Note: Descriptions are shown in the official language in which they were submitted.
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IMPROVED EVAPORATIVE CONDENSER
TECHNICAL FIELD
An improved evaporative condenser and evaporative condensation process are
disclosed for use in refrigeration and air-conditioning systems. The condenser
and
process can be employed with both chemical refrigerants (e.g.
hydrofluorocarbons) and
natural refrigerants (e.g. hydrocarbons (such as propane & isobutane), CO2,
ammonia,
etc).
BACKGROUND ART
Existing evaporative condensers are used to reject heat in a variety of
refrigeration and air-conditioning systems through the condensing of a
refrigerant.
More specifically, evaporative condensers comprise one or more wetted (e.g.
sprayed)
condensing coils for condensing the refrigerant by the passage thereover of an
airstream, and into which a portion of the water is evaporated, thereby
removing heat
from the refrigerant in the condensing coils and causing the refrigerant to
condense
therein. Evaporative condensers also comprise drift eliminators (or, more
simply,
eliminators, "drift" being water that would otherwise pass to atmosphere).
Drift
eliminators remove free water that passes with the airstream as it flows
through the
condensing coils and water spray, prior to releasing that airstream to
atmosphere.
In existing evaporative condensers the plan area of the condensing coils is
2 0 matched to the plan area of the drift eliminators to ensure constant
air flow rate and
airstream velocity through the evaporative condenser.
In existing evaporative condensers, heat exchange efficiency is limited by the
velocity of air that flows over the condensing coils. The velocity of air is
in turn limited
by the ability of the drift eliminators to remove free water from the air
passing
therethrough. In existing evaporative condensers, such removed water is
recycled for
reuse in wetting the condensing coils. However, any water that passes with the
air
flowing through the drift eliminators to atmosphere may contain bacteria, such
as
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legionella, hence the requirement to remove as much free water from the
airstream as
possible.
For example, in many existing evaporative condensers, it is known to specify a
maximum air velocity through the drift eliminators as high as 3.5 to 4 m/s to
ensure
sufficient water removal, however, it is surmised that, with such a high
maximum air
velocity, there is still a significant risk of bacteria (such as legionella)
passing with non-
eliminated free water through the drift eliminators. A safer maximum air
velocity
through the drift eliminators of 3.5 m/s is proposed. However, this will in
turn set a
limit to the velocity of air that can flow over the condensing coils.
1 0 The above
references to the background art do not necessarily constitute an
admission that the art forms part of the common general knowledge of a person
of
ordinary skill in the art. The above references are also not intended to limit
the
application of the condenser and process as disclosed herein.
SUMMARY OF THE DISCLOSURE
Disclosed herein is an evaporative condenser for use in a refrigeration or air-
conditioning system. The evaporative condenser as disclosed herein can
condense
chemical refrigerants (e.g. hydrofluorocarbons, hydrochlorofluorocarbons,
perfluorocarbons, hydrofluoroolefins, etc) and natural refrigerants (e.g.
hydrocarbons
2 0 such as propane & isobutane, CO2, ammonia, etc).
The evaporative condenser as disclosed herein comprises one or more
condensing coils for condensing therewithin the refrigerant of the system. The
one or
more condensing coils can be arranged in a condensing coil zone of the
evaporative
condenser. The condensing coil zone may comprise an air plenum having a
constant
cross-sectional area.
The evaporative condenser as disclosed herein also comprises a mechanism for
wetting the one or more condensing coils (e.g. by spraying them with water).
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The evaporative condenser as disclosed herein further comprises drift
eliminators arranged to remove free water from an airstream that has flowed
past the
one or more condensing coils and wetting mechanism.
In accordance with the present disclosure, the evaporative condenser as
disclosed herein comprises a divergent zone that diverges from the condensing
coil
zone towards the drift eliminators. The configuration of the divergent zone is
such that,
once the airstream has flowed past the one or more condensing coils, it flows
into and
through the divergent zone to the drift eliminators. For example, the
divergent zone
may comprise an air plenum having a progressively increasing cross-sectional
area.
The divergent zone is able to cause the airstream leaving the condensing coil
zone to decelerate. This means that the velocity of air passing over the
condensing coils
can be increased, relative to the velocity of air passing through the drift
eliminators.
This higher velocity can help to reduce fouling of the tube.
Further, it has been surprisingly discovered that a condensing coil bundle
with a
reduced plan area, relative to the drift eliminators can be employed. A
further
consequence of this is that less condensing coil is required for the same
condenser
performance. This means that a lower cost evaporative condenser can be
produced, as
the condensing coil bundle represents the single-most expensive component of
such a
condenser.
In addition, an increased flow of refrigerant can be passed through the
condensing coil bundle, because the greater air velocity is able to bring
about
condensation of a relatively greater amount of refrigerant.
Furthermore, this means that, as an alternative to using known hot-dipped
galvanized carbon steel condensing tube, a more expensive and/or stronger
material
(e.g. stainless steel) can be used to form the one or more condensing coils,
with the
result that longer life, less corrosion and, optionally, thinner wall material
for the coil
(tube) can be employed. Notwithstanding, and if preferred, DN 8, 10, 15, and
20
(Schedule 40) seamless, hot-dipped seamless galvanized carbon steel tube can
still be
used to form the one or more condensing coils.
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In one embodiment, each of the one or more condensing coils may employ
stainless steel tube (e.g. 304 or 316 stainless steel of 4.76 -31.8 mm outside
diameter
and 0.5 ¨ 1.6 mm thickness). The use of 304 stainless steel can offer better
conductivity, whereas 316 stainless steel can offer better corrosion
resistance. Such tube
material can perform favourably in comparison to a known condensing coil tube
of
galvanized mild carbon steel. The use of very small diameter tube can be
suitable for
certain small-scale applications.
The use of stainless steel tube material (i.e. due to corrosion/chemical
resistance,
increased refrigerant pressure capacity, etc) can also allow a natural
refrigerant, such as
a propane and/or isobutane hydrocarbon, CO2, ammonia, etc, to be employed.
In one embodiment the one or more condensing coils can be arranged as a
bundle (e.g. of two or more nested coils) in the condensing coil zone. For
example, the
condensing coil zone may comprise a section of the condenser of generally
constant
cross-sectional area (e.g. an air plenum of circular, square, rectangular, etc
hollow
1 5 section).
In one embodiment the divergent part of the zone can be configured to cause
the
airstream to decelerate in a gradually decreasing manner.
In one embodiment the divergent zone can comprise a hollow frustum (hollow
air plenum) through which the airstream flows. Such a hollow frustum may be
located
on the air exit side of the condensing coil plenum. For example, when the
condensing
coil plenum is of circular section, the divergent frustum each comprise a
conical
frustum, or a square-to-circular frustum-like prism; when the condensing coil
plenum is
of square section, the divergent frustum may comprise a square frustum; etc.
In one embodiment, the drift eliminators may be immediately located at an air
leaving side of the divergent zone.
In one embodiment, the condenser may comprise an air inlet chamber located at
an air entry side of the condensing coil zone.
In one embodiment, the mechanism for wetting the one or more condensing
coils may comprise one or more spray nozzles. The spray nozzles may be
arranged with
respect to the divergent zone to spray water onto the one or more condensing
coils in a
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direction that is counter to the airstream flow through the one or more
condensing coils.
For example, the spray nozzles may be arranged in the divergent zone, and may
spray
the water generally as a liquid cone into the condensing coil zone.
Alternatively, the mechanism for wetting the one or more condensing coils may
comprise water distribution channels, such as those with serrated edges,
internal slots,
etc.
The condenser can comprise a water collection zone (e.g. located at a base of
an
air inlet chamber). The collection zone can collect water that has passed
through
condensing coil zone.
1 0 The condenser can further comprise a recycling system for recycling the
collected water to the wetting mechanism, to maximize condenser efficiency. In
one
embodiment the recycling system can comprise a pump for pumping the collected
water
via pipework to the wetting mechanism. For example, an offtake pipe can extend
from
the base of the air inlet chamber to the pump, and a delivery pipe can extend
from the
pump outlet to the wetting mechanism (e.g. to spray nozzle, distribution
pipework, etc).
In one embodiment, the recycling system can further comprise, as necessary, a
water make-up mechanism for maintaining a predetermined amount of water (e.g.
in the
water collection zone) for effective operation of the evaporative condenser.
Such make-
up water can include that eliminated (captured) by the drift eliminators.
2 0 In one embodiment, the evaporative condenser can further comprise a
heat
exchanger (e.g. a separate, laterally located discrete heat exchange unit).
The collected
water can be passed through the heat exchanger prior to recycling it to the
wetting
mechanism. In addition, the condensed refrigerant can be passed through the
heat
exchanger to exchange heat with the recycled collected water. Such a heat
exchanger
can be used to sub-cool the condensed refrigerant to further improve the
operational
efficiency of the evaporative condenser.
Also disclosed herein is an evaporative condenser that comprises the
collection
zone for collecting water that has passed through condensing coil zone, and
that
comprises the heat exchanger through which the collected water is passed prior
to
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recycling it to the wetting mechanism, and through which the condensed
refrigerant is
passed to exchange heat with the recycled collected water.
Also disclosed herein is an evaporative condensation process forming part of a
refrigeration or air-conditioning cycle.
The process comprises passing refrigerant through one or more condensing
coils. The process also comprises wetting the one or more condensing coils
with water.
The process further comprises passing an airstream over the one or more wetted
condensing coils whereby refrigerant is caused to condense within the coils,
and
whereby a portion of the water is caused to evaporate into the airstream. The
process
additionally comprises eliminating water that is present in the airstream
leaving the one
or more condensing coils.
In accordance with the present disclosure, the process is conducted such that
the
velocity of the airstream leaving the one or more condensing coils is caused
to
decelerate prior to eliminating the water that is present in the airstream.
As outlined above, this can lead to a reduced plan area (and hence a lesser
amount) of the one or more condensing coils relative to the drift eliminators
(with the
attendant advantages as outlined above).
Also disclosed herein is an evaporative condensation process in which the
water
that passes through the one or more condensing coils is collected and recycled
to wet
2 0 the one or more condensing coils with water. Further, in such a
process, heat can be
exchanged between the condensed refrigerant and the collected water, prior to
recycling
it to wet the one or more condensing coils.
The process as disclosed herein can take place in an evaporative condenser as
set forth above.
In the process as disclosed herein, the refrigerant condensed in the one or
more
condensing coils can comprise a natural refrigerant (e.g. a hydrocarbon such
as propane
and/or isobutane, CO2, ammonia, etc) or a chemical refrigerant (e.g. a
hydrofluorocarbon, a hydrochlorofluorocarbon, a perfluorocarbon, a
hydrofluoroolefin,
etc).
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BRIEF DESCRIPTION OF THE DRAWINGS
Notwithstanding any other forms which may fall within the scope of the
condenser and process as set forth in the Summary, specific embodiments will
now be
described, by way of example only, with reference to the accompanying drawings
in
which:
Figure 1 shows cross-sectional side schematic of an evaporative condenser
having a condensing coil zone in which one or more condensing coils are
arranged, and
a divergent zone extending away from the condensing coil zone;
Figure 2 shows a detail of Figure 1, to illustrate a variant of the
evaporative
condenser that further comprises a side heat exchanger; and
Figures 3A and 3B respectively show cross-sectional and side schematics of an
evaporative condenser having a convergent-divergent zone in which one or more
condensing coils are arranged;
Figure 4 shows a cross-sectional side schematic of an evaporative condenser
that is similar to Figure 1, but for different process parameters in
accordance with the
Examples;
Figure 5 is a graph showing CO2 and water temperature profiles in accordance
with the Examples;
Figure 6 is a graph showing CO2 heat capacity profile in accordance with the
Examples;
Figure 7 is a graph showing water flow down the tube bundle in accordance
with the Examples;
Figure 8 is a graph showing overall heat transfer coefficient & pressure loss
in
accordance with the Examples;
Figure 9 is a graph showing a heat rejection profile based on a commercially
available transcritical CO2 compressor at 5 C sat. suction, with 5 K useful
suction
superheat and 5 C CO2 liquid temperature, in accordance with the Examples;
Figure 10 is a graph showing performance of a commercially available
transcritical CO2 compressor at 50 Hz. 30 kW/27.2 m3/h, in accordance with the
Examples; and
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Figure 11 is a graph showing the variation in COPs of NH3, R22 R507A,
Propane and R134a with Saturated Condensing Temperature, in accordance with
the
Examples.
DETAILED DESCRIPTION OF SPECIFIC EMBODIMENTS
Specific forms of an evaporative condenser, and an evaporative condensation
process that form part of a refrigeration or air-conditioning system/cycle,
will now be
described.
Evaporative condenser embodiments designated 10 and 100 are respectively
shown in Figure 1 & 2 and Figures 3A & 3B. The evaporative condensers
embodiments
10 and 100 are able to employ both chemical and natural refrigerants (as set
forth
above). Figures 4 to 11 relate to embodiments described in the Examples.
In Figures 1 to 3, similar components of the evaporative condensers 10 and 100
are numbered similarly, but with 100 added to the embodiment of Figure 3. It
should be
further understood that, for the sake of brevity, the following description
does not re-
describe those similar or like components that re-appear in the embodiment of
Figure 3,
which therefore should be taken to have been described.
The preferred evaporative condenser 10 of Figures 1 and 2 comprises two or
more nested condensing coil bundles 12 that have flowing (for condensing)
therewithin
the selected refrigerant of the system. The condensing coil bundles 12 are
arranged in a
2 0 condensing coil zone in the form of a rectangular airflow plenum 13.
The evaporative condenser 10 also comprises a mechanism in the form of spray
nozzles 14 formed in a distributor tube 15 for wetting the condensing coil
bundles 12 by
spraying them with cones 16 of water (e.g. at a rate of 3 kg/m2 as shown).
Alternatively,
water distribution channels, such as those having serrated edges or internal
slots, can be
employed.
The spray nozzles 14 are arranged to spray water onto the condensing coil
bundles 12 in a direction that is counter to the airstream flow therethrough
as shown.
The evaporative condenser 10 also comprises a fan arranged in a fan housing at
an upper end of the condenser. Such an arrangement is actually shown in the
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embodiment of Figure 3 as a fan 118 arranged in a fan housing 120 located at
an
uppermost end of the condenser (see Figure 3A). The same or similar
arrangement can
be employed in the embodiment of Figures 1 & 2. In this regard, the fan causes
air to be
drawn via an air inlet 21 into an air inlet chamber 22 that is arranged
towards the lower
end of the condenser 10.
In the embodiment of Figures 1 & 2 the airstream A enters at a volumetric flow
rate of e.g. 8.1 m3/s, first passing through mesh filters and then into the
air inlet
chamber 22, before it is caused by the fan to flow up to and through the
condensing coil
bundles 12. The air pressure differential can be maintained by the fan at e.g.
160Pa.
In the embodiment of Figure 3 the airstream A enters with a velocity of e.g. 3
m/s and a wet bulb temperature of e.g. 23 C, first passing through optional
mesh filters
124 depending on air contamination and air inlet slots 126 and then into the
air inlet
chamber 122, before it is caused by the fan 118 to flow up to and through the
condensing coil bundles 112.
The evaporative condenser 10 further comprises drift eliminators 30 which are
arranged within the condenser adjacent to an upper end thereof The drift
eliminators 30
remove free water from the airstream once it has flowed past the condensing
coil
bundles 12 and spray nozzles 14.
In the embodiment of Figures 1 & 2, the evaporative condenser 10 comprises
2 0 the rectangular airflow plenum 13 immediately followed by a divergent
airflow zone in
the form of a frustum-shaped plenum 40. The rectangular airflow plenum 13 can
be of
square, rectangular, etc hollow section (e.g. of bent and welded plastic or
metal
sheet/plate). The divergent plenum 40 can also be of hollow section (e.g. of
bent and
welded plastic or metal sheet/plate), but formed so as define the frustum.
When, for
example, the plenum 13 is of square section, the divergent plenum 40 comprises
a
square or rectangular frustum.
However, in the embodiment of Figure 3, the evaporative condenser 100
employs both a convergent airflow zone 135 and a divergent airflow zone 140
located
on either side of an intermediate rectangular airflow plenum 113 that contains
the
condensing coil bundles 112. The plenum 113 has a constant cross-sectional
area and
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interconnects the convergent airflow zone 135 and the divergent airflow zone
140. The
intermediate airflow plenum 113 can again be of square, rectangular, etc
hollow section
(e.g. sheet/plate). The convergent airflow zone 135 and divergent airflow zone
140 can
again be of hollow section (e.g. sheet/plate), but each formed so as define
the frustum.
When, for example, the intermediate zone 113 is of square section, the
convergent and
divergent frustums may each comprise a square or rectangular frustum.
In the embodiment of Figures 1 & 2, the fan is operated such that the
airstream
A is already at a higher velocity at the condensing coil bundles 12 relative
to the drift
eliminators 30. Having flowed past the condensing coil bundles 12, the
airstream A
flows into the divergent airflow plenum 40, passing through the water cones
16.
Because of the progressively increasing cross-section of the divergent airflow
plenum
40, the airflow is able to decelerate to an acceptable velocity before it
reaches and
passes through the drift eliminators 30. The evaporative condenser 10, and in
particular,
the divergent airflow plenum 40, is configured such that this velocity is at a
level
whereby an environmentally acceptable minimum amount of free water in the
airstream
can be eliminated therefrom. In this regard, the airflow rate at the drift
eliminators 30
can decelerate to approximately 3.5 m/s.
It will be seen that the drift eliminators 30 are arranged immediately at the
air
exit of the divergent airflow plenum 40, whereby the airflow is not permitted
to
2 0 decelerate more than is necessary.
Thus, the embodiment of Figures 1 & 2 does not employ a convergent airflow
zone. Rather, the airflow velocity from the air inlet chamber 22, and through
the
condensing coil bundles 12, is approximately 5 m/s, until the airflow reaches
the
divergent airflow plenum 40, whereupon the airflow gradually decelerates to
approximately 3.5 m/s at the drift eliminators 30.
However, in the embodiment of Figure 3, the condensing coil bundles 112 are
arranged in the intermediate airflow zone 113. The configuration of these
zones is such
that the airstream A flows through and is accelerated in the convergent
airflow zone 135
to the condensing coil bundles 112 located in the intermediate airflow plenum
113 (e.g.
to approximately 5 m/s). Having flowed past the condensing coil bundles 112,
the
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airstream A flows into the divergent airflow zone 140, passing through the
water cone
16, and decelerating before reaching the drift eliminators 130. Again, the
drift
eliminators 130 are arranged immediately at the air exit of the divergent
airflow plenum
140.
The convergent airflow zone 135 is configured to cause the airstream A to
accelerate, such as in a gradually increasing manner. Conversely, the
divergent airflow
zone 140 is configured to cause the airstream to decelerate, such as in a
gradually
decreasing manner. This means that the velocity of air passing through the
intermediate
airflow zone 113 and over the condensing coil bundles 112 is increased,
relative to the
velocity of air that passes into the air inlet chamber 122 as well as through
the drift
eliminators 130. For example, in the configuration depicted, the airflow rate
in the
intermediate airflow zone 113 is approximately double at ¨ 5 m/s, (i.e. about
45%
higher than) the 3.5 m/s air velocity through the drift eliminators.
In either embodiment, and as a result of this increased airflow rate passing
over
the condensing coil bundles 12, 112, it has surprisingly been discovered that
a
condensing coil bundle with a reduced plan area, relative to the drift
eliminators 30, 130
can be employed. As a further consequence of this increased airflow rate, it
has
surprisingly been discovered that less condensing coil is required for the
same
condenser heat rejection performance.
2 0 The result is that a lower cost evaporative condenser can be produced,
as the
condensing coil bundle represents the single-most expensive component of the
condenser. Alternatively, instead of using known thick-wall, hot-dipped
galvanized
carbon steel condensing tube for the coil bundle 12, 112 a more expensive
and/or
stronger material, such as stainless steel tube, can be used to form the coil
bundle 12,
112. In such case, the result is longer coil life, less corrosion and, if
desired, thinner
wall material for the tube in the coil bundle. In this regard, the coil bundle
12, 112 can
comprise stainless steel tube, such as 304 or 316 stainless steel of 4.76 -
31.8 mm
outside diameter and 0.5 ¨ 1.6 mm thickness. Such tube is observed to perform
well in
comparison to known condensing coil tube of galvanized mild carbon steel. The
corrosion and chemical resistance, as well as increased refrigerant pressure
capacity,
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that is provided by such stainless steel tube materials also allows a natural
refrigerant,
such as a propane and/or isobutane hydrocarbon, CO2, ammonia, etc, to be
employed in
the evaporative condenser 10, 100.
Another consequence of the increased airflow rate over the condensing coils is
that an increased flow of refrigerant can be passed through the condensing
coil bundle
12, 112 because the greater air velocity is able to bring about condensation
of a
relatively greater amount of refrigerant.
The condenser 10 also comprises a water collection zone in the form of a basin
50 located at a base of (i.e. adjacent to) the air inlet chamber 22. The basin
50 collects
excess spray water that has passed through or from condensing coils.
To maximize condenser efficiency, the condenser 10 additionally comprises a
recycling system for recycling the collected water to the distributor tube 15
for feeding
to the spray nozzles 14. In this regard, the recycling system comprises a pump
52 for
pumping the collected water via pipework to the distributor tube 15. The pump
52
draws water out of the basin 50 via an offtake pipe 54. A delivery pipe
section 56 then
extends from the pump outlet to connect with the distributor tube 15.
The recycling system also comprises water make-up 58 (e.g. at 383 kg/h) for
maintaining a predetermined amount of water in the basin 50 for effective
operation of
the evaporative condenser. Such make-up water can include a supply of water
that has
2 0 been eliminated (captured) by the drift eliminators 30.
In a variation of the evaporative condenser shown in the detail of Figure 2,
the
condenser 10 can further comprise a side heat exchanger unit 60. The water in
the basin
50 can be pumped via pump 52 and into and through the heat exchange unit 60,
prior to
being recycled to the distributor tube 15 via the delivery pipe section 56.
Such a unit
can also be fitted to the embodiment of Figure 3.
In this variation, the condensed refrigerant in the condenser tubes can also
be
passed via refrigerant delivery pipe 62 to and through the heat exchange unit
60 to
exchange heat with the recycled water from the basin 50. In the heat exchange
unit 60
the relatively cool basin water can sub-cool the condensed refrigerant, for
example,
from 30 C to around 26.5 C. This can further improve the operational
efficiency of the
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refrigerating system. The refrigerant (e.g. CO2) leaving the heat exchange
unit 60 as the
stream 64 can be at a sub-cooled temperature (e.g. around 26.5 C).
Examples
Non-limiting examples of the present condenser and process will now be
provided in order to illustrate the theoretical basis of the condenser and
process, and to
better understand the condenser and process in operation.
Example 1 ¨ Process Design Model
A design model for the application to subcritical CO2 condensing of
evaporative
condensers, such as those depicted in Figures 1 to 3, was developed. More
specifically,
the benefits of applying evaporative condensing techniques for the
condensation of
subcritical CO2 were examined. Such benefits included lower design pressures
compared to trans-critical operations, lower energy consumption, and lower
running
and operating costs. It was noted that hot gas defrosting could also become a
standard
feature of subcritical CO2 refrigeration plant operations.
Firstly, however, it was noted that ammonia can be condensed at 30 C in an
evaporative condenser with an entering air wet bulb temperature of 24 C. In
the
developed design model it was shown that an evaporative condenser for
subcritical CO2
condensing at 30 C (i.e. 1.1 K below the critical point) was able to be
designed for a
wet bulb of 24 C.
Secondly, it was noted that average climate conditions in much of Europe,
including the warmer climates in Spain, Italy, Greece and Turkey, were
suitable for
evaporative condensers to condense subcritical CO2 at 30 C. Canada, large
parts of the
USA and China, and most of Australia below the tropic of Capricorn were also
noted to
have climates suitable for the application of evaporative condensers to
subcritical CO2
condensing. The thermodynamic and transport properties of subcritical CO2 at
30 C
were noted to change significantly with temperature. Thus, the effect these
changes
have on CO2 temperature profile, heat transfer and pressure loss for a
particular design
was also shown.
For example, an examination of average climate conditions revealed that much
of Europe, including Spain, Italy, Greece and Turkey, has a climate where
evaporative
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condensers may be applied for the condensing of CO2 at subcritical conditions
at a
condensing temperature of 30 C or lower 100% of the time in many locations.
For
example, the only location in Europe where the 5% design Wet Bulb temperature
incidence exceeded 24 C was Adana in Turkey (where the 1 and 2.5% wet bulb
incidence design levels are at 26 C). At Thessaloniki in Greece the 1% wet
bulb design
incidence is at 25 C, but the 2.5% and 5% wet bulb design incidence levels
are at 24
C. The next highest 1% wet bulb design incidence level of 24 C occurred at
Gibraltar,
Barcelona, Valencia, Milan, Istanbul and Izmir.
Finally, it was concluded that the use of evaporative condensers for CO2 in
temperate and many subtropical climates could make CO2 refrigeration as
ubiquitous as
any chemical refrigerant and would compete successfully with ammonia when it
needed
to be used in an indirect application (such as the heating and cooling of
office buildings
and hospitals for example).
When CO2 refrigeration was revived about 20 years ago, air cooled gas cooling
(some with adiabatic assistance by spraying water onto the air inlet face of
the finned
coil gas cooler) was applied almost universally. It was noted that this
resulted in
virtually all CO2 refrigeration systems needing to run in trans-critical mode
because the
air cooling temperature is close to, or exceeds, the CO2 critical temperature
of 31.1 C.
More often than not the summer design CO2 exit temperatures from an air
2 0 cooled gas cooler were higher than the critical temperature, and this
resulted in the
compressors needing to operate at a pressure of 90 bar or higher to ensure a
reasonable
COP. The summer design COPs of trans-critical CO2 compressors were generally
lower than those of air cooled HFC or evaporatively cooled ammonia systems.
It was therefore proposed to reduce the temperature of the condenser cooling
medium to a level which would allow a complete subcritical CO2 refrigeration
cycle.
This was accomplished with an evaporative condenser, where the ambient air Wet
Bulb
(WB) temperature was the effective cooling medium temperature, rather than the
ambient air Dry Bulb (DB) temperature in the case of an air cooled condenser
or gas
cooler.
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Issues noted included the need for a water supply, water consumption and water
treatment, and control of a minimum condensing temperature as currently
mandated by
some compressor suppliers. Another issue was the control strategies to handle
inadvertent trans-critical conditions. Recommendations were made to address
these
issues.
A Rating Model for a CO2 Evaporative Condenser
Rating example
Figure 4 shows a schematic flow sheet of an evaporative condenser that will
now be further described. In the flow sheet of Figure 4 water was recycled
over the
1 0 tube bank, so that spray water temperature was the same as the basin
water temperature.
The specified parameters were: (a) air velocity and wet and dry bulb
temperatures, (b) spray water flow rate, (c) bundle dimensions, and (d)
leaving CO2
comprising saturated liquid at 30 C and 7.2 MPa.
Mass and energy balances in evaporative coolers
Qureshi (2006) and Heyns (2009) published five simultaneous non-linear
differential equations describing air-water-process fluid interactions in
evaporative
cooling.
dmw
¨da = hd(Wint ¨
1)
dW 1 dmvt.
da rit da
a 2)
dka lid
¨d = õ.= ¨ ha) + (1 ¨ Le)iv(Wint ¨ 147))
a
ma 3)
dl:r U,
¨da = ¨(yr ¨ Tw)
mr 4)
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dT,, 1
¨ = (11, dh, ¨ CpwTwdmw ¨ mrCprdTT)
da rnwLf PW 5)
1 1 do
fir + _______________________
U0 h d.h.
a 6)
The equations were solved by writing a program using a fourth order Runge-
Kutta routine written in Microsoft's VBA behind a Microsoft Excel spreadsheet
with
pass length divided into forty intervals. The solution was trial and error
because the
basin water temperature at air entry was guessed and adjusted iteratively
until it was the
same as the calculated water temperature at the air exit.
The solution proceeded "backwards" along a tube pass from CO2 exit to entry,
starting at the air inlet with saturated liquid refrigerant at 30 C, and
proceeded up, as if
heating, ending with superheated vapour at a calculated discharge temperature.
The
program allowed for both two-phase condensation and single phase vapour de-
1 0 superheating.
Verifying the model.
There was no analytical solution to the five equations by which the numerical
solution can be validated. However, two findings were noted: when the leaving
and
entering water temperatures were equal; (a) the CO2 enthalpy change was equal
to the
moist air enthalpy change, and (b) the heat duty calculated for ammonia
condensing at
30 C was within 9% of duty computed using the simplified Merkel model
(Merke1,1926) which is based on constant condensing temperature.
Model predictions
Figures 5 and 6 show CO2 and water temperature profiles. The shape of the CO2
2 0 temperature profile (Figure 5) was a surprise ¨ it was much flatter
than expected. Over
about 37% of exchanger surface, the CO? vapour temperature only reduced from
32 to
30 C at interval number 29. This was a consequence of the very high heat
capacity just
above 30 C (Figure 6).
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The model predicted that 67% of exchanger surface would be needed for
sensible cooling. Enthalpy data for CO2 at 30 C, close to the critical point,
showed that
68% of the rejection was sensible cooling, unlike ammonia where it is only
10%.
Water temperature profile was skewed to the left, compared to profiles where
sensible cooling was relatively small, reflecting the larger proportion of
sensible
cooling for CO2 near the critical point.
Water evaporation
Figure 7 shows the water flow down the tube bundle. Another surprise was that
water did not evaporate to air at the top of the bundle. Here, water
temperature was low
though rising, and water contacted air with a humidity ratio higher than the
humidity
ratio at the air-water interface, so some condensation occurred and water flow
increased.
Effect of property changes
Figure 8 illustrates the effects of changes in heat capacity, density,
viscosity and
thermal conductivity with temperature on overall heat transfer coefficient and
pressure
loss over each solution interval. Interval zero was where the hot discharge
gas enters.
There was considerable rise in overall heat transfer coefficient as the CO2
temperature
approached 32 C, with a corresponding decrease in pressure loss per metre as
the
vapour phase transitioned to two phases.
In the model 0.845 was used for the Lewis number in equation (3). With a
Lewis number of 1.00, the surface area required for the same heat duty was
reduced by
only 1.4%.
Remarks
The case modelled was extreme in the respect that CO2 condensing at 30 C was
very close to its critical point. It was noted that at lower condensing
temperatures the
proportion of sensible cooling would reduce and the variation of properties
with
temperature would be much reduced. Reference was made to Figure 9.
It was further noted that heat rejection predicted by the Merkel simplified
model
was about 22% lower than the differential model with CO2 condensing at 30 C,
which
was not unexpected given the significant proportion of sensible cooling.
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CO2 Compressor Subcritical Energy Performance
Effect of condensing temperature on cycle performance
In Figure 10, five COP plots were produced for a commercially available, semi
hermetic, trans-critical CO2 compressor with a swept volume of 27.2 m3/h at 50
Hz and
30 kW 4 pole motor.
Referring to curve 1 in Figure 10, the COP ranged from 6.27 at +30 C Saturated
Condensing Temperature (SCT) to 18.0 at an SCT of +16 C at a Saturated Suction
Temperature (SST) of +10 C. A +10 C SST would allow an Evaporating Temperature
(ET) of +11 C with a suction pressure drop corresponding to 1 K boiling point
suppression. 11 C was noted as a reasonably efficient evaporating temperature
for
direct cooling of Air Conditioning (AC) air, allowing a relatively large
diffusion in air
temperature across the cooling coil, thus limiting the volume of air which
would need to
be circulated, and thereby reducing fan energy consumption and the resulting
parasitic
heat load. This in turn would lead to a reduction in the required energy input
into the
compressor thereby lifting the overall energy efficiency of the system as a
whole.
Curve 2 showed the COP ranging from 4.45 to 11.67 at 30 C to 16 C SCT at an
SST of +5 C. This would allow chilled water production for AC for retrofitting
into
existing buildings and application to new buildings.
In both the above two cases the AC compressors could act also as parallel
2 0 compressors for refrigeration duties at ¨5 C SST, such as maintaining
chill storage
temperatures at around 0 C and high stage duties for two stage CO2 systems
applied to
cold storage and blast freezing applications.
In such cases the high stage compressors would operate with virtual CO2 gas
cooler exit temperatures of +5 C and +10 C, which resulted in COP curves 3 and
4
respectively. COP curve 3 ranged from 4.7 to 7.88 at an SST of ¨5 C at SCTs
ranging
from +30 to +16 C and a virtual gas cooler exit of +5 C. COP curve 4 showed
the COP
ranging 4.45 to 7.04 with a virtual gas cooler exit of +10 C, and SST of ¨5 C
and the
SCT ranging from +30 to +16 C. It was noted that this could be improved with a
Suction Heat Exchanger (SHEX) in the compressor suction to bring the
performance
closer to curve 3.
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Effect of ambient Wet Bulb temperature on condenser performance
Figure 4 shows the general details on CO2, air and water. The Effect of
ambient
Wet Bulb temperature on condenser performance is shown in the results set
forth in the
following table:
Total Sat. Ambient Leaving air Mass fluxes
Pressure Full Gross
heat Cond. air condition G kg/m2.s drops power heat
rejection Temp. condition cons., flux
(THR) (SCT) DB WB DB RH CO2 Air Water CO2 Air BkW kW/m2
kW C C C C % Gc02 Gair Gh2o kPa Pa
470 30 35 24 28.7 100 240 3.46 3.0 5 60 1.8 3.22
596 30 35 22 28.3 99.9 305 3.46 3.0 8 60 1.8 4.08
712 30 35 20 27.9 99.8 364 3.46 3.0 11 60 1.8 4.87
658 28 30 18 25.7 99.8 312 3.52 3.0 8 62 1.85 4.50
545 24 24 14 21.5 99.7 236 3.59 3.0 6 62 1.85 3.73
475 20 18 10 17.3 99.5 193 3.66 3.0 4 63 1.9 3.25
407 16 12 6 13.3 99.4 157 3.74 3.0 3 63 1.9 2.78
NB. Tube bundle: 84 circuits, 8 passes, 146.2m2
Relative energy efficiency of ammonia, R22, R507A, propane, and R134a
The COPs for these refrigerants at identical operating conditions were shown
in
1 0 Figure 11. The results confirmed that ammonia was the best refrigerant
of these. A
surprise was the low COP of R134a. At SCTs of 16 and 35 C the R134a COPs were
respectively 42 and 31% lower than those of ammonia. Furthermore the COP of an
R134a compressor at +16 C SCT was, at 3.84, about the same as the COP of an
ammonia compressor at +35 C SCT at identical suction conditions. This
confirmed
that R134a had both high direct and indirect Global Warming Potential (GWP).
The
performance of R507A was 11 to 16% less efficient than R22 at 25 to 35 C SCT.
HFC
R507A had no Ozone Depletion Potential like HCFC R22, but the 100 years GWP of
R507A was 3,895, more than double the 100 year GWP of 1,810 of R22.
Overall heat transfer factor, Uo
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Referring again to Figure 8, Uo was very considerably higher than the familiar
Uo in the case of ammonia condensing, where Uo ranges from about 450 to 550
w/m2.K at superficial air velocities of 2.6 to 3.05m/s. A 3m/s superficial air
velocity
was chosen as a maximum in the models to ensure that the drift eliminators
would be
able to catch most of the free water suspended in the upward air draft.
The average Uo in Figure 8 for the CO2 evaporative condenser in Figure 4 was
about 1,050 w/m2.K. This was noted to be more than double the average value
for
ammonia at virtually the same superficial air velocity entering the condensing
tube
bundles. This was remarkable, when it was considered that at 30 C condensing
68% of
1 0 the heat to be removed is sensible superheat and only 32% is actually
latent heat of
condensation at 30 C as shown in Figure 9. The high overall heat transfer
factor was
attributed to by the high CO2 mass flux of 338.7 kg/i-n2.s in the 76 off 55
metre
equivalent length circuits causing a calculated pressure drop of 15 kPa. This
is the
maximum value acceptable to facilitate CO2 condensers operating in parallel
without
requiring too great a drop leg to avoid liquid hold up in operating condensers
should
one condenser not be operating.
As with evaporators, the high AP/AT ratio of CO2 allowed high mass fluxes in
the condenser circuits giving high rates of heat transfer, allowing fewer
longer circuits,
which also made for more economical manufacture of the tube bundle.
It was noted that ammonia mass fluxes in evaporative condensers range from
about 25 to 40 kg/m2.s and are frequently lower than 25. The pressure drop was
a
concern with ammonia condensers, as excessive pressure drop in an ammonia
evaporative condenser lifts the discharge pressure, and thus the Saturated
Condensing
Temperature (SCT), resulting in increased energy consumption.
Consequences of minimum airflow
Again referring to Figure 4, the calculated leaving air Dry Bulb Temperature
is
29.3 C at 100% RH and hence the leaving Wet Bulb Temperature was also 29.3 C.
This was only 0.7 K lower than the SCT of 30 c. This was possible because the
top
tubes were at a temperature of 77 C and the high proportion of sensible
superheat
ensured that there was a high leaving approach TD available of 47.7 K. It was
noted
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that this was not possible in ammonia evaporative condensers, where minimum
leaving
temperatures approaches between the ammonia SCT and the leaving Wet Bulb are
rarely less than 3K and not less than 2.5K at design conditions. Little
airflow also
resulted in minimum fan energy consumption.
Conclusion
Subject to satisfactory performance testing of a full scale prototype CO2
evaporative condenser, it was concluded that the application of the
evaporative
condensers in the higher latitude subtropics with maximum design Wet Bulb (WB)
temperatures of 24 to 25 C showed a great deal of promise. The CO2 evaporative
showed even more promise in areas with more temperate and cool to cold
climates
where ambient WB temperatures are lower.
According to the above conclusion, suitable areas for the application of
evaporative condensers to the condensing of subcritical CO2 compressor
discharge
gases was thus feasible in virtually all of Europe (including the
Mediterranean
countries), the USA except for the Southern States bordering the Gulf of
Mexico and
the Atlantic Ocean, and many of the Mid West States as far North as Minnesota.
Experimentation also showed that evaporative gas cooling with ambient Wet
Bulb temperatures of 28 to 29 C and ambient air WB to CO, exit temperature
approaches of 3 K were entirely feasible. This was ascribed to the fact that,
in trans-
2 0 critical mode, there was only sensible heat transfer at a larger LMTD
without the
condensing phase (Figure 9) and relatively high trans-critical fluid densities
coupled
with high heat capacities similar to that shown in Figure 6.
The application of evaporative cooling to both the condensing of CO2 at
subcritical and gas cooling at trans-critical CO2 resulted in efficient
refrigeration at high
COPs which were comparable to, and in many cases higher than, the COPs
achieved
with conventional refrigerants operating below their critical points. This
opened the
way for the worldwide application of CO2 refrigeration. This is particularly
true in
applications where CO2 is used for Air Conditioning duties at +5 and +10 C
compressor Saturated Suction Temperature for chilled water, and DX or pumped
CO2
AC applications respectively.
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It was further noted that the AC compressors may also act as parallel
compressors for any remaining refrigeration duties in a facility such as a
supermarket
where both chilling and freezing duties are required at high to very high COPs
as shown
in Figure 10.
Indeed, when comparing Figure 10 and 11 it was clear that, in the subcritical
condensing phase, CO2 outperformed conventional chemical refrigerants such as
R22,
R507A and R134a, as also found by Pearson (2010). Additionally, CO2 rivalled
or
outperformed ammonia and propane under most operating conditions and
particularly
so where parallel compression was involved.
1 0 At a high Wet Bulb temperature of e.g. 28 C conventional evaporative
condensers would be able to operate at 40 C SCT resulting in COPs of 3.37,
3.34, 2.71,
2.96 and 2.38 for NH3, R22, R507A, propane and R134a respectively as shown in
Figure 9. Thus, to develop highly efficient CO2 refrigeration larger
compressors are
needed, for example, modified versions of CNG fuel compressors.
Nomenclature
In the Examples:
a outside surface area m2
ma air flow rate kg dry air s-1
s-i
mw water flow rate kg
mr s-i
CO2 flow rate kg
MCI SW saturated air enthalpy at air-water interface J kg-1 dry
air
ha air enthalpy J kg-1 dry air
hd mass transfer coefficient kg
water vapour enthalpy J kg-1
Tw water temperature
Le Lewis number
Tr CO2 temperature oc
Cpw heat capacity of liquid water J kg-1 K-1
Cpa heat capacity of moist air J kg-1 K-1
U0 overall heat transfer coefficient W 111-2
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W air humidity ratio kg water kg-1 dry
air
air humidity ratio at air-water interface kg water kg-1 dry
air
water-tube heat transfer coefficient W m-2 lc1
h, CO2 heat transfer coefficient W m-2 K-1
d, tube inside diameter
dõ tube outside diameter
ff fouling factor K m2 WI
Model parameters
1. NIST (2011) data were used for thermodynamic and transport properties
of saturated and superheated CO2;
2. hw in equation (6) was calculated from Mizushima and Miyasita (1967),
equation (A.8) in Qureshi and Zubair (2006);
3. hd in equation (3) was calculated from Mizushima and Miyasita (1967),
equation (A.13) in Qureshi and Zubair (2006);
4. For two phase CO2 flow, hi in equation (6) was calculated from Shah's
(2009), Qureshi and Zubair (2006) equations (A.6) and (A.7); pressure loss was
calculated from Miiller-Steinhagen and Heck correlation (ASHRAE, 2005);
5. For single phase CO2 vapour flow, hi in equation (6) was calculated from
the Dittus-Boelter correlation Nu = 0.023Re" Prm; pressure loss was calculated
from a friction factor = 0.079Re- =25;
6. Air pressure drop across the tube bank was calculated from Mills (1999),
section 4.5.1, p. 316.
The following References were used in formulating the model:
1. ASHRAE, 2005, 2005 Fundamentals, page 4.12-13
2. Heyns J, Kroger D, 2009, Performance characteristics of an air-cooled
steam condenser incorporating a hybrid (dry/wet) dephlegmator, Appendix A,
PIER Report, CEC-500-2013-065-APA
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3. Merkel, F., 1926, Verdunstungskuling, VDI-Zeitschrift, Vol. 70, pp. 123
¨ 128
4. Mills A.F., 1999, Basic Heat & Mass Transfer, 2nd ed., A.F., Prentice
Hall.
5. Mizushima, T., R. Ito and H. Miyasita, 1967, Experimental study of an
evaporative cooler, International Chemical Engineering, Vol. 7, pp. 727-732
6. NIST 2011, http://webbook.nist.gov/chemistry/fluid/ Thermophysical
Properties of Fluid Systems
7. Qureshi B, Zubair S, 2006, A comprehensive design and rating study of
1 0 evaporative coolers and condensers. Part I Performance evaluation, Int.
J.
Refrigeration, 29: 645-658.
8. Shah M, 2009, An improved and extended general correlation for heat
transfer during condensation in plain tubes, HVAC&R Research, 15 (5)
9. Pearson, S. Forbes, 2010, Use of carbon dioxide for air conditioning and
general refrigeration, IIR-IOR 1st Cold Chain Conference, Cambridge, UK.
Example 2 ¨ Design Model Outputs
The following data points were produced by the design model to illustrate
condenser capacity variation with the superficial air velocity:
0
t..)
o
,-,
Evaporative Condenser Model 1
u,
,-,
-4
t..)
,-,
Required rejection 485 kW
cio
=
Calculated rejection 485
Estimate for Condenser circuits
and pressure drop
kW
Rejection 485 kW
Refrigerant I CO2, Tc = 343=C
__...v i R744 Try R744 mass velocity 370
kg/m2 s
R744 discharge temperature Te 80 C
R744 mass flow 2.492 kg/s
R744 design condensing temp. Tc 30 C
Condenser circuits 47
MM =
16 ft 7.2
Length of tube bundle L 5,060 in 0.0% OK
Condenser pressure drop 10 kPa
Number of circuits Nc 72
P
cip Width allowance for header clearance 25 MM
-,
cH; c5r: Width of tube enclosure W 1360 mm
,
k...)
...3
Number of tube passes
Np 8
T¨
r.,
,
21
0.003 ,
,
0 c Tube selection 513 be
..= 1 ,
,
,
c,
Entering air superficial velocity Vo 3.00 m/s
Ambient air dry bulb Ta 35.0 C
Ambient air wet bulb Twb_in 24.0 C
Refrigerant CO2, Tc = 30 C
Water mass velocity at air inlet Gw 3.00 kg/m2 s
Calculated discharge temperature 80.7 C
m2
Water-side fouling allowance if 0.088 K/kW
R744 mass flux 243 kg/m2 :Iv
n
Atmospheric pressure 101.3 kPa abs
Condenser R744 pressure drop 6 kPa y
Needs higher water flow to
5;
uniformly wet tubes (Ref 2, 3.3)
w
o
Plain or finned tube Plain
Condenser exit air dry bulb 28.9 'C 1¨
vi
Plain tube OD D 15.88 mm
Condenser exit air RH 100.0%
=
Tube ID 13.48 mm
Condenser air pressure drop 61 Pa 2
- 4
Total outside surface 145.4 m2
Fan power ( at 70% efficiency) 1.8 kW --.1
0
Air mass velocity Ga 3.46
kg/m2 s
Ambient air relative humidity 41%
Basin water temperature 26.30 C
cio
Water spray into basin m.w_in 20.65 kg/s
Rise above ambient wet bulb 2.3 K
Water makeup required 929 kg/hour
Tube material n316 stainless y I
Thermal conductivity 16 W/m K
Horizontal tube pitch: PH to OD ratio 2.25
Tube horizontal pitch = 2.25 x 15.875
mm 35.72 mm
Tube to tube below: PL to OD ratio 3.90
Tube vertical pitch to tube below =
c4 3.90 x 15.875 mm 61.91 mm
Row-to-row pitch angle a 60.0 degrees
0 C4
0
t..)
o
,-,
u,
Evaporative Condenser Model 2
-4
t..)
,-,
Required rejection 631 kW
oe
o
Calculated rejection 632
Estimate for Condenser circuits
and pressure drop
kW
Rejection 631 kW
_____________________________________________ _
Refrigerant CO2, Tc = 30';C
'`' 1 R744 Try R744 mass velocity 370 kg/m2
s
__,
R744 discharge temperature Te 80 C
R744 mass flow 3.242 kg/s
R744 design condensing temp. Tc 30 C
Condenser circuits 61
I'M =
16 ft 7.2
Length of tube bundle L 5,060 in 0.2% OK
Condenser pressure drop 10 kPa
P
Number of circuits Nc 72
.
r.,
r Width allowance for header clearance 25 MITI
g
,
Width of tube enclosure W 1360 mm
k...) ...3
t \ ) 7-t-=
Number of tube passes Np 8 1-11'
,
--4
.
"
,
0 _ 31
-0.006 .
,
9 c Tube selection sts' tube
,
,
Entering air superficial velocity Vo 4.00 m/s
Ambient air dry bulb Ta 35.0 C
Ambient air wet bulb Twb_in 24.0 C
Refrigerant CO2, Tc = 30 C
Water mass velocity at air inlet Gw 3.00 kg/m2 s
Calculated discharge temperature 80.9 C
m2
Water-side fouling allowance ff 0.088 K/kW
R744 mass flux 316 kg/m2 4:.!
Atmospheric pressure 101.3 kPa abs
Condenser R744 pressure drop 9 kPa ¨H-i
P9-
Needs higher water flow to
uniformly wet tubes (Ref 2, 3.3)
w
o
Plain or or finned tube Plain
Condenser exit air dry bulb 28.8 "C vi
-a,
Plain tube OD D 15.88 mm
Condenser exit air RH 100.1% o
o
Tube ID 13.48 mm
Condenser air pressure drop 102 Pa w
--.1
Total outside surface 145.4 m2
Fan power ( at 70% efficiency) 4.0 kW --.1
0
Air mass velocity Ga 4.61
kg/m2 s
Ambient air relative humidity 41%
Basin water temperature 25.76 C
oe
Water spray into basin m.w_in 20.65 kg/s
Rise above ambient wet bulb 1.8 K
Water makeup required 1221 kg/hour
Tube material 2 316 st.ines iy I
Thermal conductivity 16 W/m K Above Ga of
3.7, water is partially held up on tubes - Ref. 2, 3.3
Horizontal tube pitch: PH to OD ratio 2.25
Tube horizontal pitch = 2.25 x 15.875
mm 35.72 mm
Tube to tube below: PL to OD ratio 3.90
Tube vertical pitch to tube below =
3.90 x 15.875 mm 61.91 mm
Row-to-row pitch angle a 60.0 degrees
Cl)
7-t-=
.
0 c
5;
0
t..)
o
,-,
u,
Evaporative Condenser Model 3
-4
t..)
,-,
cio
Required rejection 756 kW
=
Calculated rejection 756
Estimate for Condenser circuits
and pressure drop
kW
Rejection 756 kW
____________________________________________ _
Refrigerant CO2, Tc = 30';C
.L1 R744 Try R744 mass velocity 370 kg/m2 s
R744 discharge temperature Te 80 C
R744 mass flow 3.884 kg/s
R744 design condensing temp. Tc 30 C
Condenser circuits 74
mm =
16 ft 7.2
Length of tube bundle L 5,060 in -0.2% OK
Condenser pressure drop 10 kPa P
Number of circuits Nc 72
o
r.,
c4
.
Width allowance for header clearance 25 mm
.
,
ili cr'
,
t`..)
...3
u) Width of tube enclosure
W 1360 mm,.z .
Number of tube passes Np 8 F¨*
,
c) 0 31
-0.003 ,
,
----- '-' Tube selection 1 51'8' tube
0
> - 2
"
Entering air superficial velocity Vo 5.00 m/s
Ambient air dry bulb Ta 35.0 C
Ambient air wet bulb Twb in 24.0 C
Refrigerant CO2, Tc = 30 C
Water mass velocity at air inlet Gw 3.00 kg/m2 s
Calculated discharge temperature 80.3 C
m2
1-o
Water-side fouling allowance if 0.088 K/kW
R744 mass flux 378 kg/m2 s n
i-i
Atmospheric pressure 101.3 kPa abs
Condenser R744 pressure drop 13 kPa ----
Needs higher water flow to
w
uniformly wet tubes (Ref 2, 3.3)
=
Plain or or finned tube Plain
Condenser exit air dry bulb 28.5 C u,
-a,
Plain tube OD D 15.88 mm
Condenser exit air RH 100.2%
o
w
Tube ID 13.48 mm
Condenser air pressure drop 152 Pa --.1
--.1
Total outside surface 145.4 m2
Fan power ( at 70% efficiency) 7.5 kW
0
Air mass velocity Ga 5.76
kg/m2 s
Ambient air relative humidity 41%
Basin water temperature 25.39 C
cio
Water spray into basin m.w_in 20.65 kg/s
Rise above ambient wet bulb 1.4 K
Water makeup required 1487 kg/hour
Tube material 2 316 stainless iv
Thermal conductivity 16 W/m K Above Ga of
3.7, water is partially held up on tubes - Ref. 2, 3.3
Horizontal tube pitch: PH to OD ratio 2.25
Tube horizontal pitch = 2.25 x 15.875
mm 35.72 mm
Tube to tube below: PL to OD ratio 3.90
Tube vertical pitch to tube below =
3.90 x 15.875 mm 61.91 mm
Row-to-row pitch angle a 60.0 degrees
Cl)
0 c
0
t..)
o
,¨
u,
,-
-4
Evaporative Condenser Model 4
t..)
,¨
oo
o
Required rejection 485 kW
Calculated rejection 485
Estimate for Condenser circuits
and pressure drop
_____________________________________________ _ kW
Rejection 485 kW
Refrigerant CO2, Tc = 30';C
.L1 R744 Try R744 mass velocity 370 kg/m2 s
R744 discharge temperature Te 80 C
R744 mass flow 2.492 kg/s
R744 design condensing temp. Tc 30 C
Condenser circuits 47
MrT1 =
16 ft 5.3
P
Length of tube bundle L 5,012 in 0.0% OK
Condenser pressure drop 10 kPa .
r.,
r Number of circuits
, Nc 56
,
Width allowance for header clearance 25 nirT1
,
I,
O.
Width of tube enclosure W 1075 mm"
= 0 Number of tube passes Np
8 F¨* ,
,
,
0 c4,
33
0.009 i
>"
= - , Tube selection 1 51'8' tube Nor 1
Entering air superficial velocity Vo 4.00 m/s
Ambient air dry bulb Ta 35.0 C
Ambient air wet bulb Twb in 24.0 C
Refrigerant CO2, Tc = 30 C
Water mass velocity at air inlet Gw 3.00 kg/m2 s
Calculated discharge temperature 80.6 C
1-o
m2
n
Water-side fouling allowance if 0.088 K/kW
R744 mass flux 312 kg/m2 s'-
-
Atmospheric pressure 101.3 kPa abs
Condenser R744 pressure drop 9 kPa
w
Needs higher water flow to
o
uniformly wet wet tubes (Ref 2, 3.3)
u,
-a-,
Plain or finned tube Plain
Condenser exit air dry bulb 28.8 C o
o
Plain tube OD D 15.88 mm
Condenser exit air RH 99.6% w
--4
--4
Tube ID 13.48 mm
Condenser air pressure drop 102 Pa
Total outside surface 112.0 m2
Fan power ( at 70% efficiency) 3.2 kW
0
Air mass velocity Ga 4.61
kg/m2 s
Ambient air relative humidity 41%
Basin water temperature 25.98 00 re
Water spray into basin m.w_in 16.16 kg/s
Rise above ambient wet bulb 2.0 K
Water makeup required 943 kg/hour
Tube material 2 316 stainless iv
Thermal conductivity 16 W/m K Above Ga of
3.7, water is partially held up on tubes - Ref. 2, 3.3
Horizontal tube pitch: PH to OD ratio 2.25
Tube horizontal pitch = 2.25 x 15.875
mm 35.72 mm
Tube to tube below: PL to OD ratio 3.90
Tube vertical pitch to tube below =
3.90 x 15.875 mm 61.91 mm
Row-to-row pitch angle a 60.0 degrees
0 c
5;
0
t..)
o
,-,
u,
Evaporative Condenser Model 5
-4
t..)
,-,
oe
Required rejection 485 kW
o
Calculated rejection 485
Estimate for Condenser circuits
and pressure drop
______________________________________________ _ kW
Rejection 485 kW
Refrigerant CO2, Tc = 30';C '`' 1
R744 Try R744 mass velocity 370 kg/m2 s
__,
R744 discharge temperature Te 80 C
R744 mass flow 2.492 kg/s
R744 design condensing temp. Tc 30 C
Condenser circuits 47
mm =
16 ft
Length of tube bundle L 5,165 11.3 in 0.0% OK
Condenser pressure drop 11 kPa P
Number of circuits Nc 46
.
r Width allowance for header clearance 25 mm
..'
,
Width of tube enclosure W 896 mmc...)
...3
Number of tube passes Np 8 F-11'
,
L..)
.
"
,
`1) 5r8" tube '. I
35 0.002 .
,
,.µ
,.µ
9 K- Tube selection
5 ,
Entering air superficial velocity Vo 5.00 m/s
Ambient air dry bulb Ta 35.0 C
Ambient air wet bulb Twb_in 24.0 C
Refrigerant CO2, Tc = 30 C
Water mass velocity at air inlet Gw 3.00 kg/m2 s
Calculated discharge temperature 80.7 C
m2
Water-side fouling allowance ff 0.088 K/kW
R744 mass flux 380 kg/m2 A
Atmospheric pressure 101.3 kPa abs
Condenser R744 pressure drop 13 kPa .....
P9-
Needs higher water flow to
uniformly wet tubes (Ref 2, 3.3)
w
o
Plain or or finned tube Plain
Condenser exit air dry bulb 28.6 C ul
-a,
Plain tube OD D 15.88 mm
Condenser exit air RH 98.6% =
o
Tube ID 13.48 mm
Condenser air pressure drop 152 Pa t.1
--.1
Total outside surface 94.8 m2
Fan power ( at 70% efficiency) 5.0 kW
0
Air mass velocity Ga 5.76
kg/m2 s
Ambient air relative humidity 41%
Basin water temperature 25.79 00 re
Water spray into basin m.w_in 13.89 kg/s
Rise above ambient wet bulb 1.8 K
Water makeup required 966 kg/hour
Tube material 2 316 stainless iv
Thermal conductivity 16 W/m K Above Ga of
3.7, water is partially held up on tubes - Ref. 2, 3.3
Horizontal tube pitch: PH to OD ratio 2.25
Tube horizontal pitch = 2.25 x 15.875
mm 35.72 mm
Tube to tube below: PL to OD ratio 3.90
Tube vertical pitch to tube below =
c4 3.90 x 15.875 mm 61.91 mm
Row-to-row pitch angle a 60.0 degrees
4=,
0 CID
5;
CA 02947774 2016-11-02
WO 2015/172180 PCT/AU2015/000277
- 35 -
Whilst a number of condenser and process embodiments and models have been
described, it should be appreciated that the condenser and process may be
embodied in
many other forms.
For example, the plenum 13 could be of circular section, whereby the divergent
plenum 40 comprises a conical frustum, or a square to circular frustum-like
prism.
However, such a configuration is less favoured, as it does not promote free
drainage of
water within the condenser.
1 0 In the claims which follow, and in the preceding description, except
where the
context requires otherwise due to express language or necessary implication,
the word
"comprise" and variations such as "comprises" or "comprising" are used in an
inclusive
sense, i.e. to specify the presence of the stated features but not to preclude
the presence
or addition of further features in various embodiments of the condenser and
process as
disclosed herein.
