause further cooling of the product air
stream. This wet portion of the dry sides may advantageously be
placed downstream of the dry portion of the dry sides, so that
prior to being humidified in the direct evaporation cooling
process, the sensible temperature of the product air stream is
reduced as much as possible. One particular advantage of this
ordering is that below approximately 65 degrees F., modest
increases in humidity cause disproportionate reduction in air
temperatures, in accordance with standard psychometric charts. In
another embodiment of the invention, this wetted portion of the dry
sides constitutes the final 1 to 25 percent of the surface area of
the dry passages.
[0100] Some embodiments of the invention may utilise 3D printing
for construction of the device. Accordingly, in some embodiments
the present invention may reside in a digital blueprint comprising
a digital file in a format configured for use with rapid
prototyping and computer aided design (CAD) and/or manufacturing,
such as being in the STL (stereolithography) file format. Such
digital blueprint files, whether produced by performing a three
dimensional scan of an embodiment of the invention, or produced by
a CAD development software tool, or the like, are within the scope
of the present invention.
[0101] It will be appreciated by persons skilled in the art that
numerous variations and/or modifications may be made to the
invention as shown in the specific embodiments without departing
from the spirit or scope of the invention as broadly described. The
present embodiments are, therefore, to be considered in all
respects as illustrative and not restrictive.
* * * * *