U.S. patent application number 10/407415 was filed with the patent office on 2004-10-28 for convection towers for air cooled heat exchangers.
Invention is credited to Bharathan, Desikan, Gawlik, Keith M..
Application Number | 20040211184 10/407415 |
Document ID | / |
Family ID | 33298272 |
Filed Date | 2004-10-28 |
United States Patent
Application |
20040211184 |
Kind Code |
A1 |
Bharathan, Desikan ; et
al. |
October 28, 2004 |
Convection towers for air cooled heat exchangers
Abstract
An exemplary heat exchange system includes a first stage
including a first stage heat exchanger and a tower wherein the
first stage heat exchanger exchanges heat between a fluid and air
to thereby heat the air and generate air convection in the tower;
and a second stage including a second stage heat exchanger and a
powerable convection unit wherein the second stage heat exchanger
receives the fluid from the first stage and exchanges heat between
the fluid and air. An exemplary method includes transferring heat
energy from a fluid to air using a first heat exchanger to cause
air convection in a tower; and transferring heat energy from the
fluid to air using a second heat exchanger and a fan. Various other
exemplary systems and methods are also disclosed.
Inventors: |
Bharathan, Desikan; (Arvada,
CO) ; Gawlik, Keith M.; (Boulder, CO) |
Correspondence
Address: |
PAUL J WHITE, SENIOR COUNSEL
NATIONAL RENEWABLE ENERGY LABORATORY (NREL)
1617 COLE BOULEVARD
GOLDEN
CO
80401-3393
US
|
Family ID: |
33298272 |
Appl. No.: |
10/407415 |
Filed: |
April 4, 2003 |
Current U.S.
Class: |
60/651 |
Current CPC
Class: |
F28B 1/06 20130101 |
Class at
Publication: |
060/651 |
International
Class: |
F01K 025/08 |
Goverment Interests
[0001] The United States Government has rights I this invention
pursuant to Contract No. DE-AC36-99GO10337 between the U.S.
Department of Energy and the National Renewable Energy Laboratory,
a division of Midwest Research Institute.
Claims
1. A heat exchange system comprising: a first stage including a
first stage heat exchanger and a tower wherein the first stage heat
exchanger exchanges heat between a fluid and air to thereby heat
the air and generate air convection in the tower; and a second
stage including a second stage heat exchanger and a powerable
convection unit wherein the second stage heat exchanger receives
the fluid from the first stage and exchanges heat between the fluid
and air.
2. The heat exchange system of claim 1, wherein the fluid comprises
a zeotrope.
3. The heat exchange system of claim 1, wherein at least some of
the fluid enters the first stage as a vapor.
4. The heat exchange system of claim 1, wherein the fluid has a
temperature glide for a vapor to liquid phase transition.
5. The heat exchange system of claim 1, wherein the air enters the
first stage at ambient conditions.
6. The heat exchange system of claim 1, wherein the air exits the
tower at a temperature greater than the ambient air
temperature.
7. The heat exchange system of claim 1, wherein the fluid comprises
ammonia and water.
8. The heat exchange system of claim 1, wherein the fluid exits a
turbine prior to entering the first stage.
9. The heat exchange system of claim 1, wherein the first stage and
the second stage condense the fluid as part of a Kalina cycle.
10. The heat exchange system of claim 1, wherein the first stage
and the second stage condense the fluid as part of a cycle selected
from the group consisting of Lorenz cycles, Uehara cycle, Rankine
cycles, Carnot cycles and combinations thereof.
11. The heat exchange system of claim 1, wherein the tower has an
air inlet located at least at the periphery of a cylindrical
section.
12. The heat exchange system of claim 1, wherein the first stage
heat exchanger includes a plurality of heat exchangers.
13. The heat exchange system of claim 1, wherein the first stage
heat exchanger is located near an air inlet.
14. The heat exchange system of claim 1, wherein the first stage
heat exchanger is located near a periphery of a cylindrical section
of the tower.
15. The heat exchange system of claim 1, wherein the first stage
includes a powerable convection unit.
16. The heat exchange system of claim 1, wherein the second stage
includes a tower.
17. The heat exchange system of claim 1, wherein the height of the
tower depends on one or more parameters associated with the second
stage.
18. The heat exchange system of claim 1, wherein the height of the
tower depends on a pinch point.
19. The heat exchange system of claim 1, wherein air enters the
tower substantially horizontally and exits the tower substantially
vertically.
20. The heat exchange system of claim 1, wherein the convection in
the tower comprises natural convection.
21. A method comprising: transferring heat energy from a fluid to
air using a first heat exchanger to cause air convection in a
tower; and transferring heat energy from the fluid to air using a
second heat exchanger and a fan.
22. The method of claim 21, wherein the fluid comprises a
zeotrope.
23. The method of claim 21, wherein the fluid comprises ammonia and
water.
24. The method of claim 21, wherein the first heat exchanger relies
on the air convection in the tower.
25. The method of claim 21, wherein the fluid condenses from a
vapor to a liquid.
26. The method of claim 21, wherein the fluid has a temperature
glide for a vapor to liquid phase transition.
27. The method of claim 21, wherein a controller controls the
fan.
28. The method of claim 21, wherein the air enters the first heat
exchanger at ambient conditions.
29. The method of claim 21, wherein the air exits the tower at a
temperature greater than the ambient air temperature.
30. The method of claim 21, wherein the height of the tower depends
on a pinch point.
31. The method of 21, wherein the first heat exchanger and the
second heat exchanger condense the fluid as part of a Kalina
cycle.
32. The method of claim 21, wherein the first heat exchanger and
the second heat exchanger condense the fluid as part of a cycle
selected from the group consisting of Lorenz cycles, Uehara cycles,
Rankine cycles, Carnot cycles and combinations thereof.
33. The method of claim 21, wherein the air convection in the tower
comprises natural convection.
34. A heat exchange system comprising: means for transferring heat
energy from a fluid to air using only natural convection of the
air; and means for further transferring heat energy from the fluid
to air using a fan.
35. The heat exchange system of claim 34, wherein the means for
transferring heat energy comprises a heat exchanger.
36. The heat exchange system of claim 34, wherein the means for
further transferring heat energy comprises a heat exchanger.
37. A heat exchange system comprising: a first stage including a
first stage heat exchanger and a structural passageway wherein the
first stage heat exchanger exchanges heat between a fluid and air
to thereby heat the air and generate air convection in the
structural passageway; and a second stage including a second stage
heat exchanger and a powerable convection unit wherein the second
stage heat exchanger receives the fluid from the first stage and
exchanges heat between the fluid and air.
38. The heat exchange system of claim 37, wherein the structural
passageway comprises a stairway.
39. The heat exchange system of claim 37, wherein the structural
passageway is located at least partially within a building.
40. The heat exchange system of claim 37, wherein the air
convection in the structural passageway comprises natural
convection.
Description
FIELD OF INVENTION
[0002] The subject matter disclosed herein generally relates to air
cooled heat exchangers and/or air cooled heat exchange systems.
BACKGROUND
[0003] Hot air rises naturally because of its lower density in a
cool atmosphere (e.g., consider a hot air balloon); hence, a
substantial amount of air can be induced to flow naturally upward
in a properly designed flow passage. Natural draft-cooling towers
operate using this principle at many power plants.
[0004] Further, a typical residence may use a chimney to cause flue
gases (from furnaces and fireplaces) to escape into the atmosphere.
The key difference between the residential and power plant
applications is that the chimney for the residence is a lot shorter
and is made to fit naturally within the height of the building.
[0005] Because the temperature of the exhaust air from a furnace is
likely to be high, the shorter height of the chimney for a
residence is generally adequate. However, for a power plant, the
natural draft tower of a large diameter must typically rise to a
height of about 120 m (e.g., about 400 ft).
[0006] A natural draft tower uses the heat contained in the hot air
to effectively induce airflow over a plant's heat rejecting
surfaces. A natural draft tower can eliminate use of fans to induce
airflow and an associated need for parasitic power. However, the
natural draft tower, as already noted, typically needs to be quite
tall. Many sites do not offer the flexibility to use such a tower
on account of general height restrictions on structures and their
imposing visual impact. Further, natural-draft towers are normally
built using concrete (e.g., thin-shell concrete, etc.), which can
be a major cost item for the power plant that increases
substantially with respect to tower height.
[0007] Forced draft heat exchangers are commonly used to avoid
problems related to natural draft towers. Forced draft heat
exchange systems are typically used with outdoor mounted condensing
coils for air conditioners and power plants where they are placed
either vertically or horizontally above the ground with fans either
blowing or inducing airflow over the coils. Forced draft fans are
also used in wet cooling towers. While forced draft units can be
made compact, they require substantial parasitic power to run the
fans. Forced draft units also do not take advantage of the
potentially usable energy in the exhausted hot air.
[0008] With respect to water cooling, water for heat rejection is
generally on the decline, especially in newer construction. Reasons
for the decline in water use for heat rejection include lack of
water availability, very long water-rights acquisition periods, and
adverse impact of visual plumes (clouds) that arise from wet
cooling towers. Indeed, more and more power plants are required to
use air as a heat rejection medium. However, specific heat for air
is quite low in comparison to that of water. Air is also
substantially less dense than water. Hence, a large volume of air
must be induced for heat rejection from power plants. With
air-cooling emerging as a new trend in power plant construction,
effective methods, heat exchangers and/or systems to reduce power
consumption for air-cooling are becoming important in power plant
development and/or other applications.
[0009] Various exemplary heat exchange systems and exemplary
methods presented herein aim to balance aspects of natural
convection tower with aspects of forced draft units. Further,
various exemplary systems and exemplary methods presented herein
aim to reduce parasitic power requirements.
SUMMARY
[0010] An exemplary heat exchange system includes a first stage
including a first stage heat exchanger and a tower wherein the
first stage heat exchanger exchanges heat between a fluid and air
to thereby heat the air and generate air convection in the tower;
and a second stage including a second stage heat exchanger and a
powerable convection unit wherein the second stage heat exchanger
receives the fluid from the first stage and exchanges heat between
the fluid and air. An exemplary method includes transferring heat
energy from a fluid to air using a first heat exchanger to cause
air convection in a tower; and transferring heat energy from the
fluid to air using a second heat exchanger and a fan. Various other
exemplary systems and methods are also disclosed.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] Features and advantages of the described implementations can
be more readily understood by reference to the following
description taken in conjunction with the accompanying
drawings.
[0012] FIG. 1 shows an exemplary two stage heat exchange
system;
[0013] FIG. 2 shows plots of temperature versus energy for a
substantially single component working fluid and a multi-component
working fluid;
[0014] FIG. 3 shows a T-x-y diagram for a multi-component mixture
of ammonia and water;
[0015] FIG. 4 shows a plot of temperature versus energy for an
exemplary two stage heat exchange system that uses a working fluid
having a temperature glide for a vapor to liquid phase
transition;
[0016] FIG. 5 shows an exemplary heat exchange stage that includes
a powerable convection unit;
[0017] FIG. 6 shows an exemplary heat exchange stage that includes
a tower and optionally a powerable convection unit;
[0018] FIG. 7 shows an exemplary tower suitable for use in a heat
exchange stage;
[0019] FIG. 8 shows an exemplary tower suitable for use in a heat
exchange stage; and
[0020] FIG. 9 shows an exemplary two stage heat exchange system
that relies on a structural passageway of a structure or building
to help generate natural convection currents.
DETAILED DESCRIPTION
[0021] The following description includes the best mode presently
contemplated for practicing various described implementations. This
description is not to be taken in a limiting sense, but rather is
made merely for the purpose of describing the general principles of
the various implementations. The scope of the described
implementations should be ascertained with reference to the issued
claims.
[0022] Exemplary Heat Exchange System
[0023] FIG. 1 shows an exemplary two stage heat exchange system
100. A first stage, Stage 1, includes a heat exchanger 110 to
exchange heat between a working fluid and air. In Stage 1, a tower
112 enhances air convection to thereby increase heat transfer
between the working fluid and air at the heat exchanger 110.
[0024] According to Stage 1, a working fluid enters the heat
exchanger 110 at a temperature T.sub.WFS1(In) and air enters the
heat exchanger 110 at a temperature T.sub.AirS1(In) wherein
T.sub.WFS1(In) is greater than T.sub.AirS1(In). At the heat
exchanger 110, heat energy flows from the working fluid to the air.
Depending on the state of the working fluid, the loss of heat
energy normally causes the working fluid to experience at least a
partial change of state (e.g., a phase change) and/or to decrease
in temperature. For example, working fluid exiting the heat
exchanger 110 may have a temperature T.sub.WFS1(Out), which is less
than T.sub.WFS1(In).
[0025] Heat energy transferred to the air normally causes the air
to increase in temperature. Hence, air exiting the heat exchanger
110 will have a temperature T.sub.AirS1(Out), which is greater than
T.sub.AirS1(ln). In instances where the air entering the heat
exchanger is taken from an ambient source, such as, air near the
tower 112, heat energy transferred to the air at the heat exchanger
110 causes the air to expand and become less dense than the ambient
air. In turn, the heated air rises in the tower 112 according to
principles of natural convection. Thus, natural convection currents
enhance heat transfer at the heat exchanger 110.
[0026] A second stage of the two stage heat exchange system 100,
Stage 2, includes a heat exchanger 120 to exchange heat between the
working fluid and air. In Stage 2, an air duct 122 provides a flow
path for air past or through the heat exchanger 120. In this
example, a fan 124 consumes power 126 to propel air or cause air
convection in the duct 122. Air convection caused by power 126
provided to the fan 124 aims to increase heat transfer between the
working fluid and air at the heat exchanger 120.
[0027] According to Stage 2, the working fluid enters the heat
exchanger 120 at a temperature T.sub.WFS2(In) and air enters the
heat exchanger 120 at a temperature T.sub.AirS2(In) wherein
T.sub.WFS2(In) is greater than T.sub.AirS2(In). At the heat
exchanger 120, heat energy flows from the working fluid to the air.
Depending on the state of the working fluid, the loss of heat
energy normally causes the working fluid to experience at least a
partial change of state (e.g., a phase change) and/or to decrease
in temperature. For example, working fluid exiting the heat
exchanger 120 may have a temperature T.sub.WFS2(Out), which is less
than T.sub.WFS2(In).
[0028] Heat energy transferred to the air normally causes the air
to increase in temperature. Hence, air exiting the heat exchanger
120 will have a temperature T.sub.AirS2(Out), which is greater than
T.sub.AirS2(In). In instances where the air entering the heat
exchanger is taken from an ambient source, heat energy transferred
to the air at the heat exchanger 120 causes the air to expand and
become less dense than the ambient air. In turn, the heated air may
rise according to principles of natural convection. Thus, natural
convection currents may enhance heat transfer at the heat exchanger
120. However, in this example, Stage 2 consumes power from the
power source 126 to enhance heat exchange at the heat exchanger
120. In general, such power may be referred to as parasitic power.
Parasitic power demands may depend on properties of the working
fluid and/or working fluid flow requirements. For example, a
requirement may be placed on T.sub.WFS2(Out), which corresponds to
a certain vapor/liquid phase composition of the working fluid at a
given working fluid pressure.
[0029] While the exemplary two stage heat exchange system 100 shows
a single heat exchanger 110 and a single tower 112 in Stage 1, one
or more heat exchangers and/or one or more towers may be used in
alternative examples. Further, while the exemplary two stage heat
exchange system 100 shows a single heat exchanger 120, a single
duct 122, a single fan 124 and a single power source 126 in Stage
2, one or more heat exchanges, one or more ducts, one or more fans,
one or more power sources, and/or one or more towers may be used in
alternative examples. In general, Stage 1 relies predominantly on
natural convection and Stage 2 relies, at least in part, on a power
source.
[0030] FIG. 2 shows a set of temperature versus energy plots 200. A
first temperature versus heat load plot 210 corresponds to a
substantially single component working fluid whereas a second
temperature versus heat load plot 220 corresponds to a
multi-component working fluid. In these plots, heat load is given
in energy per unit time.
[0031] In the plot 210, the segment A-B corresponds to heating a
first phase of the substantially single component working fluid
from a temperature to a phase transition temperature. The energy
input for sensible heating over the segment A-B is approximated by
the equation Q.sub.S-Liq.=m.sub.Liq.C.sub.Liq..DELTA.T.sub.Liq.,
wherein m.sub.Liq. is the mass of the working fluid, C.sub.Liq. is
the specific heat capacity for the first phase, and
.DELTA.T.sub.Liq. is the temperature differential over the segment
A-B.
[0032] In the plot 210, the segment B-C corresponds to heating a
first phase of the substantially single component working fluid to
effectuate a phase transition to a second phase of the
substantially single component working fluid. In this example, the
phase transition occurs at an approximately constant temperature.
The energy input for latent heating over the segment B-C is
approximated by the equation Q.sub.L=m.sub.VL.sub.V, wherein
m.sub.V is the mass of working fluid and L.sub.V is the associated
latent heat energy per unit mass for the phase transition.
[0033] In the plot 210, the segment C-D corresponds to heating a
second phase of the substantially single component working fluid
from a phase transition temperature to a higher temperature. The
energy input for sensible heating over the segment C-D is
approximated by the equation
Q.sub.S-V=m.sub.VC.sub.V.DELTA.T.sub.V, wherein m.sub.V is the mass
of the working fluid, C.sub.V is the specific heat capacity for the
first phase, and .DELTA.T.sub.V is the temperature differential
over the segment C-D. The total energy transferred to the working
fluid over the segment A-D is approximated by the sum of energies
associated with the individual segments:
Q.sub.T=Q.sub.L+Q.sub.S-V+Q.sub.S-Liq.
[0034] As already mentioned, the plot 220 corresponds to a
multi-component working fluid. In general, a suitable
multi-component working fluid includes components that are
miscible. For example, oil and water are typically not miscible
under standard conditions (e.g., standard temperature, pressure,
gravity, etc.). Further, a suitable multi-component working fluid
may be a zeotrope or an azeotrope.
[0035] For a zeotrope, individual component concentrations in a
liquid phase and a corresponding vapor phase are never equal, which
creates a temperature glide during a liquid-vapor or vapor-liquid
phase transition. In essence, a zeotrope is a liquid mixture that
exhibits no maximum or minimum when vapor pressure is plotted
against composition at constant temperature. As a zeotrope vapor is
cooled, liquid formation commences at the dew point temperature and
stops at the bubble point temperature. The difference between the
dew point and bubble point temperatures is known as a temperature
glide. In contrast, the substantially single component working
fluid shown in plot 210 does not exhibit a temperature glide over
the phase transition represented by the segment B-C because the
phase transition occurs at an approximately constant
temperature.
[0036] An azeotrope may be defined as a multi-component mixture at
a point wherein the individual component concentrations of the
liquid phase and the vapor phase are the same for a given
temperature and pressure. An azeotrope typically behaves like a
single component fluid in that a phase transition occurs at an
approximately constant temperature. For example, an ethanol and
water mixture is an azeotrope at approximately 0.89 mole fraction
of ethanol, a pressure of approximately 101 kPa and a temperature
of approximately 78.1.degree. C. Thus, distillation of such a
mixture can never produce a pure ethanol vapor.
[0037] Referring again to FIG. 2, plot 220 shows a temperature
glide .DELTA.T.sub.Glide over a segment B-C. In addition, the plot
220 shows segments A-B and C-D, which correspond to liquid and
vapor heating, respectively. The energy associated with heating the
multi-component working fluid from A-D may still be represented as
the sum of individual sensible and latent energies:
Q.sub.T=Q.sub.L+Q.sub.S-V+Q.sub.S-Liq.
[0038] FIG. 3 shows a plot 300 of temperature versus mass fraction
for an ammonia-water zeotrope at approximately 4 bar and at
approximately 2 bar. Two pressures are shown to illustrate the
effect of increasing pressure on vapor and liquid phase composition
in terms of mass fraction. An exemplary condensation scheme is
shown for an ammonia-water zeotrope having a liquid composition of
approximately 0.35 mass fraction ammonia and approximately 0.65
mass fraction water. Of course, other compositions may be suitable.
For example, an exemplary alternative ammonia-water zeotrope has a
liquid composition of approximately 0.85 mass fraction ammonia and
0.15 mass fraction water. Yet further, other pressures may be
possible, for example, an exemplary alternative uses an
ammonia-water zeotrope at a pressure of approximately 13 bar (e.g.,
approximately 1.3 MPa or approximately 190 psia).
[0039] According to the exemplary condensation scheme, zeotrope
vapor at a pressure of approximately 2 bar and a temperature of
approximately 390 K (approx. 117.degree. C.) (e.g., point D) losses
energy to its surroundings. Upon reaching its dew point (e.g.,
point C), condensation commences wherein vapor and liquid or
condensate concentrations differ. As more energy is lost,
condensate concentration follows a line of bubble points while
vapor concentration follows a line of dew points. At the zeotrope's
bubble point (e.g., point B), the zeotrope becomes a single phase
liquid. During condensation, the segment C-B, zeotrope temperature
decreased according to a temperature glide .DELTA.T.sub.Glide. A
further loss of energy causes the zeotrope to cool as a liquid to a
temperature of approximately 310 K (approx. 37.degree. C.). Thus,
in this example, the overall temperature differential, .DELTA.T, is
approximately 80 K. Of course, other temperature differentials may
be suitable and depend on zeotrope concentrations, pressure, etc.
For example, an exemplary zeotrope operates with a temperature
differential of approximately 30 K and between approximately 343 K
(approx. 70.degree. C.) and approximately 313 K (approx. 40.degree.
C.).
[0040] Zeotropes often exhibit a nonlinear relationship between
enthalpy and temperature whereas a single component typically
exhibits a linear relationship. Consequently, a zeotrope may have a
specific heat capacity that is nonlinear with respect to
temperature. For example, an ammonia-water zeotrope may have a
specific heat capacity that increases nonlinear at temperatures
greater than approximately 100.degree. C. and that becomes more
nonlinear with respect to an increase in the mass fraction of
ammonia.
[0041] FIG. 4 shows an exemplary plot 400 of temperature versus
heat load for a working fluid (dashed line) and air (solid lines).
In the plot 400, heat load is given as energy per unit time. The
information in the plot 400 may be related to the exemplary two
stage heat exchange system 100 of FIG. 1, the exemplary working
fluid behavior shown in plot 220 of FIG. 2 and/or the exemplary
working fluid behavior shown in plot 300 of FIG. 3.
[0042] In Stage 1, a working fluid at a temperature T.sub.WFS1(In)
(e.g., point D) exchanges heat with air at a temperature
T.sub.AirS1(Out). The working fluid cools as a vapor until it
reaches point C, which corresponds to a dew point of the working
fluid. Thus, in this example, at least some condensation occurs
during Stage 1. The working fluid continues to transfer heat energy
to the air until point S.sub.1-2, which coincides with air at a
temperature T.sub.AirS1(In) and working fluid at a temperature
T.sub.WFS1(Out). While the plot 400 shows air heating and working
fluid cooling in a substantially countercurrent manner, actual heat
transfer may occur in other manners depending on heat exchanger
arrangement and/or the number of heat exchangers in Stage 1. The
plot 400, for Stage 1, primarily indicates that a working fluid
heats air to increase air temperature by a temperature differential
.DELTA.T.sub.AirS1.
[0043] Working fluid exiting Stage 1 has a temperature
T.sub.WFS1(Out) (e.g., at point S.sub.1-2). In general, the working
fluid entering Stage 2 has a temperature T.sub.WFS2(In), which is
approximately equal to T.sub.WFS1(Out). In Stage 2, the working
fluid transfers heat energy to air to increase air temperature by a
temperature differential .DELTA.T.sub.AirS2, which is the
difference between T.sub.AirS2(Out) and T.sub.AirS2(In). During
Stage 2, the working fluid continues to condense until reaching
point B, which corresponds to a bubble point of the working fluid.
Thereafter, cooling of the working fluid continues until reaching a
temperature T.sub.WFS2(Out) (e.g., point A). In Stage 2, at least
some parasitic power is used to enhance heat transfer between the
working fluid and the air. For example, as shown in FIG. 1, a power
source 126 supplies power to a fan 124 that drives air in a duct
122 whereby the air passes near or through a heat exchanger
120.
[0044] FIG. 5 shows an axisymmetric cross-sectional view of an
exemplary Stage 2 heat exchange system 500 wherein a central axis
coincides with a z-dimension. The system 500 includes a duct 510, a
heat exchanger 520, a powerable convection unit 530 and a power
source 535 that can supply power to the convection unit 530. The
duct 510 has an air inlet 512 at a position r.sub.i and having a
height .DELTA.z.sub.i and an air outlet 514 at a position z.sub.o
and having a radius r.sub.o. The heat exchanger 520 receives
working fluid at a temperature T.sub.WFS2(In) and cools the working
fluid to an exit temperature T.sub.WFS2(Out) using air provided at
a temperature of T.sub.AirS2(In). Heated air exits the heat
exchanger 520 at a temperature T.sub.AirS2(Out). The heated air
flows in the duct 510, through the powerable convection unit 530
and exits the duct air outlet 514.
[0045] A trial was conducted using a system that simulated and
approximated a heat exchange system such as the exemplary Stage 2
heat exchange system 500. To simulate a heat exchanger, a circular
unit heater (Young Radiator Inc., manufacturer, Model AV-60S) was
used. The heater consisted of a set of three parallel 16-mm
(0.625-inch) copper coils fitted with aluminum fins. The coils were
wound into a circular form. Inlet and outlet manifolds for the
coils were nominal 25-mm (1-inch) steel pipes. Air was drawn
circumferentially through the coil by an axial fan located
centrally. Portions were masked circumferentially to assure air
flowed through the fins and there was no other leakage or bypass
paths for the airflow. The exhaust nozzle was 356 mm (14 inch) in
diameter.
[0046] FIG. 6 shows an axisymmetric cross-sectional view of an
exemplary Stage 1 heat exchange system 600 wherein a central axis
coincides with a z-dimension. The system 600 includes a duct 610, a
heat exchanger 620, an optional powerable convection unit 630 and
an optional power source 635 that can supply power to the optional
convection unit 630. The duct 610 has an air inlet 612 at a
position r.sub.i and having a height .DELTA.z.sub.i and an air
outlet 614 at a position z.sub.0 and having a radius r.sub.o. The
heat exchanger 620 receives working fluid at a temperature
T.sub.WFS1(In) and cools the working fluid to an exit temperature
T.sub.WFS1(Out) using air provided at a temperature of
T.sub.AirS1(In). Heated air exits the heat exchanger 620 at a
temperature T.sub.AirS1(Out). The heated air flows in the duct 610,
through the powerable convection unit 630, if provided, and exits
the duct the air outlet 614. In this example, the duct air outlet
614 connects to a tower 640. The tower 640 receives heated air at
the duct outlet 614 and expels the heated air at a tower air outlet
644. The tower 640 has a tower height .DELTA.z.sub.To and a tower
air outlet radius of r.sub.To. The tower air outlet radius r.sub.To
may differ from the duct outlet radius r.sub.o.
[0047] A trial was conducted wherein the aforementioned system that
simulated and approximated a heat exchange system was fitted with
an insulated flexible duct having a diameter of approximately 356
mm (14 inch diameter), which received air from the exhaust nozzle
and simulated a tower.
[0048] The simulated tower duct was hung vertically from the
ceiling of a building. Trials were conducted using various heights
for this simulated tower (e.g., .DELTA.z.sub.TO).
[0049] In the trials, steam at atmospheric pressure was used to
provide heat to the coils. The heater was tested with and without
use of the fan to induce airflow. For tests without the fan, the
fan blades were removed from the motor shaft. The electric motor
powering the fan was rated at nominally 100 W and rotated at speed
of 1090 rpm.
[0050] Instrumentation for the trials included type-K thermocouples
for the air temperature measurements and an inclined tube manometer
for the measuring airside pressure drops. Barometric pressures were
also noted during the trials. The rejected heat rate was inferred
by collecting and weighing steam condensate over measured elapsed
period of time using a stopwatch.
[0051] During trials, air entered the system radially near the
bottom and then traveled upward through a plastic transition
section to the simulated tower. The simulated tower included outer
insulation having Mylar reflective film on outer surface of the
insulation.
[0052] Infrared images were taken of the system and an exiting air
plume. To visualize better the plume, a sheet of drawing paper was
placed in the path of the exiting air plume. A colored temperature
field was indicative of the velocity profile of the exiting air.
During operation, the heater was the hottest and brightest area in
the infrared images and was held at a temperature of approximately
95.degree. C.
[0053] Measured temperature rise in the air together with the
rejected heat rate were used in calculations to derive a value for
airflow. These data were then converted to variations of pressure
loss and heat transfer coefficients as functions of airflow.
[0054] Measured pressure loss (e.g., in Pa) across the coil as a
function of the incoming air volumetric flow rate (e.g., in
m.sup.3/s) was plotted. The plot showed that pressure loss
increased substantially monotonically with increasing flow. A
comparison to calculated variations in pressure loss with respect
to flow rate, based on laminar flow between parallel fins, showed
that the trial measurements were quite close to the calculated
values.
[0055] Measured overall heat transfer coefficient (e.g., in
W/m.sup.2K) was plotted as a function of incoming air volumetric
flow rate (e.g., in m.sup.3/s). The coefficient was based on the
actual heat transfer area available, rather than the tube external
surface area. The measured values fell between approximately 60
W/m.sup.2 K and approximately 90 W/m.sup.2K. Accordingly, the heat
transfer coefficient increases with increasing flow rate. At low
flow rates, a laminar flow limit (for constant temperature boundary
condition) occurred near 50 W/m.sup.2K for flow rates to
approximately 0.8 m.sup.3/s (corresponding to a Reynolds number of
2000 based on fin spacing). A corresponding turbulent limit was
determined at the higher flow rates, based on Dittus-Boelter
empirical correlation. Both predictions were for fully developed
thermal boundary layers.
[0056] The trials showed that for the tested fins with a
length-to-hydraulic-diameter ratio of approximately 11, flow does
not develop fully. Consequently, under such circumstances, the
measured heat transfer coefficients tend to be larger than those
predicted, as borne out by the trial data. The trials helped to
establish a set of base line data suitable for use in determining
flow and heat transfer performance for air-cooled heat
exchangers.
[0057] Further trials used a computational fluid dynamics (CFD)
model of an exemplary heat exchange system wherein tower heights
were varied. For the CFD model, commercially available software was
used (FLUENT.RTM., Aavid Thermal Technologies, Inc., New
Hampshire). All models used axisymmetric flow to reduce model
complexity. Flow geometries and meshes were generated using
accompanying software. Proper boundaries including the symmetry
axis, walls, the heater and fan areas were identified at mesh
generation. The mesh was then imported into the software program.
Airflow was modeled using ideal gas theory. Buoyancy forces were
accounted for by imposing gravitational forces along a downward
direction parallel to the axis of any particular exemplary
tower.
[0058] The CFD software yielded velocity profiles that were induced
in the neighborhood of the heat exchanger. Three different velocity
profiles corresponding to three different tower section heights
indicated that maximum velocity increases with increased tower
section height. In particular, tower heights of approximately 2.5
m, 4.4 m, and 5.3 m were used wherein velocity magnitude varied
from 0 m/s (e.g., at walls and the bottom) to approximately 10 m/s
at the tower air outlet.
[0059] Exemplary Power Plant Cooling System
[0060] As described herein, a geothermal power plant is used to
demonstrate various exemplary heat exchange systems and, in
particular, an exemplary two stage heat exchange system such as the
exemplary system 100 of FIG. 1. First, a geothermal power plant is
chosen for evaluation. Next, the plant is evaluated for a single
stage heat exchange system without a tower. Then, the plant is
evaluated for an exemplary single stage heat exchange system with a
tower. Finally, the plant is used to demonstrate an exemplary two
stage heat exchange system.
[0061] A nominal 12.5 MW net geothermal power plant was evaluated
for the use of an exemplary two stage heat exchange system. In this
evaluation, the plant included a bottoming binary power system that
was added to an existing flash power plant, such as at the Blundell
geothermal power plant site in Utah. Many geothermal power plants,
on account of the low temperature resource when compared to fossil
plants, reject about 85% to 90% of incoming heat. Therefore, a
substantial heat rejection system is required. Normally such a heat
rejection system occupies a considerable portion of the plant
site.
[0062] For this evaluation, the bottoming plant uses a
non-azeotropic mixed working fluid (e.g., a zeotrope). For example,
the bottoming plant may use a mixture of ammonia and water as a
working fluid for the power system and may use an air-cooled heat
rejection system. In general, ammonia-water working fluids are
suitable for plants that rely on Kalina power cycles. When compared
to a Carnot cycle (see, e.g., discussion of most single component
working fluids), a non-azeotropic working fluid in a Kalina power
cycle vaporizes and condenses at varying temperatures (e.g.,
according to a temperature glide) to result in increased work
yield. Other cycles that may be used include Rankine cycles, Uehara
cycles and Lorenz cycles.
[0063] According to the evaluation plant, an ammonia-water zeotrope
(e.g., approximately 84.62% ammonia by mass) enters a heat
exchanger at a flow rate of approximately 98 kg/s, at a temperature
of approximately 66.degree. C., a pressure of approximately 1.3 MPa
and a vapor fraction of approximately 0.67. Ambient air enters the
heat exchanger at a flow rate of approximately 5720 kg/s, a
temperature of approximately 32.degree. C. and a pressure of
approximately 81 kPa. The air cools and condenses the working fluid
to a temperature of approximately 40.degree. C. In turn, the air
heats to a temperature of approximately 47.degree. C. In this
example, approximately 86 MW of power is rejected (e.g., heat
energy per unit time).
[0064] In this example, the one stage heat exchange system causes
an airside pressure loss of approximately 140 Pa. For an assumed
overall fan efficiency of 0.54, the required fan parasitic power
for inducing the associated airflow is approximately 1.56 MW. This
amount of power represents a loss of 12.5% of the net power on
fans.
[0065] To reduce this parasitic power, an exemplary one stage heat
exchange system with a tower is implemented. Based on modeling
results, the evaluation plant requires a heat exchange system
having an overall UA of approximately 11.times.10.sup.6 W/K.
Assuming a nominal air-side heat transfer coefficient (based on
bare tube area) of approximately 625 W/m.sup.2 K, the tube area
requirement is approximately 17,600 m.sup.2. Based on data provided
in various sources, a conventional air-cooled condenser with the
fans has a cost of approximately $3.2 million (equipment
cost--non-installed cost). This cost may represent nearly one third
of the total equipment cost wherein the overall installed plant
cost is estimated to be approximately $28 million.
[0066] For purposes of demonstrating an exemplary one stage heat
exchange system having a tower, the evaluation plant is modeled
after the Blundell Geothermal Power Plant located approximately 300
km (e.g., 250 miles) south of Salt Lake City, Utah. The plant has a
dual flash power system with a nominal 23.5 MW gross output. On
account of potential precipitation from the exit brine, the plant
discharges brine at a temperature of approximately 170.degree. C.
The plant is situated on a gradually sloping ground with
neighboring mountains approximately 1 km (e.g., 0.5 mile) to the
east. The neighboring mountain range rises to a height of about 150
m (e.g., 500 ft) above the plant. Based on the terrain of the
evaluation plant, a tower height of approximately 60 m was chosen
as to be not too obtrusive and to still provide a reasonable
height.
[0067] Based on the inlet and exit air temperatures to the heat
exchanger, such a tower could provide a pressure difference of
approximately 30 Pa (0.12 inch of water). The potential parasitic
fan power reduction on account of the tower would be nominally
one-fourth of its value without the tower. This savings correspond
to 375 kW.
[0068] Of course, a taller tower and/or a hotter air exit
temperature would increase the potential benefit of such an
exemplary one stage heat exchange system having a tower, with the
power reduction being directly proportional to the product of these
variables.
[0069] FIG. 7 shows an exemplary tower 700 optionally suitable for
use in the exemplary one stage heat exchange system and/or various
other exemplary heat exchange systems described herein. To
demonstrate use of the exemplary tower, a heat exchanger (or the
condenser) was sized with a nominal inlet superficial air velocity
of approximately 2 m/s (e.g., at a tower air inlet 712) and a tower
exit velocity of approximately 5 m/s (e.g., at a tower air outlet
714). To handle 5610 m.sup.3/s of inlet air volumetric flow, the
heat exchanger frontal area is assumed to be approximately 2805
m.sup.2. In this example, the heat exchanger was arranged around
the periphery of a substantially cylindrical air inlet section 712
of the exemplary tower 700. Given the requirements, the heat
exchanger overall diameter and height were approximately 75 m and
approximately 12 m, respectively. For this example, the height of
the exemplary tower 700 was approximately 60 m and the exit
diameter of air outlet 714 was approximately 40 m, which resulted
in an exit air velocity of approximately 4.5 m/s.
[0070] The exemplary tower 700 optionally includes a flow guide
718, which has a varying radius that is larger near the bottom of
the tower (e.g., near the air inlet 712) and smaller at axial
positions the air inlet 712. The flow guide 718 may form an
internal column along the axis of the tower 700. In one example,
the flow guide 718 forms a column having a radius of approximately
several meters. The exemplary tower optionally includes a powerable
convection unit 730, for example, a unit that includes fans to help
draw air into the air inlet 712.
[0071] In one example, the powerable convection unit 730 includes
fans to induce the airflow positioned at a height of approximately
20 m (e.g., about one-third the tower height). To induce the
required airflow, six fans were arranged peripherally around the
axis of the tower. For example, fans such as those available from
Hudson Air Products, Inc., having a diameter of approximately 12 m.
To provide access to associated fan machinery, the exemplary tower
700 has a tower radius of approximately 18 m at the height of the
convection unit 730.
[0072] Overall, in this particular example, the exemplary tower 700
has a total height of approximately 60 m and effective height of
approximately 55 m, which can induce a buoyancy pressure
differential of approximately 24 Pa. The pressure differential aids
convection and hence reduces powerable convection unit
requirements. In this example, the parasitic power loss for the
powerable convection unit 730 (e.g., the fans) is approximately
1291 kW, which results in a power savings of 270 kW compared to a
heat exchange system without a tower.
[0073] While the exemplary tower 700 has a substantially
cylindrical tower section, polygonal tower sections are also
possible (e.g., 6-sided, 12-sided, etc.) and may ease construction
and reduce costs. Construction material is optionally concrete
(e.g., thin-shell, etc.) and/or other material.
[0074] As already mentioned, the evaluation plant is suitable for
demonstrating an exemplary two stage heat exchange system, such as,
but not limited to, the exemplary heat exchange system 100 of FIG.
1. In essence, the two stage approach splits the cooling
requirements into a first stage and a second stage. At times, the
first stage is referred to as a topping condenser and the second
stage is referred to as a bottoming condenser. The first stage aims
to take advantage of the highest post-expansion temperature of the
working fluid. For example, if the working fluid is exiting a
turbine wherein it expands to drive the turbine, residual heat
energy in the working fluid is at its highest value just after
exiting the turbine. Of course, due to plumbing and/or other
considerations, some heat energy may be lost prior to contacting
air in a heat exchange system having a tower.
[0075] According to an exemplary method, airflow to a first stage
of an exemplary heat exchange system is adjusted to allow the air
temperature to reach a maximum given a pinch point constraint
(e.g., accounting for second stage requirements and/or operational
parameters). Of course, a natural equilibrium may be reached
wherein the convection forces balance heat transfer to the air. For
example, an increase in air temperature causes an increase in
convection, which, in turn, may increase air flow past or through a
heat exchanger. The increase in air flow past or through the heat
exchanger may aid cooling of a working fluid; however, it will
decrease residence time of the air and hence result in a lower air
temperature. In turn, the lower air temperature will cause a
decrease in natural convection.
[0076] Various exemplary methods described herein may be
implemented using an exemplary controller or control system. In
general, a controller allows for control of parameters germane to
heat exchange (e.g., pressure drop, working fluid flow rate,
working fluid pressure, air flow rate, air pressure, parasitic
power where applicable, etc.).
[0077] An exemplary first stage includes a heat exchanger
positioned around the periphery of an exemplary tower wherein air
convection occurs due to natural convection. An alternative
exemplary first stage includes a powerable convection unit, which
may be used continuously and/or intermittently. Such a powerable
convection unit is optionally used for control and/or for emergency
cooling. Where an exemplary first stage does not include a
powerable convection unit (e.g., a fan, etc.) or wherein no power
flows to a powerable convection unit, there is no
convection-related parasitic power loss associated with transfer of
heat from a working fluid to air.
[0078] According to an exemplary two stage heat exchange system,
working fluid exiting the first stage is received by the second
stage (e.g., by a bottoming condenser, etc.). Of course, various
exemplary schemes are possible whereby working fluid condensed in
the first stage is diverted and only remaining vapor fed to the
second stage.
[0079] In this particular example, the second stage uses a
powerable convection unit (e.g., one or more fans, etc.) that
consumes power (i.e., parasitic power) to force air past or through
one or more heat exchangers. In this manner, the highest
temperature working fluid is used in a stage to generate natural
convection currents that enhance heat transfer while a lower
temperature working fluid is cooled using at least some parasitic
power. Again, in an exemplary multistage heat exchange system,
constraints may exist as to pinch points or other parameters
germane to operation of any individual stage or an entire power
plant.
[0080] Referring again to the evaluation plant examples, an
exemplary two stage heat exchange system has a first stage UA
requirement of approximately 4.3 MW/K and a second stage UA
requirement of approximately 7.5 MW/K. The (UA) requirement
translates to an estimated pressure loss for each stage:
approximately 52 Pa for the first stage and approximately 93 Pa for
the second stage. In this example, the exemplary two stage heat
exchange system results in a parasitic power reduction of
approximately 1060 kW, which may be associated with operation of
the second stage (e.g., a bottoming condenser, etc.). As such, the
exemplary two stage heat exchange system results in a saving of
approximately 500 kW.
[0081] Referring to FIG. 4, a similar plot exists for this
exemplary two stage heat exchange system. For example, ambient air
at a temperature of approximately 32.degree. C. enters a stage one
heat exchanger and is heated by a fluid to a temperature of
approximately 60.degree. C. The heated air rises in a tower, which,
in turn, draws more ambient air into stage one. During this stage
one air heating process, the fluid temperature decreases from
approximately 67.degree. C. to approximately 47.degree. C. The
associated heat duty decreases from approximately 87,000 kW to
approximately 50,000 kW. For the second stage, ambient air enters
at a temperature of approximately 32.degree. C. and is heated by
the fluid exiting stage one to a temperature of approximately
40.degree. C. During this stage two air heating process, the fluid
temperature decreases from approximately 47.degree. C. to
approximately 40.degree. C. The associated heat duty decreases from
approximately 50,000 kW to approximately 0 kW. Depending on
pressure and/or composition of the fluid, vapor cooling,
condensation and/or liquid cooling occurs during stage one and/or
stage two. For example, stage one may include vapor cooling to a
dew point followed by condensation along a temperature glide while
stage two may include condensation along a temperature glide
followed by liquid cooling.
[0082] Again, in this example, the total UA requirement for the two
stages is approximately 11.7 MW/K, which is only slightly larger
than the baseline requirement of 11 MW/K. The added area
requirement in this example may be accounted for in an economic
evaluation.
[0083] Overall, various aforementioned examples demonstrate utility
of a multi-stage approach to cooling a working fluid. In
particular, several examples demonstrate that high temperature,
post-expansion working fluid is useful to generate natural
convection to thereby reduce parasitic power requirements. Yet
further, an exemplary two stage heat exchange system that includes
powered convection in a second stage and natural convection in a
first stage demonstrates reduced parasitic power requirements when
compared to a single stage powered convection heat exchange
system.
[0084] Exemplary Tower
[0085] FIG. 8 shows a rendered view cut open to show the internal
features of an exemplary tower 800. The exemplary tower 800 has a
sheath 810 that is supported internally and/or externally by a
support system 816. The tower 800 includes an air inlet 812 and an
air outlet 814. In a specific example, the exemplary tower 800
includes 12 vertical panels, 12 radial cables to hold the structure
outward, connected by horizontal connector cables to ring cables
that are spaced apart vertically. The panels (e.g., fabric or other
material) may form a flow channel and have cables in cuffs along
vertical edges. In this example, the edges are shaped to connect to
the cable net at the intersection of ring cables and horizontal
connector cables. The panels are optionally attached with
industrial loop and hook attachments that can provide an airtight,
smooth, continuously curved surface to enclose the air passage. The
horizontal cables may permit accurate shaping of the fabric to
follow the desired shape.
[0086] In this example, the edge supports consist of two sloping
pipes that carry the load of the tower to the ground. The central
mast includes a vertical section of space frame anchored at the
bottom. Twelve stabilizing cables anchor the top ring to the mast
foundation. The weight of the exemplary tower 800 is minimal
compared to conventional cooling towers while the exemplary tower
800 has integrity sufficient to withstand cross winds.
[0087] Exemplary Multi-stage Heat Exchange for
Buildings/Structures
[0088] Many tall buildings have the potential to make use of their
natural height to help minimize the fan power used for heat
rejection. Internal heat load from a building usually ends up as
rejected condenser heat (along with the compressor work) from air
conditioners. Ambient airflow is induced through the condenser coil
using fans to carry the heat. In a typical application, hot air
from a heat rejection system may exit with an air temperature rise
of approximately 30.degree. C. to approximately 45.degree. C., and
a corresponding decrease in density of approximately 10% to
approximately 15%. Using an average value of 12% change in density,
the buoyancy of hot air causes a potential motive pressure
difference of about 1.44 Pascal per meter draft tower height. If an
air cooler causes a pressure loss of approximately 125 Pa, the
entire pressure loss can be overcome using a tower of approximately
87 m height. Thus, a tall building of 25 stories may induce all the
airflow using natural convection. Of course, a shorter building
will induce a lesser pressure difference for the airflow.
[0089] One suitable place to accommodate a draft tower is in a
stairwell of a building. Stairwells are required for most all
buildings to meet the safety related regulations. For example, an
exemplary structural passageway is formed as a central clear column
in a stairwell with stairs located around the periphery of the
column. Such a column can be used as a draft tower to allow hot air
to rise and thereby generate natural convection currents. Of
course, hot air should be generated from heat exchange equipment
located below, at or near the ground level of a building and
directed to the bottom of such a structural passageway.
[0090] At early stages of planning, adequate provisions may be made
to accommodate a structural passageway that acts as a draft tower
and to provide for reasonable access to heat-rejection equipment
for maintenance purposes. Whether the added costs related to such a
structural passageway balances potential cost savings projected for
fan use is likely to be application dependent.
[0091] FIG. 9 shows an exemplary two stage heat exchange system 900
in a structure 902 (e.g., a building, etc.). A first stage, Stage
1, includes a heat exchanger 910 to exchange heat between a working
fluid and air. In Stage 1, a structural passageway 912 enhances air
convection to thereby increase heat transfer between the working
fluid and air at the heat exchanger 910.
[0092] According to Stage 1, a working fluid enters the heat
exchanger 910 at a temperature T.sub.WFS1(In) and air enters the
heat exchanger 910 at a temperature T.sub.AirS1(In) wherein
T.sub.WFS1(In) is greater than T.sub.AirS1(In). At the heat
exchanger 910, heat energy flows from the working fluid to the air.
Depending on the state of the working fluid, the loss of heat
energy normally causes the working fluid to experience at least a
partial change of state (e.g., a phase change) and/or to decrease
in temperature. For example, working fluid exiting the heat
exchanger 910 may have a temperature T.sub.WFS1(Out), which is less
than T.sub.WFS1(In).
[0093] Heat energy transferred to the air normally causes the air
to increase in temperature. Hence, air exiting the heat exchanger
910 will have a temperature T.sub.AirS1(Out), which is greater than
T.sub.AirS1(In). In instances where the air entering the heat
exchanger is taken from an ambient source, such as, air near the
structural passageway 912, heat energy transferred to the air at
the heat exchanger 910 causes the air to expand and become less
dense than the ambient air. In turn, the heated air rises in the
structural passageway 912 according to principles of natural
convection. Thus, natural convection currents enhance heat transfer
at the heat exchanger 910.
[0094] A second stage of the two stage heat exchange system 900,
Stage 2, includes a heat exchanger 920 to exchange heat between the
working fluid and air. In Stage 2, an air duct 922 provides a flow
path for air past or through the heat exchanger 920. In this
example, a fan 924 consumes power 926 to propel air or cause air
convection in the duct 922. Air convection caused by power 926
provided to the fan 924 aims to increase heat transfer between the
working fluid and air at the heat exchanger 920.
[0095] According to Stage 2, the working fluid enters the heat
exchanger 920 at a temperature T.sub.WFS2(In) and air enters the
heat exchanger 920 at a temperature T.sub.AirS2(In) wherein
T.sub.WFS2(In) is greater than T.sub.AirS2(In). At the heat
exchanger 920, heat energy flows from the working fluid to the air.
Depending on the state of the working fluid, the loss of heat
energy normally causes the working fluid to experience at least a
partial change of state (e.g., a phase change) and/or to decrease
in temperature. For example, working fluid exiting the heat
exchanger 920 may have a temperature T.sub.WFS2(Out), which is less
than T.sub.WFS2(In).
[0096] Heat energy transferred to the air normally causes the air
to increase in temperature. Hence, air exiting the heat exchanger
920 will have a temperature T.sub.AirS2(Out), which is greater than
T.sub.AirS2(In). In instances where the air entering the heat
exchanger is taken from an ambient source, heat energy transferred
to the air at the heat exchanger 920 causes the air to expand and
become less dense than the ambient air. In turn, the heated air may
rise according to principles of natural convection. Thus, natural
convection currents may enhance heat transfer at the heat exchanger
920. However, in this example, Stage 2 consumes power from the
power source 926 to enhance heat exchange at the heat exchanger
920. In general, such power may be referred to as parasitic power.
Parasitic power demands may depend on properties of the working
fluid and/or working fluid flow requirements. For example, a
requirement may be placed on T.sub.WFS2(Out), which corresponds to
a certain vapor/liquid phase composition of the working fluid at a
given working fluid pressure.
[0097] While the exemplary two stage heat exchange system 900 shows
a single heat exchanger 910 and a single structural passageway 912
in Stage 1, one or more heat exchangers and/or one or more
structural passageways may be used in alternative examples.
Further, while the exemplary two stage heat exchange system 900
shows a single heat exchanger 920, a single duct 922, a single fan
924 and a single power source 926 in Stage 2, one or more heat
exchanges, one or more ducts, one or more fans, one or more power
sources, and/or one or more structural passageways may be used in
alternative examples. In general, Stage 1 relies predominantly on
natural convection and Stage 2 relies, at least in part, on a power
source.
[0098] Various exemplary heat exchange systems presented herein use
a hot air tower or a structural passageway, optionally in
conjunction with forced draft fans or the like, to reduce use of
parasitic power. An exemplary system that combines use of a
powerable convection unit (e.g., fan, etc.) and a tower or
structural passageway, can reduce the required height of the tower
or a structural passageway to a level that might be suitable for
any application.
[0099] Various aforementioned examples consider an evaluation plant
to demonstrate usefulness of an exemplary two stage heat exchange
system. As described, a part-load tower may provide benefits at a
reasonable cost that is considerably less than a conventional
full-load tower. In one particular example, a part-load tower
handles only a partial load and aims at achieving a maximum
temperature for cooling air (e.g., considering constraints, etc.).
Such a part-load tower optionally uses 22% of the airflow of a
full-load tower.
[0100] In addition, an exemplary two stage heat exchange system
approach is optionally suitable for implementation in conjunction
with a structure, such as a building, wherein one or more
structural passageways serve as natural convection routes that aid
cooling.
* * * * *