U.S. patent application number 13/209944 was filed with the patent office on 2013-02-21 for transferring heat between fluids.
The applicant listed for this patent is Barry R. Cole, Yunho Hwang, Hoseong Lee, Jan Muehlbaurer, Laurence Jay Shapiro. Invention is credited to Barry R. Cole, Yunho Hwang, Hoseong Lee, Jan Muehlbaurer, Laurence Jay Shapiro.
Application Number | 20130042996 13/209944 |
Document ID | / |
Family ID | 47711791 |
Filed Date | 2013-02-21 |
United States Patent
Application |
20130042996 |
Kind Code |
A1 |
Hwang; Yunho ; et
al. |
February 21, 2013 |
TRANSFERRING HEAT BETWEEN FLUIDS
Abstract
Heat exchangers can include heat exchange plates with: front and
back exterior surfaces exposed to a non-working fluid; an interior
working fluid flow channel between the front and back exterior
surfaces, comprising a first plurality of parallel flow paths in a
first direction and a second plurality of parallel flow paths in a
second direction.
Inventors: |
Hwang; Yunho; (Ellicott
City, MD) ; Muehlbaurer; Jan; (Bowie, MD) ;
Lee; Hoseong; (College Park, MD) ; Shapiro; Laurence
Jay; (Fair Lawn, NJ) ; Cole; Barry R.;
(Mineral, VA) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Hwang; Yunho
Muehlbaurer; Jan
Lee; Hoseong
Shapiro; Laurence Jay
Cole; Barry R. |
Ellicott City
Bowie
College Park
Fair Lawn
Mineral |
MD
MD
MD
NJ
VA |
US
US
US
US
US |
|
|
Family ID: |
47711791 |
Appl. No.: |
13/209944 |
Filed: |
August 15, 2011 |
Current U.S.
Class: |
165/45 ;
165/170 |
Current CPC
Class: |
F24V 50/00 20180501;
F28D 9/0037 20130101; F28F 3/14 20130101; Y02E 10/30 20130101 |
Class at
Publication: |
165/45 ;
165/170 |
International
Class: |
F28F 3/12 20060101
F28F003/12; F24J 3/08 20060101 F24J003/08 |
Claims
1. A heat exchange plate comprising: front and back exterior
surfaces exposed to a non-working fluid; and an interior working
fluid flow channel between the front and back exterior surfaces,
comprising a first plurality of parallel flow paths in a first
direction and a second plurality of parallel flow paths in a second
direction.
2. The heat exchange plate of claim 1, wherein the first direction
is opposite and parallel to the second direction.
3. The heat exchange plate of claim 1 wherein the first and second
directions of the first and second flow paths are perpendicular to
the direction of flow of the non-working fluid.
4. The heat exchange plate of claim 1 wherein the plate further
comprises a first region of relatively high working fluid mass flux
when a working fluid has a low vapor quality and a second region of
relatively low working fluid mass flux when the working fluid has a
high vapor quality.
5. The heat exchange plate of claim 1 wherein the interior flow
channel has an area of varying space in fluid contact with the
first plurality of parallel flow paths and the second plurality of
parallel flow paths.
6. The heat exchange plate of claim 5 further comprising one or
more structural walls within the first and second plurality of flow
paths, the one or more walls being generally parallel to the flow
path and terminating in the area of varying space.
7. The heat exchange plate of claim 6 wherein the one or more
structural walls comprises directional vane at the terminal end and
oriented in the direction of flow.
8. The heat exchange plate of claim 6 wherein the one or more
structural walls comprises a directional vane at the proximal end
of the structural wall.
9. The heat exchange plate of claim 1 wherein a flow path of the
first and second plurality of flow paths comprises a void having a
cross sectional area of between 155 mm and 60 mm.
10. The heat exchange plate of claim 1 wherein the heat exchange
plate is a composite blow molded plate.
11. The heat exchange plate of claim 1 wherein the heat exchange
plate is aluminum.
12. The heat exchange plate of claim 1 wherein the working fluid
pressure drop across the plate is about 0.2 psi/ft.
13. The heat exchange plate of claim 1 wherein the non-working
fluid heat transfer coefficient ranges from 900 to 1400 Btu/ft2
Rhr.
14. The heat exchange plate of claim 1 wherein the pattern of a
first plurality of parallel flow paths in a first direction and a
second plurality of parallel flow paths in a second direction
repeats across the length of the heat exchange plate.
15. The heat exchange plate of claim 14 wherein the number of flow
paths in the first and second plurality of flow paths increases as
the pattern repeats across the length of the heat exchange
plate.
16. The heat exchange plate of claim 14 wherein the number of flow
paths in the first plurality of flow paths increases from four flow
paths per first direction to six flow paths per first
direction.
17. The heat exchange plate of claim 14 wherein the number of flow
paths in the first plurality of flow paths increases from two flow
paths per first direction to four flow paths per first
direction.
18. The heat exchange plate of claim 1 wherein the non-working
fluid is sea water.
19. The heat exchange plate of claim 1 wherein the working fluid is
ammonia.
20. The heat exchange plate of claim 1 wherein the plate is an OTEC
heat exchange plate.
Description
TECHNICAL FIELD
[0001] This invention relates to transferring heat between fluids
and, more specifically, to transferring heat between fluids using
heat exchange plates.
BACKGROUND
[0002] Energy consumption and demand throughout the world has grown
at an exponential rate. This demand is expected to continue to
rise, particularly in developing countries in Asia and Latin
America. At the same time, traditional sources of energy, namely
fossil fuels, are being depleted at an accelerating rate and the
cost of exploiting fossil fuels continues to rise. Environmental
and regulatory concerns are exacerbating that problem.
[0003] Solar-related renewable energy is one alternative energy
source that may provide a portion of the solution to the growing
demand for energy. Solar-related renewable energy is appealing
because, unlike fossil fuels, uranium, or even thermal "green"
energy, there are few or no climatic risks associated with its use.
In addition, solar related energy is free and vastly abundant.
[0004] Ocean Thermal Energy Conversion ("OTEC") is a manner of
producing renewable energy using solar energy stored as heat in the
oceans' tropical regions. Tropical oceans and seas around the world
offer a unique renewable energy resource. In many tropical areas
(between approximately 20.degree. north and 20.degree. south
latitude), the temperature of the surface sea water remains nearly
constant. To depths of approximately 100 ft the average surface
temperature of the sea water varies seasonally between 75.degree.
F. and 85.degree. F. or more. In the same regions, deep ocean water
(between 2500 ft and 4200 ft or more) remains a fairly constant
40.degree. F. Thus, the tropical ocean structure offers a large
warm water reservoir at the surface and a large cold water
reservoir at depth, with a temperature difference between the warm
and cold reservoirs of between 35.degree. F. to 45.degree. F. This
temperature difference (.DELTA.T) remains fairly constant
throughout the day and night, with small seasonal changes.
[0005] The OTEC process uses the temperature difference between
surface and deep sea tropical waters to drive a heat engine to
produce electrical energy. OTEC power generation was identified in
the late 1970's as a possible renewable energy source having a low
to zero carbon footprint for the energy produced. An OTEC power
plant, however, has a low thermodynamic efficiency compared to more
traditional, high pressure, high temperature power generation
plants. For example, using the average ocean surface temperatures
between 80.degree. F. and 85.degree. F. and a constant deep water
temperature of 40.degree. F., the maximum ideal Carnot efficiency
of an OTEC power plant will be 7.5 to 8%. In practical operation,
the gross power efficiency of an OTEC power system has been
estimated to be about half the Carnot limit, or approximately 3.5
to 4.0%. Additionally, analysis performed by leading investigators
in the 1970's and 1980's, and documented in "Renewable Energy from
the Ocean, a Guide to OTEC" William Avery and Chih Wu, Oxford
University Press, 1994 (incorporated herein by reference),
indicates that between one quarter to one half (or more) of the
gross electrical power generated by an OTEC plant operating with a
.DELTA.T of 40.degree. F. would be required to run the water and
working fluid pumps and to supply power to other auxiliary needs of
the plant. On this basis, the low overall net efficiency of an OTEC
power plant converting the thermal energy stored in the ocean
surface waters to net electric energy has not been a commercially
viable energy production option.
[0006] An additional factor resulting in further reductions in
overall thermodynamic efficiency is the loss associated with
providing necessary controls on the turbine for precise frequency
regulation. This introduces pressure losses in the turbine cycle
that limit the work that can be extracted from the warm sea water.
The resulting net plant efficiency would then be between 1.5% and
2.0%
[0007] This low OTEC net efficiency compared with efficiencies
typical of heat engines that operate at high temperatures and
pressures has led to the widely held assumption by energy planners
that OTEC power is too costly to compete with more traditional
methods of power production.
[0008] Indeed, the parasitic electrical power requirements are
particularly important in an OTEC power plant because of the
relatively small temperature difference between the hot and cold
water. To achieve maximum heat transfer between the warm sea water
and the working fluid, and between the cold sea water and the
working fluid large heat exchange surface areas are required, along
with high fluid velocities. Increasing any one of these factors can
increase the parasitic load on the OTEC plant, thereby decreasing
net efficiency. An efficient heat transfer system that maximizes
the energy transfer in the limited temperature differential between
the sea water and the working fluid would increase the commercial
viability of an OTEC power plant.
[0009] In addition to the relatively low efficiencies with
seemingly inherent large parasitic loads, the operating environment
of OTEC plants presents design and operating challenges that also
decrease the commercial viability of such operations. As previously
mentioned, the warm water needed for the OTEC heat engine is found
at the surface of the ocean, to a depth of 100 ft or less. The
constant source of cold water for cooling the OTEC engine is found
at a depth of between 2700 ft and 4200 ft or more. Such depths are
not typically found in close proximity to population centers or
even land masses. An offshore power plant is required.
[0010] Whether the plant is floating or fixed to an underwater
feature, a long cold water intake pipe of 2000 ft or longer is
required. Moreover, because of the large volume of water required
in commercially viable OTEC operations, the cold water intake pipe
requires a large diameter (typically between 6 and 35 feet or
more). Suspending a large diameter pipe from an offshore structure
presents stability, connection and construction challenges which
have previously driven OTEC costs beyond commercial viability.
[0011] Additionally, a pipe having significant length to diameter
ratio that is suspended in a dynamic ocean environment can be
subjected to temperature differences and varying ocean currents
along the length of the pipe. Stresses from bending and vortex
shedding along the pipe also present challenges. And surface
influences such as wave action present further challenges with the
connection between the pipe and floating platform. A cold water
pipe intake system having desirable performance, connection, and
construction consideration would increase the commercial viability
of an OTEC power plant.
[0012] Environmental concerns associated with an OTEC plant have
also been an impediment to OTEC operations. Traditional OTEC
systems draw in large volumes of nutrient rich cold water from the
ocean depths and discharge this water at or near the surface. Such
discharge can effect, in a positive or adverse manner, the ocean
environment near the OTEC plant, impacting fish stocks and reef
systems that may be down current from the OTEC discharge.
SUMMARY
[0013] In some aspects, power generation plant uses ocean thermal
energy conversion processes as a power source.
[0014] Further aspects relate to an offshore OTEC power plant
having improved overall efficiencies with reduced parasitic loads,
greater stability, lower construction and operating costs, and
improved environmental footprint. Other aspects include large
volume water conduits that are integral with the floating
structure. Modularity and compartmentation of the multi-stage OTEC
heat engine reduces construction and maintenance costs, limits
off-grid operation and improves operating performance. Still
further aspects provide for a floating platform having structurally
integrated heat exchange compartments and provides for low movement
of the platform due to wave action. The integrated floating
platform may also provide for efficient flow of the warm water or
cool water through the multi-stage heat exchanger, increasing
efficiency and reducing the parasitic power demand. Associated
system can promote an environmentally neutral thermal footprint by
discharging warm and cold water at appropriate depth/temperature
ranges. Energy extracted in the form of electricity reduces the
bulk temperature to the ocean.
[0015] Further aspects relate to a floating, low heave OTEC power
plant having a high efficiency, multi-stage heat exchange system,
wherein the warm and cold water supply conduits and heat exchanger
cabinets are structurally integrated into the floating platform or
structure of the power plant.
[0016] In some aspects, heat exchange plates include: front and
back exterior surfaces exposed to a non-working fluid; and an
interior working fluid flow channel between the front and back
exterior surfaces, comprising a first plurality of parallel flow
paths in a first direction and a second plurality of parallel flow
paths in a second direction. Embodiments of these systems can
include one or more of the following features.
[0017] In some embodiments, the first direction is opposite and
parallel to the second direction.
[0018] In some embodiments, the first and second directions of the
first and second flow paths are perpendicular to the direction of
flow of the non-working fluid.
[0019] In some embodiments, wherein the plate further comprises a
first region of relatively high working fluid mass flux when a
working fluid has a low vapor quality and a second region of
relatively low working fluid mass flux when the working fluid has a
high vapor quality.
[0020] In some embodiments, the interior flow channel has an area
of varying space in fluid contact with the first plurality of
parallel flow paths and the second plurality of parallel flow
paths. In some cases, the plates also include one or more
structural walls within the first and second plurality of flow
paths, the one or more walls being generally parallel to the flow
path and terminating in the area of varying space. In some cases,
the one or more structural walls comprises directional vane at the
terminal end and oriented in the direction of flow. In some cases,
the one or more structural walls comprise a directional vane at the
proximal end of the structural wall.
[0021] In some embodiments, a flow path of the first and second
plurality of flow paths comprises a void having a cross sectional
area of between 155 mm and 60 mm.
[0022] In some embodiments, the heat exchange plate is a composite
blow molded plate.
[0023] In some embodiments, the heat exchange plate is
aluminum.
[0024] In some embodiments, the working fluid pressure drop across
the plate is about 0.2 psi/ft.
[0025] In some embodiments, the non-working fluid heat transfer
coefficient ranges from 900 to 1400 Btu/ft2 Rhr.
[0026] In some embodiments, the pattern of a first plurality of
parallel flow paths in a first direction and a second plurality of
parallel flow paths in a second direction repeats across the length
of the heat exchange plate. In some cases, the number of flow paths
in the first and second plurality of flow paths increases as the
pattern repeats across the length of the heat exchange plate. In
some cases, the number of flow paths in the first plurality of flow
paths increases from four flow paths per first direction to six
flow paths per first direction. In some cases, the number of flow
paths in the first plurality of flow paths increases from two flow
paths per first direction to four flow paths per first
direction.
[0027] In some embodiments, the non-working fluid is sea water.
[0028] In some embodiments, the working fluid is ammonia.
[0029] In some embodiments, the plate is an OTEC heat exchange
plate.
[0030] Still further aspects include a floating ocean thermal
energy conversion power plant. A low heave structure, such as a
spar, or modified semi-submersible offshore structure may comprise
a first deck portion having structurally integral warm sea water
passages, multi-stage heat exchange surfaces, and working fluid
passages, wherein the first deck portion provides for the
evaporation of the working fluid. A second deck portion is also
provided having structurally integral cold sea water passages,
multi-stage heat exchange surfaces, and working fluid passages,
wherein the second deck portion provides a condensing system for
condensing the working fluid from a vapor to a liquid. The first
and second deck working fluid passages are in communication with a
third deck portion comprising one or more vapor turbine driven
electric generators for power generation.
[0031] In one aspect, an offshore power generation structure is
provided comprising a submerged portion. The submerged portion
further comprises a first deck portion comprising an integral
multi-stage evaporator system, a second deck portion comprising an
integral multi-stage condensing system; a third deck portion
housing power generation and transformation equipment; a cold water
pipe and a cold water pipe connection.
[0032] In a further aspect, the first deck portion further
comprises a first stage warm water structural passage forming a
high volume warm water conduit. The first deck portion also
comprises a first stage working fluid passage arranged in
cooperation with the first stage warm water structural passage to
warm a working fluid to a vapor. The first deck portion also
comprises a first stage warm water discharge directly coupled to a
second stage warm water structural passage. The second stage warm
water structural passage forms a high volume warm water conduit and
comprises a second stage warm water intake coupled to the first
stage warm water discharge. The arrangement of the first stage warm
water discharge to the second stage warm water intake provides low
pressure loss in the warm water flow between the first and second
stage. The first deck portion also comprises a second stage working
fluid passage arranged in cooperation with the second stage warm
water structural passage to warm the working fluid to a vapor. The
first deck portion also comprises a second stage warm water
discharge.
[0033] In a further aspect, the submerged portion further comprises
a second deck portion comprising a first stage cold water
structural passage forming a high volume cold water conduit. The
first stage cold water passage further comprises a first stage cold
water intake. The second deck portion also comprises a first stage
working fluid passage in communication with the first stage working
fluid passage of the first deck portion. The first stage working
fluid passage of the second deck portion in cooperation with the
first stage cold water structural passage cools the working fluid
to a liquid. The second deck portion also comprises a first stage
cold water discharge directly coupled to a second stage cold water
structural passage forming a high volume cold water conduit. The
second stage cold water structural passage comprises a second stage
cold water intake. The first stage cold water discharge and the
second stage cold water intake are arranged to provide low pressure
loss in the cold water flow from the first stage cold water
discharge to the second stage cold water intake. The second deck
portion also comprises a second stage working fluid passage in
communication with the second stage working fluid passage of the
first deck portion. The second stage working fluid passage in
cooperation with the second stage cold water structural passage
cool the working fluid within the second stage working fluid
passage to a liquid. The second deck portion also comprises a
second stage cold water discharge.
[0034] In a further aspect, the third deck portion may comprise a
first and second vapor turbine, wherein the first stage working
fluid passage of the first deck portion is in communication with
the first turbine and the second stage working fluid passage of the
first deck portion is in communication with the second turbine. The
first and second turbine can be coupled to one or more electric
generators.
[0035] In still further aspects, an offshore power generation
structure is provided comprising a submerged portion, the submerged
portion further comprises a four stage evaporator portion, a four
stage condenser portion, a four stage power generation portion, a
cold water pipe connection, and a cold water pipe.
[0036] In one aspect, the four stage evaporator portion comprises a
warm water conduit including, a first stage heat exchange surface,
a second stage heat exchange surface, a third stage heat exchange
surface, and fourth stage heat exchange surface. The warm water
conduit comprises a vertical structural member of the submerged
portion. The first, second, third and fourth heat exchange surfaces
are in cooperation with first, second, third and fourth stage
portions of a working fluid conduit, wherein a working fluid
flowing through the working fluid conduit is heated to a vapor at
each of the first, second, third, and fourth stage portions.
[0037] In one aspect, the four stage condenser portion comprises a
cold water conduit including a first stage heat exchange surface, a
second stage heat exchange surface, a third stage heat exchange
surface, and fourth stage heat exchange surface. The cold water
conduit comprises a vertical structural member of the submerged
portion. The first, second, third and fourth heat exchange surfaces
are in cooperation with first, second, third and fourth stage
portions of a working fluid conduit, wherein a working fluid
flowing through the working fluid conduit is cooled to a liquid at
each of the first, second, third, and fourth stage portions, with a
progressively higher temperature at each successive stage.
[0038] In yet another aspect, first, second, third and fourth stage
working fluid conduits of the evaporator portion are in
communication with first, second, third and fourth vapor turbines,
wherein the evaporator portion first stage working fluid conduit is
in communication with a first vapor turbine and exhausts to the
fourth stage working fluid conduit of the condenser portion.
[0039] In yet another aspect, first, second, third and fourth stage
working fluid conduits of the evaporator portion are in
communication with first, second, third and fourth vapor turbines,
wherein the evaporator portion second stage working fluid conduit
is in communication with a second vapor turbine and exhausts to the
third stage working fluid conduit of the condenser portion.
[0040] In yet another aspect, first, second, third and fourth stage
working fluid conduits of the evaporator portion are in
communication with first, second, third and fourth vapor turbines,
wherein the evaporator portion third stage working fluid conduit is
in communication with a third vapor turbine and exhausts to the
second stage working fluid conduit of the condenser portion.
[0041] In yet another aspect, first, second, third and fourth stage
working fluid conduits of the evaporator portion are in
communication with first, second, third and fourth vapor turbines,
wherein the evaporator portion fourth stage working fluid conduit
is in communication with a fourth vapor turbine and exhausts to the
first stage working fluid conduit of the condenser portion.
[0042] In still a further aspect, a first electrical generator is
driven by the first turbine, the fourth turbine, or a combination
of the first and fourth turbine.
[0043] In still a further aspect, a second electrical generator is
driven by the second turbine, the third turbine, or a combination
of both the second and third turbine.
[0044] Additional aspects can incorporate one or more of the
following features: the first and fourth turbines or the second and
third turbines produce between 9 MW and 60 MW of electrical power;
the first and second turbines produce approximately 55 MW of
electrical power; the first and second turbines form one of a
plurality of turbine-generator sets in an Ocean Thermal Energy
Conversion power plant; the first stage warm water intake is free
of interference from the second stage cold water discharge; the
first stage cold water intake is free of interference from the
second stage warm water discharge; the working fluid within the
first or second stage working fluid passages comprises a commercial
refrigerant. The working fluid comprises any fluid with suitable
thermodynamic properties such as ammonia, propylene, butane, R-134,
or R-22; the working fluid in the first and second stage working
fluid passages increases in temperature between 12.degree. F. and
24.degree. F.; a first working fluid flows through the first stage
working fluid passage and a second working fluid flows through the
second stage working fluid passage, wherein the second working
fluid enters the second vapor turbine at a lower temperature than
the first working fluid enters the first vapor turbine; the working
fluid in the first and second stage working fluid passages
decreases in temperature between 12.degree. F. and 24.degree. F.; a
first working fluid flows through the first stage working fluid
passage and a second working fluid flows through the second stage
working fluid passage, wherein the second working fluid enters the
second deck portion at a lower temperature than the first working
fluid enters the second deck portion.
[0045] Further aspects can also incorporate one or more of the
following features: the warm water flowing within the first or
second stage warm water structural passage comprises, warm sea
water, geo-thermally heated water, solar heated reservoir water;
heated industrial cooling water, or a combination thereof; the warm
water flows between 500,000 and 6,000,000 gpm; the warm water flows
at 5,440,000 gpm; the warm water flows between 300,000,000 lb/hr
and 1,000,000,000 lb/hr; the warm water flows at 2,720,000 lb/hr;
the cold water flowing within the first or second stage cold water
structural passage comprises cold sea water, cold fresh water, cold
subterranean water or a combination thereof; the cold water flows
between 250,000 and 3,000,000 gpm; the cold water flows at
3,420,000 gpm; the cold water flows between 125,000,000 lb/hr and
1,750,000,000 lb/hr; the cold water flows at 1,710,000 lb/hr.
[0046] Aspects can also incorporate one or more of the following
features: the offshore structure is a low heave structure; the
offshore structure is a floating spar structure; the offshore
structure is a semi-submersible structure.
[0047] A still further aspect can include a high-volume,
low-velocity heat exchange system for use in an ocean thermal
energy conversion power plant, comprising: a first stage cabinet
that further comprises a first water flow passage for heat exchange
with a working fluid; and a first working fluid passage; and a
second stage cabinet coupled to the first stage cabinet, that
further comprises a second water flow passage for heat exchange
with a working fluid and coupled to the first water flow passage in
a manner to limit pressure drop of water flowing from the first
water flow passage to the second water flow passage; and a second
working fluid passage. The first and second stage cabinets comprise
structural members of the power plant.
[0048] In one aspect, water flows from the first stage cabinet to
the second stage cabinet and the second stage cabinet is beneath
the first stage cabinet evaporator. In another aspect, water flows
from the first stage cabinet to the second stage cabinet and the
second stage cabinet is above the first stage cabinet in the
condensers and below the first stage cabinet in the
evaporators.
[0049] In still a further aspect, a cold water pipe provides cold
water from ocean depths to the cold water intake of the OTEC. The
cold water intake can be in the second deck portion of the
submerged portion of the OTEC plant. The cold water pipe can be a
segmented construction. The cold water pipe can be a continuous
pipe. The cold water pipe can comprise: an elongated tubular
structure having an outer surface, a top end and a bottom end. The
tubular structure can further comprise a plurality of first and
second stave segments wherein each stave segment has a top portion
and a bottom portion, and wherein the top portion of the second
stave segment is offset from the top portion of the first staved
segment. The cold water pipe can include a strake or ribbon at
least partially wound spirally about the outer surface. The first
and second staves and/or the strake can comprise polyvinyl chloride
(PVC), chlorinated polyvinyl chloride (CPVC), fiber reinforced
plastic (FRP), reinforced polymer mortar (RPMP), polypropylene
(PP), polyethylene (PE), cross-linked high-density polyethylene
(PEX), polybutylene (PB), acrylonitrile butadiene styrene (ABS);
polyester, fiber reinforced polyester, vinyl ester, reinforced
vinyl ester, concrete, ceramic, or a composite of one or more
thereof.
[0050] Further aspects include a dynamic connection between the
submerged portion of the OTEC plant and the cold water pipe. The
dynamic connection can support the weight and dynamic forces of the
cold water pipe while it is suspended from the OTEC platform. The
dynamic pipe connection can allow for relative movement between the
OTEC platform and the cold water pipe. The relative movement can be
between 0.5.degree. and 30.degree. from vertical. In one aspect the
relative movement can be between 0.5.degree. and 5.degree. from
vertical. The dynamic pipe connection can include a spherical or
arcuate bearing surface.
[0051] In some embodiments, a static connection is provided between
the submerged portion of the OTEC plant and the cold water pipe. In
these systems, the top of the cold water pipe can be conical and is
retracted into a conical receptacle using lines and winches lowered
from within the spar. The old water pipe can be retained using
locking mechanisms such that the lines can be detached for use in
lifting equipment from the lower decks of the spar to the mid-body
decks.
[0052] In an aspect, a submerged vertical pipe connection comprises
a floating structure having a vertical pipe receiving bay, wherein
the receiving bay has a first diameter, a vertical pipe for
insertion into the pipe receiving bay, the vertical pipe having a
second diameter smaller than the first diameter of the pipe
receiving bay; a bearing surface; and one or more detents operable
with the bearing surface, wherein the detents define a diameter
that is different than the first or second diameter when in contact
with the bearing surface.
[0053] More details of other aspects are described in U.S. patent
application Ser. No. ______ (Attorney Docket No. 25667-016001)
entitled Staved Ocean Thermal Energy Conversion Power Plant--Cold
Water Pipe Connection, and U.S. patent application Ser. No. ______
(Attorney Docket No. 25667-009001) entitled Ocean Thermal Energy
Conversion Power Plant, filed simultaneously with the present
application and incorporated herein by reference in their
entirety.
[0054] Aspects may have one or more of the following advantages:
OTEC power production requires little to no fuel costs for energy
production; the low pressures and low temperatures involved in the
OTEC heat engine reduce component costs and require ordinary
materials compared to the high-cost, exotic materials used in high
pressure, high temperature power generation plants; plant
reliability is comparable to commercial refrigeration systems,
operating continuously for several years without significant
maintenance; reduced construction times compared to high pressure,
high temperature plants; and safe, environmentally benign operation
and power production. Additional advantages may include, increased
net efficiency compared to traditional OTEC systems, lower
sacrificial electrical loads; reduced pressure loss in warm and
cold water passages as well as working fluid flow passages; modular
components; less frequent off-grid production time; low heave and
reduced susceptibility to wave action; discharge of cooling water
below surface levels, intake of warm water free from interference
from cold water discharge.
[0055] The details of one or more embodiments are set forth in the
accompanying drawings and the description below. Other features,
objects, and advantages will be apparent from the description and
drawings, and from the claims.
DESCRIPTION OF DRAWINGS
[0056] FIG. 1 illustrates an exemplary prior-art OTEC heat
engine.
[0057] FIG. 2 illustrates an exemplary prior-art OTEC power
plant.
[0058] FIG. 3 illustrates OTEC structure.
[0059] FIG. 4 illustrates a deck plan of a heat exchanger deck.
[0060] FIG. 5 illustrates a cabinet heat exchanger.
[0061] FIG. 6A illustrates a conventional heat exchange cycle.
[0062] FIG. 6B illustrates a cascading multi-stage heat exchange
cycle.
[0063] FIG. 6C illustrates a hybrid cascading multi-stage heat
exchange cycle.
[0064] FIG. 6D illustrates the evaporator pressure drop and
associate power production.
[0065] FIGS. 7A and B illustrate an exemplary OTEC heat engine.
[0066] FIG. 8 illustrates a conventional shell and tube heat
exchanger.
[0067] FIG. 9 illustrates a conventional plate heat exchanger.
[0068] FIG. 10 illustrates a cabinet heat exchanger.
[0069] FIG. 11 illustrates a perspective view of a heat exchange
plate arrangement.
[0070] FIG. 12 illustrates a perspective view of a heat exchange
plate arrangement.
[0071] FIG. 13 illustrates a side view of a heat exchange plate
configuration.
[0072] FIG. 14 illustrates a P-h diagram of a conventional high
temperature steam cycle.
[0073] FIG. 15 illustrates a P-h diagram of a heat cycle.
[0074] FIG. 16 illustrates an embodiment of a heat exchange
plate.
[0075] FIG. 17 illustrates an embodiment of a heat exchange
plate.
[0076] FIG. 18 illustrates a portion of a heat exchange plate.
[0077] FIGS. 19A and 19B illustrates an embodiment of a pair of
heat exchange plates.
[0078] FIGS. 20A and 20B illustrates an embodiment of a pair of
heat exchange plates.
[0079] FIGS. 21A-21D illustrate exemplary layout of heat exchange
plates for an OTEC plant.
[0080] FIGS. 22A and 22B are, respectively, a schematic plan view
of a heat exchange cassette and a cross-section of a working fluid
channel of the heat exchange plate.
[0081] FIGS. 23A and 23B are, respectively, a schematic plan view
of a heat exchange cassette and a cross-section of a working fluid
channel of the heat exchange plate.
[0082] FIGS. 24A and 24B are, respectively, a schematic plan view
of a heat exchange cassette and a cross-section of a working fluid
channel of the heat exchange plate.
[0083] FIGS. 25A and 25B are, respectively, a schematic plan view
of a heat exchange cassette and a cross-section of a working fluid
channel of the heat exchange plate.
[0084] FIGS. 26A and 26B are, respectively, a schematic plan view
of a heat exchange cassette and a cross-section of a working fluid
channel of the heat exchange plate.
[0085] FIGS. 27A and 27B compare a full-size heat exchange cassette
and a scaled-down heat exchange cassette.
[0086] FIGS. 28-32 are schematic plan views of embodiments of heat
exchange cassettes.
[0087] FIGS. 33-35 illustrate cross-sections of working fluid
channels.
[0088] Like reference symbols in the various drawings indicate like
elements.
DETAILED DESCRIPTION
[0089] This disclosure relates to electrical power generation using
Ocean Thermal Energy Conversion (OTEC) technology. Aspects relate
to a floating OTEC power plant having improved overall efficiencies
with reduced parasitic loads, greater stability, and lower
construction and operating costs over convention OTEC power plants.
Other aspects include large volume water conduits that are integral
with the floating structure. Modularity and compartmentation of the
multi-stage OTEC heat engine reduces construction and maintenance
costs, limits off-grid operation and improves operating
performance. Still further aspects provide for a floating platform
having integrated heat exchange compartments and provides for low
movement of the platform due to wave action. The integrated
floating platform may also provide for efficient flow of the warm
water or cool water through the multi-stage heat exchanger,
increasing efficiency and reducing the parasitic power demand. In
particular, highly efficient heat exchange plates can provide
increased overall efficiency, thus further reducing parasitic power
demand. Aspects promote a neutral thermal footprint by discharging
warm and cold water at appropriate depth/temperature ranges. Energy
extracted in the form of electricity reduces the bulk temperature
to the ocean.
[0090] OTEC is a process that uses heat energy from the sun that is
stored in the Earth's oceans to generate electricity. OTEC uses the
temperature difference between the warmer, top layer of the ocean
and the colder, deep ocean water. Typically this difference is at
least 36.degree. F. (20.degree. C.). These conditions exist in
tropical areas, roughly between the Tropic of Capricorn and the
Tropic of Cancer, or even 20.degree. north and south latitude. The
OTEC process uses the temperature difference to power a Rankine
cycle, with the warm surface water serving as the heat source and
the cold deep water serving as the heat sink. Rankine cycle
turbines drive generators which produce electrical power.
[0091] FIG. 1 illustrates a typical OTEC Rankine cycle heat engine
10 which includes warm sea water inlet 12, evaporator 14, warm sea
water outlet 15, turbine 16, cold sea water inlet 18, condenser 20,
cold sea water outlet 21, working fluid conduit 22 and working
fluid pump 24.
[0092] In operation, heat engine 10 can use any one of a number of
working fluids, for example commercial refrigerants such as
ammonia. Other working fluids can include propylene, butane, R-22
and R-134a. Other commercial refrigerants can be used. Warm sea
water between approximately 75.degree. F. and 85.degree. F., or
more, is drawn from the ocean surface or just below the ocean
surface through warm sea water inlet 12 and in turn warms the
ammonia working fluid passing through evaporator 14. The ammonia
boils to a vapor pressure of approximately 9.3 atm. The vapor is
carried along working fluid conduit 22 to turbine 16. The ammonia
vapor expands as it passes through the turbine 16, producing power
to drive an electric generator 25. The ammonia vapor then enters
condenser 20 where it is cooled to a liquid by cold sea water drawn
from a deep ocean depth of approximately 3000 ft. The cold sea
water enters the condenser at a temperature of approximately
40.degree. F. The vapor pressure of the ammonia working fluid at
the temperature in the condenser 20, approximately 51.degree. F.,
is 6.1 atm. Thus, a significant pressure difference is available to
drive the turbine 16 and generate electric power. As the ammonia
working fluid condenses, the liquid working fluid is pumped back
into the evaporator 14 by working fluid pump 24 via working fluid
conduit 22.
[0093] The heat engine 10 of FIG. 1 is essentially the same as the
Rankine cycle of most steam turbines, except that OTEC differs by
using different working fluids and lower temperatures and
pressures. The heat engine 10 of the FIG. 1 is also similar to
commercial refrigeration plants, except that the OTEC cycle is run
in the opposite direction so that a heat source (e.g., warm ocean
water) and a cold heat sink (e.g., deep ocean water) are used to
produce electric power.
[0094] FIG. 2 illustrates the components of a floating OTEC power
plant 200, which include: the vessel or platform 210, warm sea
water inlet 212, warm water pump 213, evaporator 214, warm sea
water outlet 215, turbine-generator 216, cold water pipe 217, cold
water inlet 218, cold water pump 219, condenser 220, cold water
outlet 221, working fluid conduit 222, working fluid pump 224, and
pipe connections 230. OTEC plant 200 can also include electrical
generation, transformation and transmission systems, position
control systems such as propulsion, thrusters, or mooring systems,
as well as various auxiliary and support systems (for example,
personnel accommodations, emergency power, potable water, black and
grey water, firefighting, damage control, reserve buoyancy, and
other common shipboard or marine systems).
[0095] Implementations of OTEC power plants utilizing the basic
heat engine and system of FIGS. 1 and 2 have a relatively low
overall efficiency of 3% or below. Because of this low thermal
efficiency, OTEC operations require the flow of large amounts of
water through the power system per kilowatt of power generated.
This in turn requires large heat exchangers having large heat
exchange surface areas.
[0096] Such large volumes of water and large surface areas require
considerable pumping capacity in the warm water pump 213 and cold
water pump 219, reducing the net electrical power available for
distribution to a shore-based facility or on board industrial
purposes. Moreover, the limited space of most surface vessels does
not easily facilitate large volumes of water directed to and
flowing through the evaporator or condenser. Indeed, large volumes
of water require large diameter pipes and conduits. Putting such
structures in limited space requires multiple bends to accommodate
other machinery. And the limited space of typical surface vessels
or structures does not easily facilitate the large heat exchange
surface area required for maximum efficiency in an OTEC plant. Thus
the OTEC systems and vessel or platform have traditionally been
large and costly. This has led to an industry conclusion that OTEC
operations are a high cost, low yield energy production option when
compared to other energy production options using higher
temperatures and pressures.
[0097] The systems and approaches described herein address
technical challenges in order to improve the efficiency of OTEC
operations and reduce the cost of construction and operation.
[0098] The vessel or platform 210 requires low motions to limit
dynamic forces between the cold water pipe 217 and the vessel or
platform 210 and to provide a benign operating environment for the
OTEC equipment in the platform or vessel. The vessel or platform
210 should also support cold and warm water inlet (218 and 212)
volume flows, bringing in sufficient cold and warm water at
appropriate levels to ensure OTEC process efficiency. The vessel or
platform 210 should also enable cold and warm water discharge via
cold and warm water outlets (221 and 215) well below the waterline
of vessel or platform 210 to avoid thermal recirculation into the
ocean surface layer. Additionally, the vessel or platform 210
should survive heavy weather without disrupting power generating
operations.
[0099] The OTEC heat engine 10 described herein uses a highly
efficient thermal cycle for maximum efficiency and power
production. Heat transfer in boiling and condensing processes, as
well as the heat exchanger materials and design, limit the amount
of energy that can be extracted from each pound of warm seawater.
The heat exchangers used in the evaporator 214 and the condenser
220 use high volumes of warm and cold water flow with low head loss
to limit parasitic loads. The heat exchangers also provide high
coefficients of heat transfer to enhance efficiency. The heat
exchangers incorporate materials and designs tailored to the warm
and cold water inlet temperatures to enhance efficiency. The heat
exchanger design can use a simple construction method with low
amounts of material to reduce cost and volume.
[0100] The turbo generators 216 are highly efficient with low
internal losses and may also be tailored to the working fluid to
enhance efficiency
[0101] FIG. 3 illustrates an implementation of an OTEC system that
enhances the efficiency of previous OTEC power plants and overcomes
many of the technical challenges associated therewith. This
implementation comprises a spar for the vessel or platform, with
heat exchangers and associated warm and cold water piping integral
to the spar.
[0102] OTEC Spar 310 houses an integral multi-stage heat exchange
system for use with an OTEC power generation plant. Spar 310
includes a submerged portion 311 below waterline 305. Submerged
portion 311 comprises warm water intake portion 340, evaporator
portion 344, warm water discharge portion 346, condenser portion
348, cold water intake portion 350, cold water pipe 351, cold water
discharge portion 352, machinery deck portion 354, and deck house
360.
[0103] In operation, warm sea water of between 75.degree. F. and
85.degree. F. is drawn through warm water intake portion 340 and
flows down the spar though structurally integral warm water
conduits (not shown). Due to the high volume water flow
requirements of OTEC heat engines, the warm water conduits direct
flow to the evaporator portion 344 of between 500,000 gpm and
6,000,000 gpm. The warm water conduits have a diameter of between 6
ft and 35 ft, or more. Due to this size, the warm water conduits
are vertical structural members of spar 310. Warm water conduits
can be large diameter pipes of sufficient strength to vertically
support spar 310. Alternatively, the warm water conduits can be
passages integral to the construction of the spar 310.
[0104] Warm water then flows through the evaporator portion 344
which houses one or more stacked, multi-stage heat exchangers for
warming a working fluid to a vapor. The warm sea water is then
discharged from spar 310 via warm water discharge 346. Warm water
discharge can be located or directed via a warm water discharge
pipe to a depth at or close to an ocean thermal layer that is
approximately the same temperature as the warm water discharge
temperature to limit environmental impacts. The warm water
discharge can be directed to a sufficient depth to avoid thermal
recirculation with either the warm water intake or cold water
intake.
[0105] Cold sea water is drawn from a depth of between 2500 and
4200 ft, or more, at a temperature of approximately 40.degree. F.,
via cold water pipe 351. The cold sea water enters spar 310 via
cold water intake portion 350. Due to the high volume water flow
requirements of OTEC heat engines, the cold sea water conduits
direct flow to the condenser portion 348 of between 500,000 gpm and
3,500,000 gpm. Such cold sea water conduits have a diameter of
between 6 ft and 35 ft, or more. Due to this size, the cold sea
water conduits are vertical structural members of spar 310. Cold
water conduits can be large diameter pipes of sufficient strength
to vertically support spar 310. Alternatively, the cold water
conduits can be passages integral to the construction of the spar
310.
[0106] Cold sea water then flows upward to stacked multi-stage
condenser portion 348, where the cold sea water cools a working
fluid to a liquid. The cold sea water is then discharged from spar
310 via cold sea water discharge 352. Cold water discharge can be
located or directed via a cold sea water discharge pipe to depth at
or close to an ocean thermal layer that is approximately the same
temperature as the cold sea water discharge temperature. The cold
water discharge can be directed to a sufficient depth to avoid
thermal recirculation with either the warm water intake or cold
water intake.
[0107] Machinery deck portion 354 can be positioned vertically
between the evaporator portion 344 and the condenser portion 348.
Positioning machinery deck portion 354 beneath evaporator portion
344 allows nearly straight line warm water flow from intake,
through the multi-stage evaporators, and to discharge. Positioning
machinery deck portion 354 above condenser portion 348 allows
nearly straight line cold water flow from intake, through the
multi-stage condensers, and to discharge. Machinery deck portion
354 includes turbo generators 356. In operation, warm working fluid
heated to a vapor flows from evaporator portion 344 to one or more
turbo generators 356. The working fluid expands in turbo generator
356 thereby driving a turbine for the production of electrical
power. The working fluid then flows to condenser portion 348 where
it is cooled to a liquid and pumped to evaporator portion 344.
[0108] FIG. 4 illustrates an implementation of an OTEC system
wherein a plurality of multi-stage heat exchangers 420 is arranged
about the periphery of OTEC spar 410. Heat exchangers 420 can be
evaporators or condensers used in an OTEC heat engine. The
peripheral layout of heat exchanges can be used with evaporator
portion 344 or condenser portion 348 of an OTEC spar platform (as
shown in FIG. 3). The peripheral arrangement can support any number
of heat exchangers (e.g., 1 heat exchanger, between 2 and 8 heat
exchangers, 8-16 heat exchanger, 16-32 heat exchangers, or 32 or
more heat exchangers). One or more heat exchangers can be
peripherally arranged on a single deck or on multiple decks (e.g.,
on 2, 3, 4, 5, or 6 or more decks) of the OTEC spar 410. One or
more heat exchangers can be peripherally offset between two or more
decks such that no two heat exchangers are vertically aligned over
one another. One or more heat exchangers can be peripherally
arranged so that heat exchangers in one deck are vertically aligned
with heat exchanges on another adjacent deck.
[0109] Individual heat exchangers 420 can comprise a multi-stage
heat exchange system (e.g., a 2, 3, 4, 5, or 6 or more heat
exchange system). In some embodiments, individual heat exchangers
420 are cabinet heat exchangers constructed to provide low pressure
loss in the warm sea water flow, cold sea water flow, and working
fluid flow through the heat exchanger.
[0110] Referring to FIG. 5, an embodiment of a cabinet heat
exchanger 520 includes multiple heat exchange stages, 521, 522, 523
and 524. In some implementations, the stacked heat exchangers
accommodate warm sea water flowing down through the cabinet, from
first evaporator stage 521, to second evaporator stage 522, to
third evaporator stage 523 to fourth evaporator stage 524. In
another embodiment of the stacked heat exchange cabinet, cold sea
water flows up through the cabinet from first condenser stage 531,
to second condenser stage 532, to third condenser stage 533, to
fourth condenser stage 534. Working fluid flows through working
fluid supply conduits 538 and working fluid discharge conduits 539.
In an embodiment, working fluid conduits 538 and 539 enter and exit
each heat exchanger stage horizontally as compared to the vertical
flow of the warm sea water or cold sea water. The vertical
multi-stage heat exchange design of cabinet heat exchanger 520
facilitates an integrated vessel (e.g., spar) and heat exchanger
design, removes the requirement for interconnecting piping between
heat exchanger stages, and ensures that virtually all of the heat
exchanger system pressure drop occurs over the heat transfer
surface.
[0111] The heat transfer surface efficiency can be improved using
surface shape, treatment and spacing as described herein. Material
selection such as alloys of aluminum offer superior economic
performance over traditional titanium base designs. The heat
transfer surface can comprise 100 Series, 3000 Series, or 5000
Series aluminum alloys. The heat transfer surface can comprise
titanium and titanium alloys.
[0112] It has been found that the multi-stage heat exchanger
cabinet enables high energy transfer to the working fluid from the
sea water within the relatively low available temperature
differential of the OTEC heat engine. The thermodynamic efficiency
of any OTEC power plant is a function of how close the temperature
of the working fluid approaches that of the sea water. The physics
of the heat transfer dictate that the area required to transfer the
energy increases as the temperature of the working fluid approaches
that of the sea water. Increasing the velocity of the sea water can
increase the heat transfer coefficient to offset the increase in
surface area. However, increasing the velocity of the sea water can
greatly increases the power required for pumping, thereby
increasing the parasitic electrical load on the OTEC plant.
[0113] FIG. 6A illustrates an OTEC cycle wherein the working fluid
is boiled in a heat exchanger using warm surface sea water. The
fluid properties in this conventional Rankine cycle are constrained
by the boiling process that limits the leaving working fluid to
approximately 3.degree. F. below the leaving warm seawater
temperature. In a similar fashion, the condensing side of the cycle
is limited to being no close than 2.degree. F. higher than the
leaving cold seawater temperature. The total available temperature
drop for the working fluid is approximately 12.degree. F. (between
68.degree. F. and 56.degree. F.).
[0114] It has been found that a cascading multi-stage OTEC cycle
allows the working fluid temperatures to more closely match that of
the sea water. This increase in temperature differential increases
the amount of work that can be done by the turbines associated with
the OTEC heat engine.
[0115] FIG. 6B illustrates a cascading multi-stage OTEC cycle using
multiple steps of boiling and condensing to expand the available
working fluid temperature drop. Each step requires an independent
heat exchanger, or a dedicated heat exchanger stage in the cabinet
heat exchanger 520 of FIG. 5. The cascading multi-stage OTEC cycle
of FIG. 6b allows for matching the output of the turbines with the
expected pumping loads for the sea water and working fluid. This
highly efficient design would require dedicated and customized
turbines.
[0116] FIG. 6C illustrates a hybrid yet still efficient cascading
OTEC cycle that facilitates the use of identical equipment (e.g.,
turbines, generators, pumps) while retaining the thermodynamic
efficiencies or optimization of the true cascade arrangement of
FIG. 6B. In the hybrid cascade cycle of FIG. 6C, the available
temperature differential for the working fluid ranges from about
18.degree. F. to about 22.degree. F. This narrow range allows the
turbines in the heat engine to have identical performance
specifications, thereby lowering construction and operation
costs.
[0117] System performance and power output is greatly increased
using the hybrid cascade cycle in an OTEC power plant. Table A
compares the performance of the conventional cycle of FIG. 6A with
that of the hybrid cascading cycle of FIG. 6C.
TABLE-US-00001 TABLE A Estimated Performance for 100 MW Net Output
Four Stage Hybrid Conventional Cycle Cascade Cycle Warm Sea Water
4,800,000 GPM 3,800,000 GPM Flow Cold Sea Water 3,520,000 GPM
2,280,000 GPM Flow Gross Heat Rate 163,000 BTU/kWH 110,500
BTU/kWH
Utilizing the four stage hybrid cascade heat exchange cycle reduces
the amount of energy that needs to be transferred between the
fluids. This in turn serves to reduce the amount of heat exchange
surface that is required.
[0118] The performance of heat exchangers is affected by the
available temperature difference between the fluids as well as the
heat transfer coefficient at the surfaces of the heat exchanger.
The heat transfer coefficient generally varies with the velocity of
the fluid across the heat transfer surfaces. Higher fluid
velocities require higher pumping power, thereby reducing the net
efficiency of the plant. A hybrid cascading multi-stage heat
exchange system facilitates lower fluid velocities and greater
plant efficiencies. The stacked hybrid cascade heat exchange design
also facilitates lower pressure drops through the heat exchanger.
And the vertical plant design facilitates lower pressure drop
across the whole system.
[0119] FIG. 6D illustrates the impact of heat exchanger pressure
drop on the total OTEC plant generation to deliver 100 MW to a
power grid. Limiting pressure drop through the heat exchanger
greatly enhances the OTEC power plant's performance. Pressure drop
is reduced by providing an integrated vessel or platform-heat
exchanger system, wherein the sea water conduits form structural
members of the vessel and allow for sea water flow from one heat
exchanger stage to another in series. An approximate straight line
sea water flow, with low changes in direction from intake into the
vessel, through the pump, through the heat exchange cabinets and in
turn through each heat exchange stage in series, and ultimate
discharging from the plant, allows for low pressure drop.
[0120] Cascade Hybrid OTEC Power Generation:
[0121] An integrated multi-stage OTEC power plant can produce
electricity using the temperature differential between the surface
water and deep ocean water in tropical and subtropical regions.
Traditional piping runs for sea water can be eliminated by using
the off-shore vessel's or platform's structure as a conduit or flow
passage. Alternatively, the warm and cold sea water piping runs can
use conduits or pipes of sufficient size and strength to provide
vertical or other structural support to the vessel or platform.
These integral sea water conduit sections or passages serve as
structural members of the vessel, thereby reducing the requirements
for additional steel. As part of the integral sea water passages,
multi-stage cabinet heat exchangers provides multiple stages of
working fluid evaporation without the need for external water
nozzles or piping connections. The integrated multi-stage OTEC
power plant allows the warm and cold sea water to flow in their
natural directions. The warm sea water flows downward through the
vessel as it is cooled before being discharged into a cooler zone
of the ocean. In a similar fashion, the cold sea water from deep in
the ocean flows upward through the vessel as it is warmed before
discharging into a warmer zone of the ocean. This arrangement
avoids the need for changes in sea water flow direction and
associated pressure losses. The arrangement also reduces the
pumping energy required.
[0122] Multi-stage cabinet heat exchangers allow for the use of a
hybrid cascade OTEC cycle. These stacks of heat exchangers comprise
multiple heat exchanger stages or sections that have sea water
passing through them in series to boil or condense the working
fluid as appropriate. In the evaporator section, the warm sea water
passes through a first stage where it boils off some of the working
fluid as the sea water is cooled. The warm sea water then flows
down the stack into the next heat exchanger stage and boils off
additional working fluid at a slightly lower pressure and
temperature. This occurs sequentially through the entire stack.
Each stage or section of the cabinet heat exchanger supplies
working fluid vapor to a dedicated turbine that generates
electrical power. Each of the evaporator stages has a corresponding
condenser stage at the exhaust of the turbine. The cold sea water
passes through the condenser stacks in a reverse order to the
evaporators.
[0123] Referring to FIGS. 7A and 7B, an exemplary multi-stage OTEC
heat engine 710 utilizes a hybrid cascading heat exchange cycles.
Warm sea water is pumped from a warm sea water intake (not shown)
by warm water pump 712, discharging from the pump at approximately
1,360,000 gpm and at a temperature of approximately 79.degree. F.
All or parts of the warm water conduit from the warm water intake
to the warm water pump, and from the warm water pump to the stacked
heat exchanger cabinet can form integral structural members of the
vessel.
[0124] From the warm water pump 712, the warm sea water then enters
a first stage evaporator 714 where it boils a first working fluid.
The warm water exits the first stage evaporator 714 at a
temperature of approximately 76.8.degree. F. and flows down to a
second stage evaporator 715.
[0125] The warm water enters the second stage evaporator 715 at
approximately 76.8.degree. F. where it boils a second working fluid
and exits the second stage evaporator 715 at a temperature of
approximately 74.5.degree..
[0126] The warm water flows down to a third stage evaporator 716
from the second stage evaporator 715, entering at a temperature of
approximately 74.5.degree. F., where it boils a third working
fluid. The warm water exits the third stage evaporator 716 at a
temperature of approximately 72.3.degree. F.
[0127] The warm water then flows from the third stage evaporator
716 down to the fourth stage evaporator 717, entering at a
temperature of approximately 72.3.degree. F., where it boils a
fourth working fluid. The warm water exits the fourth stage
evaporator 717 at a temperature of approximately 70.1.degree. F.
and then discharges from the vessel. Though not shown, the
discharge can be directed to a thermal layer at an ocean depth of
approximately the same temperature as the discharge temperature of
the warm sea water. Alternately, the portion of the power plant
housing the multi-stage evaporator can be located at a depth within
the structure so that the warm water is discharged to an
appropriate ocean thermal layer. In some embodiments, the warm
water conduit from the fourth stage evaporator to the warm water
discharge of the vessel can comprise structural members of the
vessel.
[0128] Similarly, cold sea water is pumped from a cold sea water
intake (not shown) via cold sea water pump 722, discharging from
the pump at approximately 855,003 gpm and at a temperature of
approximately 40.0.degree. F. The cold sea water is drawn from
ocean depths of between approximately 2700 and 4200 ft, or more.
The cold water conduit carrying cold sea water from the cold water
intake of the vessel to the cold water pump, and from the cold
water pump to the first stage condenser can comprise in its
entirety or in part structural members of the vessel.
[0129] From cold sea water pump 722, the cold sea water enters a
first stage condenser 724, where it condenses the fourth working
fluid from the fourth stage boiler 717. The cold seawater exits the
first stage condenser at a temperature of approximately
43.5.degree. F. and flows up to a second stage condenser 725.
[0130] The cold sea water enters the second stage condenser 725 at
approximately 43.5.degree. F. where it condenses the third working
fluid from third stage evaporator 716. The cold sea water exits the
second stage condenser 725 at a temperature approximately
46.9.degree. F. and flows up to a third stage condenser 726.
[0131] The cold sea water enters the third stage condenser 726 at a
temperature of approximately 46.9.degree. F. where it condenses the
second working fluid from second stage evaporator 715. The cold sea
water exits the third stage condenser 726 at a temperature
approximately 50.4.degree. F.
[0132] The cold sea water then flows up from the third stage
condenser 726 to a fourth stage condenser 727, entering at a
temperature of approximately 50.4.degree. F. In the fourth stage
condenser, the cold sea water condenses the first working fluid
from the first stage evaporator 714. The cold sea water then exits
the fourth stage condenser at a temperature of approximately
54.0.degree. F. and ultimately discharges from the vessel. The cold
sea water discharge can be directed to a thermal layer at an ocean
depth of or approximately the same temperature as the discharge
temperature of the cold sea water. Alternately, the portion of the
power plant housing the multi-stage condenser can be located at a
depth within the structure so that the cold sea water is discharged
to an appropriate ocean thermal layer.
[0133] The first working fluid enters the first stage evaporator
714 at a temperature of 56.7.degree. F. where it is heated to a
vapor with a temperature of 74.7.degree. F. The first working fluid
then flows to first turbine 731 and then to the fourth stage
condenser 727 where the first working fluid is condensed to a
liquid with a temperature of approximately 56.5.degree. F. The
liquid first working fluid is then pumped via first working fluid
pump 741 back to the first stage evaporator 714.
[0134] The second working fluid enters the second stage evaporator
715 at a temperature approximately 53.0.degree. F. where it is
heated to a vapor. The second working fluid exits the second stage
evaporator 715 at a temperature approximately 72.4.degree. F. The
second working fluid then flow to a second turbine 732 and then to
the third stage condenser 726. The second working fluid exits the
third stage condenser at a temperature approximately 53.0.degree.
F. and flows to working fluid pump 742, which in turn pumps the
second working fluid back to the second stage evaporator 715.
[0135] The third working fluid enters the third stage evaporator
716 at a temperature approximately 49.5.degree. F. where it will be
heated to a vapor and exit the third stage evaporator 716 at a
temperature of approximately 70.2.degree. F. The third working
fluid then flows to third turbine 733 and then to the second stage
condenser 725 where the third working fluid is condensed to a fluid
at a temperature approximately 49.5.degree. F. The third working
fluid exits the second stage condenser 725 and is pumped back to
the third stage evaporator 716 via third working fluid pump
743.
[0136] The fourth working fluid enters the fourth stage evaporator
717 at a temperature of approximately 46.0.degree. F. where it will
be heated to a vapor. The fourth working fluid exits the fourth
stage evaporator 717 at a temperature approximately 68.0.degree. F.
and flow to a fourth turbine 734. The fourth working fluid exits
fourth turbine 734 and flows to the first stage condenser 724 where
it is condensed to a liquid with a temperature approximately
46.0.degree. F. The fourth working fluid exits the first stage
condenser 724 and is pumped back to the fourth stage evaporator 717
via fourth working fluid pump 744.
[0137] The first turbine 731 and the fourth turbine 734
cooperatively drive a first generator 751 and form first turbo
generator pair 761. First turbo generator pair will produce
approximately 25 MW of electric power.
[0138] The second turbine 732 and the third turbine 733
cooperatively drive a second generator 752 and form second turbo
generator pair 762. Second turbo generator pair 762 will produce
approximately 25 MW of electric power.
[0139] The four stage hybrid cascade heat exchange cycle of FIG. 7
allows the maximum amount of energy to be extracted from the
relatively low temperature differential between the warm sea water
and the cold sea water. Moreover, all heat exchangers can directly
support turbo generator pairs that produce electricity using the
same component turbines and generators.
[0140] It will be appreciated that multiple multi-stage hybrid
cascading heat exchangers and turbo generator pairs can be
incorporated into a vessel or platform design.
[0141] Multi-Stage, Open-Flow, Heat Exchange Cabinets
[0142] OTEC systems, by their nature require large volumes of
water, for example, a 100 megawatt OTEC power plant can require,
for example, up to orders of magnitude more water than required for
a similarly sized combustion fired steam power plant. In an
exemplary implementation, a 25 MW OTEC power plant can require
approximately 1,000,000 gallons per minute of warm water supply to
the evaporators and approximately 875,000 gallons per minute of
cold water to the condensers. The energy required for pumping water
together with the small temperature differentials (approximately 35
to 45 degrees F.) act to drive down efficiency while raising the
cost of construction.
[0143] Presently available heat exchangers are insufficient to
handle the large volumes of water and high efficiencies required
for OTEC heat exchange operations. Shell and tube heat exchangers
consist of a series of tubes. One set of these tubes contains the
working fluid that must be either heated or cooled. The second
non-working fluid runs over the tubes that are being heated or
cooled so that it can either provide the heat or absorb the heat
required. A set of tubes is called the tube bundle and can be made
up of several types of tubes: plain, longitudinally finned, etc.
Shell and tube heat exchangers are typically used for high-pressure
applications. This is because the shell and tube heat exchangers
are robust due to their shape. Shell and tube exchangers are not
ideal for the low temperature differential, low pressure, high
volume nature of OTEC operations. For example, shell and tube heat
exchangers, as shown in FIG. 8, typically require complicated
piping arrangements with high pressure losses and associated piping
energy. These types of heat exchangers are difficult to fabricate,
install and maintain, particularly in a dynamic environment such as
an offshore platform. Shell and tube heat exchanges also require
precision assembly particularly for the shell to tube connections
and for the internal supports. Moreover, shell and tube heat
exchangers often have a low heat transfer coefficient and are
restricted in the volume of water that can be accommodated.
[0144] FIG. 9 depicts a plate heat exchanger. Plate heat exchangers
can include multiple, thin, slightly-separated plates that have
very large surface areas and fluid flow passages for heat transfer.
This stacked-plate arrangement can be more effective, in a given
space, than the shell and tube heat exchanger. Advances in gasket
and brazing technology have made the plate-type heat exchanger
increasingly practical. In HVAC applications for example, large
heat exchangers of this type are called plate-and-frame; when used
in open loops, these heat exchangers are normally of the gasket
type to allow periodic disassembly, cleaning, and inspection.
Permanently-bonded plate heat exchangers, such as dip-brazed and
vacuum-brazed plate varieties, are often specified for closed-loop
applications such as refrigeration. Plate heat exchangers also
differ in the types of plates that are used, and in the
configurations of those plates. Some plates may be stamped with
"chevron" or other patterns, where others may have machined fins
and/or grooves.
[0145] Plate heat exchangers, however, have some significant
disadvantages in OTEC applications. For example, these types of
heat exchangers can require complicated piping arrangements that do
not easily accommodate the large volumes of water needed with OTEC
systems. Often, gaskets must be precisely fitted and maintained
between each plate pair, and significant bolting is needed to
maintain the gasket seals. Plate heat exchangers typically require
complete disassembly to inspect and repair even one faulty plate.
Materials needed for plate heat exchangers are can be limited to
costly titanium and/or stainless steel. These types of heat
exchangers require relatively equal flow areas between the working
and non-working fluids. Flow ratios between the fluids are
typically 1:1. As can be seen in FIG. 9, supply and discharge ports
are typically provided on the face of the plate, reducing the total
heat exchange surface area and complicating the flow path of each
of the working and non-working fluids. Moreover, plate heat
exchangers include complex internal circuiting for nozzles that
penetrate all plates.
[0146] In order to overcome the limitations of such conventional
heat exchangers, a gasket-free, open flow heat exchanger is
provided. In some implementations, individual plates are
horizontally aligned in a cabinet such that a gap exists between
each plate. A flow path for the working fluid runs through the
interior of each plate in a pattern providing high heat transfer
(e.g., alternating serpentine, chevrons, z-patterns, and the like).
The working fluid enters each plate through connections on the side
of the plates so as to reduce obstructions in the face of the plate
or impediments to the water flow by the working fluid. The
non-working fluid, such as raw water, flows vertically through the
cabinet and fills the gaps between each of the open-flow plates. In
some implementations, the non-working fluid is in contact with all
sides of the open-flow plates or in contact with just the front and
back surfaces of the open-flow plates.
[0147] FIG. 10 illustrates a stacked cabinet arrangement 520 of
heat exchangers, similar to the arrangement as described in FIG. 5,
with a detail of a single cabinet 524 having a rack of multiple
heat exchange plates 1022. The non-working fluid flows vertically
through the cabinet 524 and past each of the plates 1022 in the
rack. Arrow 525 indicates the flow direction of the water. The flow
direction of the water can be from top to bottom or bottom to top.
In some embodiments, the flow direction can be in the natural
direction of the water as it is heated or cooled. For example, when
condensing a working fluid, the water can flow through the cabinet
arrangement from bottom to top in the natural flow of convection as
the water is warmed. In another example, when evaporating a working
fluid, the water can flow from top to bottom as the water cools. In
still other embodiments, the non-working fluid flow can be
horizontally across the cabinet, that is, from left to right or
right to left.
[0148] Referring to FIG. 10, open-flow heat exchange cabinet 524
includes cabinet face 1030 and cabinet side 1031. Opposite of
cabinet face 1030 is cabinet face 1032 (not shown) and opposite of
cabinet side 1031 is cabinet side 1033 (not shown). The cabinet
faces and sides form a plenum or water conduit through which the
raw water non-working fluid flows with little to no pressure losses
due to piping. In contrast to the gasket heat exchanger described
above with respect to FIG. 9, the open flow heat exchanger uses the
cabinet to form a flow chamber containing the non-working fluid
(e.g., sea water) rather than using gaskets between plates to form
the flow chamber containing the non-working fluid. Thus, the
open-flow heat exchange cabinet 524 is effectively gasket-free.
This aspect of this system provides significant advantages over
other plate and frame heat exchangers that rely on gaskets to
isolate the working fluid from the energy providing medium (e.g.,
sea water). Corrosion testing of aluminum plate and frame heat
exchangers done at NELHA in the 1980s and 1990s had to stop after
only six months because there was so much leakage around the
gaskets where biological deposits caused extensive erosion. The
applicants identified gasket issues as a major impediment to using
a plate and frame design in an OTEC system.
[0149] In addition, the cabinet approach combined with side mounted
inlet and outlet ports for the heat exchange plates avoids the
needs for the supply and discharge ports typically provided on the
face of the plate heat exchange systems (see, e.g., FIG. 9). This
approach increases the total heat exchange surface area of each
plate as well as simplifying the flow path of both the working and
non-working fluids. Removing the gaskets between the plates also
removes significant obstructions that can cause resistance to flow.
The gasket-free open-flow heat exchange cabinets can reduce back
pressure and associated pumping demand, thus reducing the parasitic
load of an OTEC plant and resulting in increased power that can be
delivered to the utility company.
[0150] In the case of an OTEC condenser, cabinet 524 is open on the
bottom to the cold raw water supply, and open on the top to provide
unobstructed fluid communication with the cabinet 523 above. The
final cabinet in the vertical series 521 is open at the top to the
raw water discharge system.
[0151] In the case of an evaporator, cabinet 521 is open on the top
to the warm raw water supply and open at the bottom to provide
unobstructed fluid communication to the cabinet below 522. The
final cabinet 524 in the vertical series is open on the bottom to
the warm raw water discharge system.
[0152] Within each of the heat exchange cabinets, a plurality of
open-flow heat exchange plates 1022 are arranged in horizontal
alignment to provide a gap 1025 between each pair of plates 1022.
Each open flow plate has a front face, a back face, a top surface,
a bottom surface, and left and right sides. The plates 1022 are
arranged in horizontal alignment so that the back face of a first
plate faces the front face of a second plate immediately behind the
first plate. A working fluid supply and discharge are provided on
the sides of each of the plates to avoid impediments to the flow of
the raw water through the gaps 1025 as the raw water flows past the
front and back faces of the plurality of plates 1022 in the rack.
Each of the plates 1022 includes a working fluid flow passage that
is internal to the plate. Open-flow plates 1022 are described in
greater detail further below.
[0153] In some implementations, each individual plate 1022 has a
dedicated working fluid supply and discharge such that the working
fluid flows through a single plate. Supply of the working fluid is
directly to one or more of working fluid supply passages. In other
implementations, the working fluid can flow through two or more
plates in series before being discharged from the heat exchange
cabinet to the reminder of the working fluid system.
[0154] It will be appreciated that each heat exchanger cabinet 524,
523, 522, and 521 has similar components and is vertically aligned
such that the horizontally aligned plates 1022 in one cabinet
vertically align over the plates in the cabinet below. The gaps
1025 between plates 1022 on one cabinet vertically align over the
gaps 1025 between plates 1022 in the cabinet below.
[0155] Referring to FIGS. 11 and 12, an exemplary implementation of
the plate arrangement in heat exchange cabinet 524 includes a first
open-flow heat exchange plate 1051 having an exterior surface
including at least a front and back face. The exterior surface is
in fluid communication with and surrounded by a non-working fluid
1057 such as cold raw water. The first open flow plate also
includes an internal passage in fluid communication with a working
fluid 1058 flowing through the internal passage. At least one more
second open-flow heat exchange plate 1052 is horizontally aligned
with the first open-flow heat exchange plate 1051 such that the
front exterior surface of the second plate 1052 faces the back
exterior surface of the first plate 1051. Like the first plate, the
at least one second plate 1052 includes an exterior surface in
fluid communication with and surrounded by the non-working fluid
1057, and an internal passage in fluid communication with a working
fluid 1058 flowing through the internal passage. The first
open-flow heat exchange plate 1051 is separated from the second
heat exchange plate 1052 by a gap 1053. The non-working fluid 1057
flows through the gap.
[0156] FIG. 13 depicts a side view of an exemplary open-flow heat
exchange cabinet 524 including a first open flow heat exchange
plate 1051, a second heat exchange plate 1052, and gaps 1053
separating each first plate 1051 and 1052. The working fluid 1058
flows through the internal working fluid flow passages 1055.
[0157] As described above, in some implementations, a single heat
exchange cabinet can be dedicated to a single stage of a hybrid
cascade OTEC cycle. In some implementations, four heat exchange
cabinets are vertically aligned, as depicted and described in FIG.
5. In other implementations, cabinets having working fluid supply
and discharge lines connected to the sides of each plate can be
used. This avoids working fluid conduits being on the face of the
plates and impeding the flow of both the working fluid and the
non-working fluid.
[0158] For example, a gasket-free multi-stage heat exchange system
can include a first stage heat exchange rack comprising one or more
open-flow plates in fluid communication with a first working fluid
flowing through an internal passage in each of the one or more
open-flow plates. The working fluid can be supplied and discharged
from each plate via supply and discharge lines dedicate to each
individual plate. A second stage heat exchange rack vertically
aligned with the first heat exchange rack is also included. The
second stage heat exchange rack comprising one or more open-flow
plates in fluid communication with a second working fluid flowing
through an internal passage in each of the one or more open-flow
plates. Again, the second working fluid is supplied and discharged
to and from each individual plate through lines dedicated to each
individual plate. A non-working fluid, such as raw water, flows
first through the first stage heat exchange rack and around each of
the one or more open-flow plates allowing for thermal exchange with
the first working fluid. The non-working fluid then passes through
the second heat exchange rack and around each of the open-flow
plates allowing for thermal exchange with the second working
fluid.
[0159] The first stage rack includes a plurality of open-flow
plates in horizontal alignment having a gap between each plate. The
second stage rack also includes a plurality of open-flow plates in
horizontal alignment having a gap between each plate within the
second stage racks. The plurality of open-flow plates and gaps in
the second stage rack are vertically aligned with the plurality of
open-flow plates and gaps in the first stage rack. This reduces
pressure losses in the flow of the non-working fluid through the
first and second stage racks. Pressure losses in the non-working
fluid are also reduced by having the non-working fluid directly
discharge from one cabinet to the next thereby eliminating the need
for extensive and massive piping systems. In some embodiments, the
walls of the cabinets containing the first and second stage racks
of heat exchange plates form the conduit through which the
non-working fluid flows.
[0160] Due to the open-flow arrangement of the plates in each rack
of each stage of an exemplary four stage OTEC system, the flow
ratio of the non-working fluid to the working fluid is increased
from the typical 1:1 of most conventional plate heat exchanger
systems. In some implementations the flow ratio of the non-working
fluid is greater than 1:1, (e.g., greater than 2:1, greater than
10:1, greater than 20:1, greater than 30:1, greater than 40:1,
greater than 50:1, greater than 60:1, greater than 70:1, greater
than 80:1, greater than 90:1 or greater than 100:1).
[0161] When a multi-stage arrangement of heat exchange cabinets is
used as a condenser, the non-working fluid (e.g., the cold sea
water) generally enters the first stage cabinet at a temperature
lower than when the non-working fluid enters the second stage
cabinet, and the non-working fluid then enters the second stage
cabinet at a temperature lower than when the non-working fluid
entered the third stage cabinet; and the non-working fluid enters
the third stage cabinet at a temperature generally lower than when
it enters the fourth stage cabinet.
[0162] When a multi-stage arrangement of heat exchange cabinets are
used as an evaporator, the non-working fluid (e.g., the warm sea
water) generally enters the first stage cabinet at a temperature
higher than when the non-working fluid enters the second stage
cabinet, and the non-working fluid then enters the second stage
cabinet at a temperature higher than when the non-working fluid
enters the third stage cabinet; and the non-working fluid enters
the third stage cabinet at a temperature generally higher than when
it enters the fourth stage cabinet.
[0163] When a multi-stage arrangement of heat exchange cabinets are
used as an condenser, the working fluid (e.g., the ammonia)
generally exits the first stage cabinet a temperature lower than
when the working fluid exits the second stage cabinet, and the
working fluid exits the second stage cabinet at a temperature lower
than the working fluid exits the third stage cabinet; and the
working fluid exits the third stage cabinet at a temperature
generally lower than when it exits the fourth stage cabinet.
[0164] When a multi-stage arrangement of heat exchange cabinets are
used as an evaporator, the working fluid (e.g., the ammonia)
generally exits the first stage cabinet at a temperature higher
than the working fluid exiting the second stage cabinet, and the
working fluid exits the second stage cabinet at a temperature
generally higher than the working fluid exits the third stage
cabinet; and the working fluid exits the third stage cabinet at a
temperature generally higher than when it exits the fourth stage
cabinet.
[0165] An exemplary heat balance of an implementation of a four
stage OTEC cycle is described herein and generally illustrates
these concepts.
[0166] In some implementations, a four stage, gasket-free, heat
exchange system includes a first stage heat exchange rack having
one or more open-flow plates, each plate includes an exterior
surface having at least a front and back face surrounded by a
non-working fluid. Each plate also includes an internal passage in
fluid communication with a first working fluid flowing through the
internal passage. The working fluid is supplied and discharged from
each plate by supply and discharge lines dedicated to each
plate.
[0167] The four-stage heat exchange system also includes second
stage heat exchange rack vertically aligned with the first heat
exchange rack, the second stage heat exchange rack includes one or
more open-flow heat exchange plates substantially similar to those
of the first stage and vertically aligned with the plates of the
first stage.
[0168] A third stage heat exchange rack, substantially similar to
the first and second stage racks is also included and is vertically
aligned with the second stage heat exchange rack. A fourth stage
heat exchange rack substantially similar to the first, second and
third stage racks is included and vertically aligned with the third
stage heat exchange rack.
[0169] In operation, the non-working fluid flows through the first
stage heat exchange rack and surrounds each open-flow plate therein
for thermal interaction with the first working fluid flowing within
the internal flow passages of each plate. The non-working fluid
then flows through the second stage heat exchange rack for thermal
interaction with the second working fluid. The non-working fluid
then flows through the second stage heat exchange rack for thermal
interaction with the second working fluid before flowing through
the third stage heat exchange rack for thermal interaction with the
third working fluid. The non-working fluid flows through the third
stage heat exchange rack for thermal interaction with the third
working fluid before flowing through the fourth stage heat exchange
rack for thermal interaction with the fourth working fluid. The
non-working fluid is then discharged from the heat exchange
system.
[0170] Free-Flow Heat Exchange Plates:
[0171] The low temperature differential of OTEC operations
(typically between 35 degrees F. and 85 degrees F.) requires a heat
exchange plate design free of obstructions in the flow of the
non-working fluid and the working fluid. Moreover the plate must
provide enough surface area to support the low temperature lift
energy conversion of the working fluid.
[0172] Conventional power generation systems typically use
combustion process with a large temperature lift system such as a
steam power cycle. As environmental issues and unbalanced fossil
fuel supply issues become more prevalent, Low Temperature Lift
Energy Conversion (LTLEC) systems, such as the implementations of
OTEC systems described herein, and which use renewable energy
sources such as solar thermal and ocean thermal, will become more
important. While conventional steam power cycles uses exhaust gas
from combustion process and are usually at very high temperatures,
the LTLEC cycles use low temperature energy sources ranging from 30
to 100 degrees C. Therefore, the temperature difference between the
heat source and heat sink of the LTLEC cycle is much smaller than
that of the steam power cycle.
[0173] FIG. 14 shows the process of a conventional high temperature
steam power cycle in a pressure-enthalpy (P-h) diagram. Thermal
efficiency of the steam power cycle is in the range of 30 to
35%.
[0174] In contrast, FIG. 15 shows the P-h diagram of an LTLEC
cycle, such as those used in OTEC operations. Typical thermal
efficiency for an LTLEC cycle is 2 to 10%. This is almost one-third
to one-tenth that of a conventional high temperature steam power
cycle. Hence, an LTLEC cycle needs much larger size heat exchangers
than conventional power cycles.
[0175] The heat exchange plates described below provide high heat
transfer performance and also low pressure drop in heat source and
heat sink fluid sides to limit the pumping power requirements which
affect the system efficiency. These heat exchange plates, designed
for OTEC and other LTLEC cycles, can include the following
features:
[0176] 1) A working fluid flow path having a mini-channel design.
This can be provided in a roll-bonded aluminum heat exchange plate
and provides a large active heat transfer area between the working
and non-working fluids;
[0177] 2) A gap provided between plates and/or offsetting the
roll-bond plates between even number and odd number plates so as to
significantly reduce the pressure drop in heat source and heat sink
non-working fluids. In this way, a relatively wide fluid flow area
for heat source and heat sink fluid sides can be provided, while
maintaining a relatively narrow fluid flow area for the working
fluid of the power cycle;
[0178] 3) A configuration of progressively changing channel numbers
per pass within the flow passages of the working fluid can reduce
the pressure drop of the phase-changing working fluid along the
flow. The number of channels in the plate can be designed according
to the working fluid, operating conditions, and heat exchanger
geometry.
[0179] 4) A wavy working fluid flow passages or channel
configuration can enhance the heat transfer performance.
[0180] 5) Within the working fluid flow channels and among parallel
channels, both ends of channel's inner walls of the flow channel
can be curved to smoothly direct the fluid to subsequent channels
when the flow direction is reversed, and non-uniform distances from
the ends of channel's inner walls to the side wall can be used
among parallel channels.
[0181] The above features can reduce the pumping power needed in
the system, and enhance the heat transfer performance.
[0182] Referring again to FIG. 11, mini-channel roll-bonded heat
exchange plates 1051 and 1052 are shown in perspective view. A
cross-counter flow between the working fluid and the non-working
fluid is provided. When used as an evaporator, the non-working
fluid 1057 (e.g., seawater) enters the top of the plates and leaves
from the bottom of the plates. The working fluid 1058 (e.g.,
ammonia) enters the bottom side of the plates in liquid state, and
evaporates and finally becomes vapor phase by absorbing thermal
energy from the higher temperature non-working fluid. The generated
vapor 1059 leaves the plates from the top side.
[0183] FIG. 13 shows fluid flows in a side view. The working fluid
flow channels 1055 have relatively wide width w and relatively low
height h in order to increase the active heat transfer area between
the two fluids while reducing the volume of the entire heat
exchange plate. The width w of the channels can range between about
10 and about 15 mm (e.g., more than 11 mm, more than 12 mm, more
than 13 mm, less than 14 mm, less than 13 mm, and/or less than 12
mm). The height h of the channels can range between about 1 and
about 3 mm (e.g., more than 1.25 mm, more than 1.5 mm, more than
1.75 mm, more than 2 mm, less than 2.75 mm, less than 2.5 mm, less
than 2.25 mm and/or less than 2 mm). The spacing between channels
can be between about 4 and about 8 mm (e.g., more than 4.5 mm, more
than 5 mm, more than 5.5 mm, less than 7.5 mm, less than 7 mm,
and/or less than 6.5 mm). The roll-bonded plates are arranged in an
even plate 1051 and odd plate 1052 distribution with offset working
fluid flow passages 1055 in order to provide a smooth flow path for
the non-working fluid 1057 and provide a wider non-working fluid
flow area than the working fluid flow area in working fluid flow
channels 1055. This arrangement reduces the pressure drop in the
heat source and heat sink fluid sides.
[0184] FIG. 16 illustrates an undulating or wavy working fluid flow
path designed to enhance the heat transfer performance of the
plate.
[0185] FIG. 17 illustrates an embodiment of a heat exchange plate
with two inlets receiving working fluid 1058 and two outlets
discharging heated or cooled fluid 1059. The internal flow paths
within each open-flow plate are arranged in an alternating
serpentine pattern so that the flow of the working fluid is
substantially perpendicular or cross-flow to the flow direction of
the non-working fluid. In addition, the progression of the working
fluid through the serpentine patter can be generally parallel to
the flow of the non-working fluid or opposite the direction of flow
of the non-working fluid. In some embodiments, flow distribution
between channels can be improved by the use of guide vanes. FIG. 18
illustrates an embodiment of a heat exchange plate in which an area
1710 of varying space in the flow path 1701 is provided to even the
flow distribution among parallel channels 1705. Furthermore, both
ends 1715 of the channel's inner walls 1712 are curved to smoothly
direct the fluid to subsequent channels when the flow direction is
reversed, and non-uniform distances from the ends of channel's
inner walls 1712 to the side wall 1702 can be used among parallel
channels. These guide vanes and varying flow path dimensions can be
implemented in heat exchange plates such as, for example, the heat
exchange plates shown in FIGS. 17, 19A and B, and 20A and B.
[0186] In some embodiments, it has been found that the working
fluid changes its phase from liquid to vapor along the flow path,
and consequently the working fluid pressure drop will increase
significantly if the same flow passage area is used throughout the
entire heat exchange plate like. In order to reduce the
fluid-pressure drop increase along the flow associated with its
vapor quality change, the number of parallel flow passages per pass
can be increased along the flow path of the working fluid.
[0187] FIGS. 19A and 19B illustrate a pair of heat exchange plates
1905, 1910 implementing this approach in an evaporator. The heat
exchange plate 1905 in FIG. 19A has two inlets 1911 which each feed
into two mini-channels 1912. The mini-channels 1912 extend along
the plate in a serpentine fashion that is similar to the channels
of the heat exchange plate shown in FIG. 17. However, in the heat
exchange plate shown in FIG. 19A, the flow from two mini-channels
feeds into three mini-channels at a first transition point 1914.
The flow from the three mini-channels feeds into four mini-channels
at a second transition point 1916. As the heat exchange plate
includes two separate, complementary flow paths, these expansions
result in eight mini-channels which discharge through four outlets
1918.
[0188] The four outlets 1918 of the heat exchange plate 1905 are
hydraulically connected to the four inlets 1920 of heat exchange
plate 1910 shown in FIG. 19B. The flow from four mini-channels
feeds into five mini-channels at a third transition point 1922. The
flow from the five mini-channels feeds into six mini-channels at a
fourth transition point 1924. As this heat exchange plate also
includes two separate, complementary flow paths, these expansions
result in twelve mini-channels which discharge through six outlets
1926. Connecting the heat exchange plates 1905, 1910 in series
provides the equivalent of a single long heat exchange plate but is
easier to manufacture.
[0189] The plates 1905, 1910 have a length L of between about 1200
mm and 1800 mm (e.g., more than 1300 mm, more than 1400 mm, more
than 1450 mm, more than 1475 mm, less than 1700 mm, less than 1600
mm, less than about 1550 mm and/or less than 1525 mm). The width W
of the plates can range between about 250 and about 450 mm (e.g.,
more than 275 mm, more than 300 mm, more than 325 mm, more than 350
mm, less than 425 mm, less than 400 mm, less than 375 mm and/or
less than 350 mm).
[0190] In some embodiments, different size plates and different
numbers of inlets and outlets are used to provide the desired heat
exchange area and expansion/contraction characteristics. For
example, the paired plates 1905, 1910 are sized in part based on
the limitations of the current vendor. In some embodiments, a
single plate will replace the paired plates 1905, 1010 thus
removing the need for the outlets 1920 and inlets 1918 that are
used to transfer working fluid from plate 1905 to plate 1910. The
larger plates can have a length L of between about 2700 mm and 3300
mm (e.g., more than 2800 mm, more than 2900 mm, more than 2950 mm,
more than 2975 mm, less than 3200 mm, less than 3100 mm, less than
about 3050 mm and/or less than 3025 mm). The larger plates can have
a width W between about 550 and about 850 mm (e.g., more than 575
mm, more than 600 mm, more than 625 mm, more than 650 mm, less than
825 mm, less than 800 mm, less than 775 mm and/or less than 750
mm). In some embodiments, a single larger inlet 1918 replaces the 2
inlets of plate 1905 and feeds working fluid to all four
mini-channels 1912. Because the inlets 1918 and outlets 1920 can be
sources of head losses that decrease the efficiency of the heat
exchange plates, reducing the number of inlets 1918 and outlets
1920 will reduce the overall pumping requirement and, thus
parasitic load, of a given OTEC system. The flow through heat
exchange plates 1905, 1910 is described for an evaporator. The heat
exchange plates 1905, 1910 could also be used in a condenser.
However, the flow of fluid through a condenser would be the reverse
of the flow described for the evaporator.
[0191] Some heat exchange plates include meandering mini-channels
which can increase residence time for the working fluid (e.g.,
ammonia) passing through the heat exchange plates as well as
providing additional surface area for heat transfer. FIGS. 20A and
20B illustrate a pair of heat exchange plates 2005, 2010 that is
generally similar to the heat exchange plates 1905, 1910 shown in
FIGS. 19A and 19B. However, the mini-channels of heat exchange
plates 2005, 2010 include a meandering pattern. Based laboratory
testing and numerical modeling, the heat exchange plates 2005, 2010
including a sinusoidal meandering pattern are estimated to provide
the same heat exchange as plates 1905, 1910 with an approximately
10% reduction in the number of plates.
[0192] Both plates 1905, 1910 and plates 2005, 2010 include
channels arranged in relatively sinusoidal curve patterns. These
patterns appear to provide several advantages. The relatively
sinusoidal curve patterns cause the water flow over the plates to
take a more turbulent and longer path between the plates enabling
the working fluid (e.g., ammonia) side to theoretically extract
more thermal energy from the water. Moreover, the sinusoidal flow
patterns are configured such that the plates can be turned in
opposite directions or staggered (e.g., alternating left and right)
so that the inlet and outlet fittings do not interfere with each
other.
[0193] Heat exchange plates incorporating the various features
discussed above can be manufactured using a roll-bonded process.
Roll bonding is a manufacturing process by which two metal plates
are fused together by heat and pressure then expanded with high
pressure air so that flow channels are created between the two
panels. A carbon-based material is printed on the bottom panel in
the desired flow pattern. A second panel is then laid atop the
first panel and the two panels are then rolled through a hot
rolling press where the two panels are fused everywhere except
where the carbon material is present. At least one channel is
printed to the edge where a vibrating mandrel is inserted between
the two plates creating a port into which pressurized air is
injected. The pressurized air causes the metal to deform and expand
so that channels are created where the two plates are prevented
from fusing together. There are two ways that roll bonding can be
done: continuous, wherein the metal is run continuously through hot
roll presses off rolls of sheet metal; or discontinuous wherein
precut panels are individually processed.
[0194] In a prototype, two metal sheets, each approximately
1.05-1.2 mm thick, 1545 mm long, and 350 mm wide, were roll-bonded
together to form plates. Channels, in the patterns shown in FIGS.
19A and 19B, were formed between the joined metal sheets by
blow-molding. The channels were formed with a width w of between
12-13.5 mm and a height h of about 2 mm. The plates exhibit good
heat exchange properties using ammonia as the working fluid and
water as the non-working fluid.
[0195] Heat Exchange Plate Examples
[0196] Heat exchangers are large and costly components for OTEC
power production and OTEC water desalination systems. Their size
and performance dominate all other aspects of the plant design.
Although theoretical heat transfer coefficients range from 500 to
3,500 Btu/hft2 R, the actual heat transfer coefficients of the
state-of-the-art plate heat exchangers (Pheat exchanger) under
given operating condition for an OTEC system are only 215-383
Btu/ft2 hrR. This poor performance was mainly due to unbalanced
flow area ratio between ammonia and water sides.
[0197] Laboratory and modeling studies were performed and it was
found that when plates similar to the embodiment depicted in FIG.
19A were arranged as an evaporator having a warm water inlet
temperature of 24.3 degrees C. and an ammonia inlet temperature of
10.7 degrees C., with distance between plates of approximately 6.3
to 6.4 mm, the ammonia quality was maintained between 0.0 and 0.5
and the following performance results were observed as listed in
Table 1 below:
TABLE-US-00002 TABLE 1 HX Plate Performance Results under Varying
Flow Conditions Property Test 1 Test 2 Test 3 Ai (m.sup.2) 3.232
3.232 3.232 Ui (W/m.sup.2k) 3842 3822 4053 Ao (m.sup.2) 5.471 5.471
5.471 Uo (W/m.sup.2k) 2270 2258 2395 DP water (KPa) 3.85 3.94 4.06
DP per length Water (KPa/m) 2.5 2.6 2.6 DP Ammonia (KPa) 10.4 13.2
10.9 MFR Water (kg/s) 11 12.0 13.0 MFR Ammonia (g/s) 31.0 31.0 31.0
Heat Tranfer Coefficient Range of Range of Range of
(BTU/ft.sup.2hrR) 360 to 440 360 to 435 370 to 460 Pressure Drop
(psi/ft) Range of Range of Range of 0.109 to 0.111 to 0.115 to
0.113 0.116 0.119.
In a comparison of rollbonded heat exchange plates and traditional
gasket heat exchange plates, the water flow rate for the rollbonded
design was almost quadrupled while the water side pressure drop was
only one quarter that of the gasketed heat exchange plate design.
The Ui value was also enhanced by approximately 72%. Higher Ui
values can be obtained when the water flow rate is increased.
[0198] In further examples, a combined laboratory/numerical
modeling study was performed to assess novel heat exchanger plate
design concepts in order to reduce heat exchanger surface area and
the associated costs for a planned 100 MW OTEC plant. Initial
assessments of novel heat exchanger were performed and were
followed by testing of several prototype heat exchangers.
[0199] Planned Plant Design and Operating Conditions
[0200] The study estimated required heat exchanger plate surface
area for a planned 100 MWe OTEC based generally on the approach
described with respect to FIGS. 1-4. The layout of evaporators in
the OTEC system for the planned 100 MWe plant consists of 16
chambers as shown in FIG. 21A. Each chamber contains 4 stacks with
each stack including 3 modules with dimensions 3 ft (width) 28 ft
(depth) by 10 ft (height) as shown in FIGS. 21B-21D. As used in the
study, a "cartridge" indicates a single plate; a "cassette"
indicates 2 cartridges forming ammonia flow channel; a "module"
indicates 28' assemblies of cassettes; a "stack" indicates a
four-stage module; and a "chamber" includes 4 stacks.
[0201] The designs were assessed in terms of the plate surface area
and volume with a high heat transfer coefficient and low pressure
drop desired. Each of the four-stage cycles producing 25 MWe are
identical in design and is found to require approximately 10,080
ft.sup.3 of the evaporator volume is required.
[0202] Assumed operating conditions of the second-stage of a
four-stage cycle for 25 MWe were as follows: warm water inlet
temperature--76.3.degree. F.; warm water outlet
temperature--73.4.degree. F.; evaporating pressure--927.3 kPa; Log
Mean Temperature Difference (LMTD)--1.7 K; and warm water flow
rate--1,100,000 gpm.
[0203] Initial Designs
[0204] Five different designs were suggested and investigated. The
flow patterns and pass types are described in Table 2 below. While
a single-pass design in ammonia-side was applied to two base cases,
a multi-pass design in ammonia-side was applied to three designs
(Design A, Design B1, and Design B2). The difference between the
Design A and the Design B is the ammonia channel direction. While
ammonia channels are vertically oriented in the Design A, they are
horizontally oriented in the Design B so that the flow direction is
in cross flow with respect to the water flow.
[0205] Cartridge size, numbers, and heat exchanger volumes were
calculated while maintaining the pressure drop per each stage at 2
psi.
TABLE-US-00003 TABLE 2 Flow pattern and pass type of heat
exchangers Base case - Base case - Design- Design- Item low area
ratio high area ratio A B1 Design-B2 Pass Single Single Multi Multi
Multi Flow Counter Counter Mixed Cross Cross pattern
[0206] FIG. 22A shows a schematic of the base case with low flow
area ratio. The flow area ratio (the ratio between water flow area
and ammonia flow area) is set to 1.47. Ammonia (shown as the dark
lines in the figure) flows upward through the ammonia channels and
water flows downward through the water channels (the gaps between
ammonia channels). FIG. 22B shows a cross-sectional view of the
ammonia channel.
[0207] FIG. 23A shows a schematic of the base case with high flow
area ratio. The flow area ratio is set to 7.14. While ammonia flows
upward through the ammonia channels, water flows downward through
the water channels so that two fluids form the counter flow heat
transfer configuration. FIG. 23B shows a cross-sectional view of
the ammonia channel. In this design, the water channel width was
increased in order to increase the flow area ratio.
[0208] FIG. 24A shows a schematic of Design A. While ammonia flows
through the ammonia channels so that it flows up and down, water
flows only downward over the ammonia channels. FIG. 24B shows a
cross-sectional view of the ammonia channel. In this design, the
ammonia mass flux is increased as compared to the base cases. In
this design, two fluids form counter flow and parallel heat
transfer configurations alternatively.
[0209] FIG. 25A shows a schematic of Design B1. In this design,
ammonia flow channel direction is rotated in 90 from the Design A
so that two fluids form a cross-counter flow heat transfer
configuration. In this design, ammonia flows through the ammonia
channels horizontally and gradually moves up, while water flows
downward over the ammonia channel so that the potential issue of
vapor stagnation on the top horizontal portion of the channels in
the Design A is addressed. FIG. 25B shows a cross-sectional view of
the ammonia channel.
[0210] FIG. 26A shows a schematic of the Design B2 which is also
cross-counter flow as used in the Design B1. FIG. 26B shows a
cross-sectional view of ammonia channel. The main difference
between the Design B1 and Design B2 is the ammonia flow channel
area. The Design B2 has a reduced ammonia flow channel area as
compared to the Design B1. In order to control the ammonia mass
flux in the Design B2, the number of ammonia passes per cassette is
doubled by having two inlets and two outlets (see FIG. 26A).
[0211] Table 3 shows a system level comparison of the five
different initial designs of roll-bonded heat exchangers. For the
base designs, the cartridge size was extremely large due to the
relatively small overall heat transfer coefficient (U) mainly due
to the low ammonia heat transfer coefficient (HTC). Therefore, the
base designs needed a large plate area to provide the required
total heat transfer capacity. While the heat exchanger surface
areas of the base cases are in an unacceptable range, the heat
exchanger surface areas of Design A, Design B1 and Design B2 are
less than one million ft.sup.2, which is in the acceptable range
for the planned OTEC system. In Design A heat exchangers, the
ammonia vapor can possibly be trapped in the horizontal portion of
channels potentially causing significant problems. The Design B2
heat exchangers provided the required total heat transfer capacity
with a smaller footprint and lower cost than the Design B1 heat
exchangers. Accordingly, the Design B2 heat exchanger design was
chosen as a final design.
TABLE-US-00004 TABLE 3 Comparison of Initial Designs Base case -
Base case- high low flow flow area Design- Design- Parameters Unit
area ratio ratio Design-A B1 B2 Water-side Btu/ft.sup.2hrR 545 480
941 1,171 1,191 HTC Ammonia Btu/ft.sup.2hrR 116 379 1,870 3,917
7,619 HTC U Btu/ft.sup.2hrR 95 210 611 871 990 Cartridge ft .times.
ft 3 .times. 34.58 3 .times. 67.05 3 .times. 10.67 3 .times. 10 3
.times. 9.5 Size No. of EA 40,778 28,692 20,584 16,300 16,296
cartridge heat Ft.sup.3 83,272 113,616 12,970 10,585 7,391
exchanger Volume Plate Ft.sup.2 4,767,465 6,198,791 742,576 483,959
464,550 surface area Comments Unacceptable results Infeasible OK OK
design * Water-side HTC was obtained from CFD.
[0212] Prototype Evaporator Design for Laboratory Testing
[0213] The evaluation of the initial design alternatives led to the
selection of Design B2 as the flow pattern to be used in prototype
heat exchangers for laboratory testing. Several additional designs
based on the Design B2 were also tested.
[0214] The original cartridge size was 3 by 10 ft. However, in
order to evaluate these heat exchangers in the laboratory test
facility, the heat exchanger scale has to be cut down due to the
height limitations of the test facility.
[0215] The typical way to downscale an object undergoing hydraulic
testing is using "similarity analysis". This method relies on
maintaining dimensional aspect ratios as well as maintaining
thermal properties and dimensionless numbers such as Reynolds
number and Prandtl number. However, the similarity analysis
approach could not be used for this application as the plate
thickness for the heat exchangers is fixed at 1.5 mm based on a
minimum required 1.0 mm for manufacturing and ammonia gap size
cannot be reduced further. Accordingly, the following alternative
method was used.
[0216] As mentioned above, the original cartridge size was 3 by 10
ft. The test cartridge size was selected as a quarter size of the
original cartridge size based on test facility constraints (see
FIGS. 27A and 27B). For the test heat exchanger, the water
temperature change across the heat exchanger was reduced
proportionally to the cartridge height reduction, and the ammonia
vapor quality change was also reduced proportionally at the same
evaporating temperature. Ammonia and water Reynolds numbers were
kept constant.
[0217] The prototype heat exchange was designed with four cassettes
to minimize edge effects.
[0218] FIG. 28 illustrates pattern 1 which included 40 ammonia
channels and 2 dummy channels. The figure is rotated
counter-clockwise by 90 degrees, so that left hand side of the
figure is the top of the heat exchanger and right hand side is the
bottom. Ammonia entered though the bottom and exited from the top.
The ammonia mass flux was designed to be constant along the flow
path. Two dummy channels were designed to form water channels.
Pressure in the ammonia channel was measured through the dummy
channels. The dummy channels and associated pressure transducers
are for testing purposes and will not be included in production
models.
[0219] FIG. 29A illustrates pattern 2 which included four ammonia
channels connected together. For the better flow distribution among
parallel ammonia channels, the cross-sectional area of header
section is designed to be varied along the flow as previously
discussed with respect to FIG. 18. The ammonia mass flux was
designed to be uniform throughout the heat exchanger. In an
alternate ammonia channel design, the corners are rounded, and
guide vanes are applied in order to minimize the pressure drop due
to sharp corners as shown in FIG. 29B.
[0220] FIG. 30A illustrates pattern 3 which included an increasing
number of ammonia channels along the flow path such that, as the
ammonia vapor quality increases, the ammonia pressure drop
increases as well. Therefore, the ammonia mass flux was designed to
be varied to the vapor quality with high ammonia mass flux at low
vapor quality and low ammonia mass flux at high vapor quality. In
an alternate ammonia channel design, the corners are rounded, and
guide vanes are applied in order to minimize the pressure drop due
to sharp corners as shown in FIG. 29B.
[0221] FIG. 31 illustrates pattern 4. Several surface enhancements
such as dimples, offsets, and a wavy pattern were considered to
improve the performance of the heat exchanger. These kinds of
surface enhancement options are anticipated to increase the system
heat transfer coefficients but also to increase the pressure drop.
When the pressure drop was regulated to the target value, the heat
transfer coefficient value can be higher or lower than that of
pattern 3. A wavy pattern was applied in the pattern 4 and its
performance characteristics were evaluated through the experimental
test.
[0222] FIG. 32 illustrates pattern 5 which included an offset
pattern applied to the ammonia channel as an option for the surface
enhancement. As same as for the pattern 4, the relation between the
heat transfer coefficient and pressure drop were investigated
through the experimental test.
[0223] The performance and physical properties of suggested
patterns are investigated for the OTEC system. Table 3 shows the
summary of different pattern designs for one evaporator of
four-stage cycle for 25 MWe. The water channel gap and number of
plates were calculated such that the HX stack depth does not exceed
28' limitation.
[0224] For the pattern 1, the U value of the evaporator was
calculated at 990 Btu/ft2 hrR due to the high ammonia heat transfer
coefficient. However, the ammonia-side pressure drop was predicted
as 975 psi, which is even larger than the absolute working
pressure. Thus, the ammonia pressure drop would be the limiting
factor to make this pattern feasible.
[0225] The pattern 2 was suggested with the intention of reducing
the large ammonia-side pressure drop of the pattern 1. Ammonia-side
pressure drop through the evaporator was reduced to less than 50
kPa by regulating the ammonia mass flux. The U value was calculated
as 660 Btu/ft2 hrR. The plate number and evaporator volume of one
of four-stage cycle for 25 MW were calculated as 22,250 EA and
9,526 ft3, respectively.
[0226] The U value of the pattern 3 was calculated as 629 Btu/ft2
hrR, which is lower than that of the pattern 2 by 5%, while the
ammonia-side pressure drop was calculated as 38 kPa that is smaller
than that of the pattern 2 by 16%.
[0227] For patterns 4 and 5, water and ammonia heat transfer
coefficient were assumed to be increased by 5% as compared to the
pattern 3. When these values are increased, the U value increases
too. This results in a smaller active heat transfer area needed, so
that plate numbers can be reduced and, thus, the water and ammonia
mass fluxes can be increased. While the water-side HTC is increased
by 5%, the ammonia-side HTC is increased by 13.5% due to the
increased ammonia mass flux and surface enhancement. Finally, the U
value was estimated to 688 Btu/ft2 hrR.
[0228] Table 5 shows the number of plates and evaporator volume for
the pattern 2 design for the 100 MW OTEC system as an example.
TABLE-US-00005 TABLE 4 Calculated Heat Exchanger Properties
Parameter Unit Pattern 1 Pattern 2 Pattern 3 Pattern 4/5 Water HTC
W/m.sup.2K 6,760 6,339, 6,380 6,699 Btu/ft.sup.2hrR 1,191 1,116
1,124 1,180 Ammonia HTC W/m.sup.2K 43,261 9,768 8,833 10,026
Btu/ft.sup.2hrR 7,619 1,720 1,556 1,766 U value W/m.sup.2K 5,622
3,746 3,572 3,909 Btu/ft.sup.2hrR 990 660 629 688 Ammonia mass flux
kg/m.sup.2s 588.7 86.3 134 -> 58 149 -> 64 Lb/ft.sup.2s 121
18 27 -> 12 31 -> 13 Ammonia pressure kPa 6,032 45.5 38.2 46
drop in evaporator Psi 875 6.6 5.5 6.7 Plate numbers EA 16,296
22,248 23,784 21,520 Evaporator volume ft.sup.3 7,391 9,528 9,956
9,320 (one of four-stage cycle for 25 MWe) Plate size ft by ft 3 by
9.5 3 by 10 3 by 10 3 by 10 Cassette numbers per EA/ft 24.3 33.1
35.4 32.0 unit length Comments -- Infeasible Feasible OK Estimated
Design design *: Water-side HTC was obtained from CFD.
TABLE-US-00006 TABLE 5 Estimated Evaporators with Pattern 2 Designs
for 100 MWe OTEC Plant 16 Chambers One One One 4 chambers for
Parameters Module Stack Chamber for 25 MWe 100 MWe No. of plate
1,854 5,562 22,248 88,992 355,968 (EA) No. of 927 2,781 11,124
44,496 177,984 cassette (EA) Evaporator 794 2,382 9,528 38,112
152,448 volume (ft.sup.3)
[0229] Table 6 shows the measured heat exchanger properties.
Ammonia-side HTC was calculated as an average value between
Kandlikar (1990) and Shah (1982) correlations, and ammonia-side
pressure drop was calculated as an average value between Lockhard
and Martinelli (1949) and Friedel (1980) correlations. Eight
cartridges with dimensions of 1.5 by 5 ft were to be used for the
prototype evaporators.
TABLE-US-00007 TABLE 6 Measured Heat Exchanger Properties Parameter
Unit Pattern 1 Pattern 2 Pattern 3 Pattern 4/5 Water HTC W/m.sup.2K
6,760 6,339, 6,380 6,699 Btu/ft.sup.2hrR 1,191 1,116 1,116 1,180
Ammonia HTC W/m.sup.2K 34,297 9,301 8,337 8,985 Btu/ft.sup.2hrR
6,040 1,638 1,468 1,582 U value W/m.sup.2K 5,437 3,675 3,515 3,740
Btu/ft.sup.2hrR 958 647 619 659 Ammonia mass flux kg/m.sup.2s 440.8
81 162 -> 54 167 -> 56 Lb/ft.sup.2s 90 16.6 33 -> 11 34
-> 11 Ammonia pressure kPa 1,908 25.1 20 21.7 drop in evaporator
Psi 277 3.6 3 3.15 Plate size ft by ft 1.5 by 5 1.5 by 5 1.5 by 5
1.5 by 5 Plate numbers EA 8 8 8 8 Comments -- One long 4 channels
Connected Surface channel connected channel Enhanced numbers
increased *: Water-side HTC was obtained from CFD.
[0230] Computational Fluid Dynamics Simulation on Water-Side
[0231] CFD simulation was conducted on the water-side of rollbond
heat exchanger, in order to investigate the heat transfer
coefficient and pressure drop characteristics. Several different
designs were evaluated.
[0232] FIG. 33 shows the dual gap design in ammonia side. For the
base case with high flow are ratio, low fin efficiency resulted in
needing many number of plates and huge heat exchanger volume.
Therefore, in order to increase the fin-efficiency high, the
concept of increasing ammonia gap came up and evaluated. The
detailed geometry of the proposed concept is shown in Table 7.
TABLE-US-00008 TABLE 7 Detailed geometry of dual gap design in
ammonia side Parameter A b c D t Unit Mm mm mm mm mm Value 10 0.5
0.75 0.5 0.5
[0233] Table 8 shows the CFD result of dual gas design in ammonia
side. Basically, it shows high water HTC and fin efficiency.
However, the stress analysis results showed that the design could
not stand the ammonia pressure since the gap (d) was relatively
small and the width (a) was large. Therefore, it was concluded that
this design was infeasible.
TABLE-US-00009 TABLE 8 CFD result of dual gap design in ammonia
side Water HTC DP Mass flux Efficiency Re SI unit W/m2K Pa/m
kg/m.sup.2s -- -- 10,721 5,682 2,915 94% 76,242 IP unit
BTU/hft.sup.2.degree. F. Psi/ft Lb/ft.sup.2s -- -- 1,888 0.25 597
94% 76,242
[0234] The detailed geometry of the Design B1 is shown in FIG. 34.
This is a cross counter flow design. Several different gap sizes
were evaluated the heat exchanger. Since it is a cross counter flow
design, the channel pattern was designed and then the ammonia-side
heat transfer coefficient and pressure drop were determined. Then,
the heat exchangers with different water gap sizes were evaluated
for the water-side.
[0235] The CFD simulation results of different gap sizes are shown
in Table 9. Since cartridge size was fixed to 3 by 10 ft, the
pressure drop was maintained at 0.2 psi/ft. Water heat transfer
coefficient ranged from 1,000 to 1,200 Btu/ft.sup.2Rhr.
TABLE-US-00010 TABLE 9 CFD result of Design B1 Water Mass Fin HTC
DP flux Efficiency Re g = SI unit W/m.sup.2K Pa/m kg/m.sup.2s -- --
5.3 mm 5,974 4,575 799 96% 4,400 IP unit BTU/ psi/ft Lb/ft.sup.2s
-- -- hft.sup.2 .degree. F. 1,052 0.2 164 96% 4,400 g = SI unit
W/m.sup.2K Pa/m kg/m.sup.2s -- -- 8.7 mm 6,009 4,614 860 96% 10,652
IP unit BTU/ psi/ft Lb/ft.sup.2s -- -- hft.sup.2 .degree. F. 1,058
0.20 176 96% 10,652 g = SI unit W/m.sup.2K Pa/m kg/m.sup.2s -- --
9.4 mm 6,757 4,542 1,108 96% 14,782 IP unit BTU/ Psi/ft
Lb/ft.sup.2s -- -- hft.sup.2 .degree. F. 1,190 0.2 227 96%
14,782
[0236] The detailed geometry of the Design B2 is shown in FIG. 35.
Compared to the Design B1, the ammonia channel gap height of Design
B2 was reduced from 6.5 to 20.0 mm. This approach reduces the
water-side pressure drop. Therefore, the water mass flux can be
increased in order to enhance the water-side HTC.
[0237] Table 10 shows the CFD result of the Design B2 while the
pressure drop was maintained at 0.2 psi/ft for all cases. The
water-side HTC ranged from 1,200 to 1,400 Btu/ft.sup.2Rhr.
TABLE-US-00011 TABLE 10 CFD result of Design B2 Water Mass Fin HTC
DP flux Efficiency Re g = 5 mm SI unit W/m.sup.2K Pa/m kg/m.sup.2s
-- -- 6,663 4,585 1,732 94% 12,349 IP unit BTU/ psi/ft Lb/ft.sup.2s
-- -- hft.sup.2 .degree. F. 1,173 0.203 355 94% 12,349 g = 6 mm SI
unit W/m.sup.2K Pa/m kg/m.sup.2s -- -- 7,131 4,593 1,936 94% 16,520
IP unit BTU/ psi/ft Lb/ft.sup.2s -- -- hft.sup.2 .degree. F. 1,256
0.203 397 94% 16,520 g = 7 mm SI unit W/m.sup.2K Pa/m kg/m.sup.2s
-- -- 7,488 4,518 2,096 94% 20,829 IP unit BTU/ Psi/ft Lb/ft.sup.2s
-- -- hft.sup.2 .degree. F. 1,319 0.2 429 94% 20,829 g = 8 mm SI
unit W/m.sup.2K Pa/m kg/m.sup.2s -- -- 7,948 4,454 2,226 94% 25,238
IP unit BTU/ Psi/ft Lb/ft.sup.2s -- -- hft.sup.2 .degree. F. 1,400
0.197 456 94% 25,238
[0238] Additional OTEC Features:
[0239] In an exemplary implementation of an OTEC power plant, an
offshore OTEC spar platform includes four separate power modules,
each generating about 25 MWe Net at the rated design condition.
Each power module comprises four separate power cycles or cascading
thermodynamic stages that operate at different pressure and
temperature levels and pick up heat from the sea water system in
four different stages. The four different stages operate in series.
The approximate pressure and temperature levels of the four stages
at the rated design conditions (Full Load--Summer Conditions)
are:
TABLE-US-00012 Turbine inlet Condenser Pressure/Temp.
Pressure/Temp. (Psia)/(.degree. F.) (Psia)/(.degree. F.) 1.sup.st
Stage 137.9/74.7 100.2/56.5 2.sup.nd Stage 132.5/72.4 93.7/53
3.sup.rd Stage 127.3/70.2 87.6/49.5 4.sup.th Stage 122.4/68
81.9/46
[0240] The working fluid is boiled in multiple evaporators by
picking up heat from warm sea water (WSW). Saturated vapor is
separated in a vapor separator and led to an ammonia turbine by STD
schedule, seamless carbon steel pipe. The liquid condensed in the
condenser is pumped back to the evaporator by 2.times.100% electric
motor driven constant speed feed pumps. The turbines of cycle-1 and
4 drive a common electric generator. Similarly the turbines of
cycle-2 and 3 drive another common generator. In some embodiments,
there are two generators in each plant module and a total of 8 in
the 100 MWe plant. The feed to the evaporators is controlled by
feed control valves to maintain the level in the vapor separator.
The condenser level is controlled by cycle fluid make up control
valves. The feed pump minimum flow is maintained by recirculation
lines led to the condenser through control valves regulated by the
flow meter on the feed line.
[0241] In operation, the four (4) power cycles of the modules
operate independently. Any of the cycles can be shut down without
hampering operation of the other cycles if needed, for example in
case of a fault or for maintenance. Such partial shut downs will
reduce the net power generation of the overall power module.
[0242] The system requires large volumes of seawater and includes
separate systems for handling cold and warm seawater, each with its
pumping equipment, water ducts, piping, valves, heat exchangers,
etc. Seawater is more corrosive than fresh water and all materials
that may come in contact with it need to be selected carefully
considering this. The materials of construction for the major
components of the seawater systems will be:
[0243] Large bore piping: Fiberglass Reinforced Plastic (FRP)
[0244] Large seawater ducts & chambers: Epoxy-coated carbon
steel
[0245] Large bore valves: Rubber lined butterfly type
[0246] Pump impellers: Suitable bronze alloy
[0247] Unless controlled by suitable means, biological growths
inside the seawater systems can cause significant loss of plant
performance and can cause fouling of the heat transfer surfaces
leading to lower outputs from the plant. This internal growth can
also increase resistance to water flows causing greater pumping
power requirements, lower system flows, etc. and even complete
blockages of flow paths in more severe cases.
[0248] The Cold Sea Water ("CSW") system using water drawn in from
the deep ocean should have very little or no bio-fouling problems.
Water in those depths does not receive much sunlight and lacks
oxygen, and so there are fewer living organisms in it. Some types
of anaerobic bacteria may, however, be able to grow in it under
some conditions. Shock chlorination will be used to combat
bio-fouling.
[0249] The Warm Sea Water ("WSW") system handling warm seawater
from near the surface will have to be protected from bio-fouling.
It has been found that fouling rates are much lower in tropical
open ocean waters suitable for OTEC operations than in coastal
waters. When necessary, chemical agents can be used to control
bio-fouling in OTEC systems at very low doses that will be
environmentally acceptable. Dosing of small amounts of chlorine has
proved to be very effective in combating bio-fouling in seawater.
Dosages of chlorine at the rate of about 70 ppb for one hour per
day, is quite effective in preventing growth of marine organisms.
This dosage rate is only 1/20th of the environmentally safe level
stipulated by EPA. Other types of treatment (thermal shock, shock
chlorination, other biocides, etc.) can be used from time to time
in-between the regimes of the low dosage treatment to get rid of
chlorine-resistant organisms.
[0250] Necessary chlorine for dosing the seawater streams is
generated on-board the plant ship by electrolysis of seawater.
Electro-chlorination plants of this type are available commercially
and have been used successfully to produce hypochlorite solution to
be used for dosing. The electro-chlorination plant can operate
continuously to fill-up storage tanks and contents of these tanks
are used for the periodic dosing described above.
[0251] The seawater conduits are designed to avoid any dead pockets
where sediments can deposit or organisms can settle to start a
colony. Sluicing arrangements are provided from the low points of
the water ducts to blow out deposits that may get collected there.
High points of the ducts and water chambers are vented to allow
trapped gases to escape.
[0252] The Cold Seawater (CSW) system will consist of a common deep
water intake for the plant ship, and water pumping/distribution
systems, the condensers with their associated water piping, and
discharge ducts for returning the water back to the sea. The cold
water intake pipe extends down to a depth of more than 2700 ft,
(e.g., between 2700 ft to 4200 ft), where the sea water temperature
is approximately a constant 40.degree. F. The entrance to the pipe
is protected by screens to stop large organisms from being sucked
in to it. After entering the pipe, cold water flows up towards the
sea surface and is delivered to a cold well chamber near the bottom
of the vessel or spar.
[0253] The CSW supply pumps, distribution ducts, condensers, etc.
are located on the lowest level of the plant. The pumps take
suction from the cross duct and send the cold water to the
distribution duct system. 4.times.25% CSW supply pumps are provided
for each module. Each pump is independently circuited with inlet
valves so that they can be isolated and opened up for inspection,
maintenance, etc. when required. The pumps are driven by
high-efficiency electric motors.
[0254] The cold seawater flows through the condensers of the cycles
in series and then the CSW effluent is discharged back to the sea.
CSW flows through the condenser heat exchangers of the four plant
cycles in series in the required order. The condenser installations
is arranged to allow them to be isolated and opened up for cleaning
and maintenance when needed.
[0255] The WSW system comprises underwater intake grills located
below the sea surface, an intake plenum for conveying the incoming
water to the pumps, water pumps, biocide dosing system to control
fouling of the heat transfer surfaces, water straining system to
prevent blockages by suspended materials, the evaporators with
their associated water piping, and discharge ducts for returning
the water back to the sea.
[0256] Intake grills are provided in the outside wall of the plant
modules to draw in warm water from near the sea surface. Face
velocity at the intake grills is kept to less than 0.5 ft/sec. to
limit entrainment of marine organisms. These grills also prevent
entry of large floating debris and their clear openings are based
on the maximum size of solids that can pass through the pumps and
heat exchangers safely. After passing through these grills, water
enters the intake plenum located behind the grills and is routed to
the suctions of the WSW supply pumps.
[0257] The WSW pumps are located in two groups on opposite sides of
the pump floor. Half of the pumps are located on each side with
separate suction connections from the intake plenum for each group.
This arrangement limits the maximum flow rate through any portion
of the intake plenum to about 1/16th of the total flow and so
reduces the friction losses in the intake system. Each of the pumps
is provided with valves on inlet sides so that they can be isolated
and opened up for inspection, maintenance, etc. when required. The
pumps are driven by high-efficiency electric motors with variable
frequency drives to match pump output to load.
[0258] It is necessary to control bio-fouling of the WSW system and
particularly its heat transfer surfaces, and suitable biocides will
be dosed at the suction of the pumps for this.
[0259] The warm water stream may need to be strained to remove the
larger suspended particles that can block the narrow passages in
the heat exchangers. Large automatic filters or `Debris Filters`
can be used for this if required. Suspended materials can be
retained on screens and then removed by backwashing. The
backwashing effluents carrying the suspended solids will be routed
to the discharge stream of the plant to be returned to the ocean.
The exact requirements for this will be decided during further
development of the design after collection of more data regarding
the seawater quality.
[0260] The strained warm seawater (WSW) is distributed to the
evaporator heat exchangers. WSW flows through the evaporators of
the four plant cycles in series in the required order. WSW effluent
from the last cycle is discharged at a depth of approximately 175
feet or more below the sea surface. It then sinks slowly to a depth
where temperature (and therefore density) of the seawater will
match that of the effluent.
[0261] Additional Aspects:
[0262] The baseline cold water intake pipe is a staved, segmented,
pultruded fiberglass pipe. Each stave segment can be 40-50' long.
Stave segments can be joined by staggering staves to create an
interlocking seam. Pipe staves can be extruded in panels up to
52-inches wide and at least 50-feet in length and can incorporate
e-glass or s-glass with polyurethane, polyester, or vinylester
resin. In some aspects, the stave segments can be concrete. Staves
can be solid construction. The staves can be a cored or honeycombed
construction. The staves will be designed to interlock with each
other and at the ends of the staves will be staggered there by
eliminating the use of flanges between sections of the cold water
pipe. In some embodiments, the staves can be 40-ft long and
staggered by 5-ft and 10-ft where the pipe sections are joined. The
staves and pipe sections can be bonded together, e.g., using
polyurethane or polyester adhesive. 3-M and other companies make
suitable adhesives. If sandwich construction is used, polycarbonate
foam or syntactic foam can be used as the core material. Spider
cracking is to be avoided and the use of polyurethane helps to
provide a reliable design.
[0263] In some embodiments, the envisioned CWP is continuous, i.e.
it does not have flanges between sections.
[0264] The CWP can be connected to the spar via a spherical bearing
joint. The cold water pipe can also be connected to the spar using
a combination of lifting cables and a ram or dead-bolt system.
[0265] One of the significant advantages of using the spar buoy as
the platform is that doing so results in relatively small rotations
between the spar itself and the CWP even in the most severe
100-year storm conditions. In addition, the vertical and lateral
forces between the spar and the CWP are such that the downward
force between the spherical ball and its seat keeps the bearing
surfaces in contact at all times. This bearing, which also acts as
the water seal, does not come out of contact with its mating
spherical seat. Thus, there is no need to install a mechanism to
hold the CWP in place vertically. This helps to simplify the
spherical bearing design and also limits the pressure losses that
would otherwise be caused by any additional CWP pipe restraining
structures or hardware. The lateral forces transferred through the
spherical bearing are also low enough that they can be adequately
accommodated without the need for vertical restraint of the
CWP.
[0266] Though embodiments herein have described multi-stage heat
exchanger in a floating offshore vessel or platform, it will be
appreciated that other embodiments are within the scope of the
invention. For example, the multi-stage heat exchanger and
integrated flow passages can be incorporated into shore based
facilities including shore based OTEC facilities. Moreover, the
warm water can be warm fresh water, geo-thermally heated water, or
industrial discharge water (e.g., discharged cooling water from a
nuclear power plant or other industrial plant). The cold water can
be cold fresh water. The OTEC system and components described
herein can be used for electrical energy production or in other
fields of use including: salt water desalination: water
purification; deep water reclamation; aquaculture; the production
of biomass or biofuels; and still other industries.
[0267] All references mentioned herein are incorporated by
reference in their entirety.
[0268] Other embodiments are within the scope of the following
claims.
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