U.S. patent application number 10/314135 was filed with the patent office on 2004-06-10 for simple and compact low-temperature power cycle.
This patent application is currently assigned to TENNESSEE VALLEY AUTHORITY. Invention is credited to Crim, Michael C., McClanahan, Timmons S..
Application Number | 20040107700 10/314135 |
Document ID | / |
Family ID | 32468425 |
Filed Date | 2004-06-10 |
United States Patent
Application |
20040107700 |
Kind Code |
A1 |
McClanahan, Timmons S. ; et
al. |
June 10, 2004 |
SIMPLE AND COMPACT LOW-TEMPERATURE POWER CYCLE
Abstract
A simple, compact, and relatively efficient thermodynamic power
cycle system and process for extracting heat from a heat source
stream and converting a portion of the heat to mechanical power.
The system and process are composed of the same series of four
processing units or steps found in the most basic form of a Rankine
power cycle: (1) heating (means) of a pressurized working fluid to
produce a superheated gas, (2) expansion (means) to a lower
pressure to produce power, (3) condensation (means) of the low
pressure gas to a liquid, and (4) pumping (means) of the liquid to
high pressure to complete the cycle. The working fluid is heated
under pressures above critical. The working fluid must have a
critical temperature more than 40.degree. F. lower than the
temperature of the heat source stream and a normal boiling point
less than 32.degree. F.
Inventors: |
McClanahan, Timmons S.;
(Florence, AL) ; Crim, Michael C.; (Florence,
AL) |
Correspondence
Address: |
SUGHRUE MION, PLLC
2100 PENNSYLVANIA AVENUE, N.W.
SUITE 800
WASHINGTON
DC
20037
US
|
Assignee: |
TENNESSEE VALLEY AUTHORITY
|
Family ID: |
32468425 |
Appl. No.: |
10/314135 |
Filed: |
December 9, 2002 |
Current U.S.
Class: |
60/670 ;
60/676 |
Current CPC
Class: |
F01K 25/106 20130101;
F01K 25/08 20130101 |
Class at
Publication: |
060/670 ;
060/676 |
International
Class: |
F01K 001/00; F01K
013/00 |
Claims
What is claimed:
1. A thermodynamic power cycle system for extracting a flow of heat
from a heat source stream and generating mechanical power from the
flow of heat by means of a working fluid flowing within a
closed-loop cycle comprising: means for transferring heat from the
beat source stream to the working fluid such that the working fluid
warms from a first temperature to a second temperature that is more
than 30.degree. F. greater than the critical temperature of the
working fluid wherein the working fluid has a critical temperature
more than 40.degree. F. lower than a temperature of the heat source
stream and has a normal boiling point less than 32.degree. F.;
means for expanding the working fluid and converting work of
expansion of the working fluid to mechanical power; said means for
expanding and converting work of expansion also throttling the
working fluid such that a pressure of the working fluid exceeds the
critical pressure of the working fluid by an amount greater than 5%
of the critical pressure of the working fluid as the working fluid
emerges from the means for transferring heat; means for cooling to
condense and subcool the working fluid after the means for
expanding; means for returning the working fluid to the means for
transferring heat; wherein the means for transferring heat, the
means for expanding, the means for cooling, and the means for
returning the working fluid to the means for transferring heat are
the only four means in which energy is removed from or transferred
into the working fluid in the form of heat or work.
2. The thermodynamic power cycle system of claim 1, wherein the
heat source stream comprises a gas, liquid, solid or mixture
thereof.
3. The thermodynamic power cycle system of claim 1, further
comprising an additional means for throttling the working fluid
after the means for transferring heat and before the means for
expanding.
4. The thermodynamic power cycle system of claim 1, further
comprising means for controlling the flow rate of the working
fluid.
5. The thermodynamic power cycle system of claim 1, further
comprising means for containing excess of the working fluid in the
liquid state after the means for cooling to condense the working
fluid.
6. The thermodynamic power cycle system of claim 1, further
comprising means for redirecting the flow of the working fluid
after the working fluid has exited the means for transferring heat
to bypass the means for expanding, the means for redirecting the
flow containing a means for throttling the working fluid such that
a pressure of the working fluid exceeds the critical pressure of
the working fluid by an amount greater than 5% of the critical
pressure of the working fluid as the working fluid emerges from the
means for transferring heat.
7. The thermodynamic power cycle system of claim 1, which comprises
two or more heat source streams, with the thermodynamic power cycle
system comprising additional means for transferring heat, each of
which means for transferring heat is dedicated to a single heat
source stream; wherein the working fluid is divided into separate
streams, with each of the separate streams of working fluid being
dedicated to a separate means for transferring heat; and wherein
the separate streams of working fluid, after having been heated by
transfer of heat from the heat source streams, are combined into a
single working fluid stream.
8. The thermodynamic power cycle system of claim 2, further
comprising means for increasing the pressure of the heat source
stream to restore pressure lost by the heat source stream as it
flows through the means for transferring heat.
9. The thermodynamic power cycle system of claim 8, wherein the
heat source stream is a gas, and wherein the thermodynamic power
cycle system further comprises ducting means to transport the gas
to the means for transferring heat and wherein the means for
increasing the pressure is a fan or compressor.
10. The thermodynamic power cycle system of claim 1, wherein the
working fluid is ammonia, chlorodifluoromethane, sulfur dioxide or
bromotrifluoromethane.
11. The thermodynamic power cycle system of claim 1, wherein the
working fluid is ammonia.
12. The thermodynamic power cycle system of claim 1, wherein the
working fluid is chlorodifluoromethane.
13. The thermodynamic power cycle system of claim 1, wherein the
working fluid is sulfur dioxide.
14. The thermodynamic power cycle system of claim 1, wherein all
means of the system except for the means for transferring heat are
mounted on one or more portable transportation means.
15. The thermodynamic power cycle system of claim 14, wherein the
mechanical power is 4 MW or less and wherein there is only one
portable transportation means.
16. The thermodynamic power cycle system of claim 15, wherein the
working fluid is ammonia.
17. The thermodynamic power cycle system of claim 15, wherein the
working fluid is chlordifluoromethane.
18. The thermodynamic power cycle system of claim 1, wherein the
heat source stream is a gas which contains a condensable vapor.
19. The thermodynamic power cycle system of claim 18, wherein the
heat source stream is a stream of pressurized hot gas which has
been quenched in water.
20. The thermodynamic power cycle system of claim 19, wherein the
stream of pressurized hot gas has been produced by the reaction of
coal and oxygen in a coal gasifier.
21. The thermodynamic power cycle system of claim 20, wherein the
working fluid is ammonia.
22. The thermodynamic power cycle system of claim 20, wherein the
mechanical power is utilized to provide supplemental drive power to
an air compressor of an air separation unit, the air separation
unit being employed to provide oxygen to the coal gasifier.
23. The thermodynamic power cycle system of claim 22, wherein the
working fluid is ammonia.
24. The thermodynamic power cycle system of claim 1, wherein the
heat source stream is generated by a topping cycle.
25. The thermodynamic power cycle system of claim 24, wherein the
topping cycle comprises a combustion turbine and wherein the heat
source stream is exhaust gas from a combustion turbine.
26. The thermodynamic power cycle system of claim 25, wherein the
combustion turbine is a peaking unit.
27. The thermodynamic power cycle system of claim 26, wherein the
working fluid is sulfur dioxide.
28. The thermodynamic power cycle system of claim 26, wherein the
working fluid is ammonia.
29. The thermodynamic power cycle system of claim 25, wherein the
exhaust gas has been partially cooled by heat exchange with
compressed air.
30. The thermodynamic power cycle system of claim 29, wherein the
working fluid is ammonia.
31. A thermodynamic process for the production of mechanical power
from a heat source stream of gas, liquid solid, or mixture thereof
comprising: a. transferring heat from the heat source stream to a
working fluid; wherein the working fluid is at a pressure more than
5 percent greater than the critical pressure of the working fluid;
wherein the working fluid has been heated to a temperature more
than 30.degree. F. greater than the critical temperature of the
working fluid; and wherein the working fluid has a critical
temperature more than 40.degree. F. lower than the temperature of
the heat source stream and the working fluid has a normal boiling
point less than 32.degree. F.; b. expanding the working fluid to
produce mechanical power; c. cooling to condense and subcool the
working fluid; d. pressurizing the working fluid; e. directing the
flow of the working fluid in a continuous loop through the above
described process steps a, b, c, d, in that order, and returning to
step a to continue the continuous loop; f. process steps a, b, c,
and d being the only four process steps in which energy is removed
from or transferred into the working fluid in the form of heat or
work.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] The field of the invention is a thermodynamic power cycle in
which the working fluid (also called motive fluid) is energized by
externally applied heat. Within that field, the present invention
is a process in which the working fluid in the course of power
production reaches a pressure and a temperature above that at which
its vapor and liquid have the same density (i.e., a fluid state
that is above its critical pressure and temperature, also called
supercritical conditions). The present invention is also a process
in which the working fluid is other than water or steam. Ammonia is
the working fluid of choice for most applications contemplated for
the present invention, but other fluid types which, like ammonia,
have boiling points below 32.degree. F. at a pressure of one
atmosphere, absolute, may also be selected for reasons of obtaining
greater efficiency, operability, or economics.
[0003] 2. Description of the Prior Art
[0004] Many industrial processes have flowing streams of liquids,
solids, or gases that contain heat which must be exhausted to the
environment or removed in some way to facilitate proper operation
of the process. Typically, the process designer for these
industrial processes will use heat exchange devices to capture the
heat and recycle it back into the process via other process
streams. Often, however, there are not streams suitable to capture
and recycle this heat, because they are either already too high in
temperature or they contain insufficient mass flow. Any heat which
cannot be recycled into the process is typically referred to as
waste heat. Most often waste heat is simply discharged to the
environment, either directly as an exhaust stream, or indirectly
via a cooling medium, such as cooling water.
[0005] One method of utilizing waste heat is to raise steam in a
boiler to drive a turbine, a known method well recognized by
practitioners of the art known as the Rankine cycle. The
steam-based Rankine cycle, however, is only economic when it is
applied to heat source streams that are relatively high in
temperature (generally 600.degree. F. or higher) or are large in
overall heat content. In other words, high thermal efficiency or
significantly large scale is generally needed to make the Rankine
cycle economic. A major reason for this is that efficient removal
of waste heat from a process stream requires boiling water at
multiple pressures/temperatures to capture heat at multiple
temperature levels as the heat source stream is cooled. This
complexity is costly from standpoints of both equipment cost and
operating labor. Overall, the steam-based Rankine cycle is either
too expensive or too inefficient or some combination of the two to
be applied to streams of small flow rate and/or
low-temperature.
[0006] Some process developers have substituted other working
fluids for steam in the Rankine cycle to obtain greater
compatibility with heat source streams of low or moderate
temperature. Typically, an organic fluid such as propane is used.
Although improved over steam, organic cycles present the same
fundamental inadequacies of the Rankine cycle described above.
[0007] Accordingly, there is a need for a relatively simple,
low-cost, and relatively efficient method of capturing and
utilizing waste heat from process streams that are low in
temperature or low in overall heat content.
[0008] The advantages of using supercritical conditions in a power
cycle have been recognized for many years. For example, a 1927
patent (U.S. Pat. No. 1,632,575, Jun. 14, 1927, Abendroth)
describes a system for generating power from supercritical steam.
Even then the inventor, Abendroth, did not claim supercritical
steam power generation as the invention, rather he claimed a
variation of it.
[0009] Abendroth, '575, above, highlighted the advantages of
supercritical steam generation when he stated, "The advantage of
this process resides in the fact that a separation of steam and
liquid of equal temperatures but of different physical properties
cannot take place at any point in the process. In this way the
dangers are eliminated which are caused by the well-known
ebullition or boiling phenomena." In other words, a heated fluid
does not boil when it is at a pressure above critical, instead the
fluid simply transitions from liquid to vapor as its temperature
rises through the critical temperature. Indeed, the properties of
the liquid and the vapor are identical at the critical temperature.
And, although the dangers of boiling are well-understood and easily
controlled in today's power plants, boiling requires specialized
equipment to separate the liquid phase from the vapor phase. Under
supercritical conditions, in which no such separation takes place,
the equipment is simplified. Moreover, as will be explained later
in detail, supercritical operation can have thermal efficiency
advantages over boiling operation. Generally, with most working
fluids, multiple pressure boiling stages are needed to achieve the
same thermal efficiency that supercritical operation can achieve in
one stage.
[0010] A disadvantage of the supercritical steam cycle is that the
heat source must be above 705.degree. F., the critical temperature
of water. This eliminates many moderate and low temperature heat
sources as potential applications for the cycle. A supercritical
ammonia cycle, however, is applicable to these h eat sources
because of the relatively low critical temperature of ammonia,
270.degree. F.
[0011] The use of ammonia as a working fluid is also known. In U.S.
Pat. No. 781,481, Jan. 31, 1905, Windhausen, Jr., a basic form of
Rankine cycle is described in which ammonia is the working fluid.
The patent covers generally all pure working fluids in which their
normal boiling temperature is less than 32.degree. F. These "low
boiling" fluids are, in general, a good match for Rankine cycle
applications in which the heat source temperature and/or the
condensing temperature is relatively low. Since Windhausen's patent
was first issued in 1905, there have been variations of using
ammonia and other low boiling liquids to capture low temperature
heat. Many of these were patented during the late 1970's and early
1980's at the height of the energy crisis in the U.S. Exemplary of
these is an invention to convert natural heat sources (solar,
geothermal, etc.) to power (U.S. Pat. No. 4,100,744, Jul. 18, 1978,
DeMunari), an invention to produce power from low temperature heat
sources in a petroleum refinery (U.S. Pat. No. 4,109,469, Aug. 29,
1978, Carson), an invention to exploit natural temperature
differences on the earth such as a mountain top and a desert valley
(U.S. Pat. No. 3,953,971, May 4, 1976, Parker), and an invention
that is another variation of the use of natural heat sources (U.S.
Pat. No. 4,192,145, Mar. 11, 1980, Tanaka).
[0012] In 1969, William L. Minto discussed the value of low boiling
compounds as working fluids in his patent of a Low Entropy Engine
(U.S. Pat. No. 3,479,817, Nov. 25, 1969, Minto). Minto's engine was
in essence a basic Rankine cycle with certain low boiling
compounds, mainly halogenated hydrocarbon compounds such as carbon
tetrachloride, as the universe of working fluids from which to
select. (One example of an acceptable working fluid for Minto's
invention, chlorodifluoromethane, is specifically cited as a
potential selection of working fluid for our invention.) Minto's
engine vaporizes the working fluid by ordinary means of boiling at
subcritical pressure. In his patent, Minto stated that low boiling
compounds could endow a power cycle with certain characteristics
when he stated that an "object of the present invention is to
provide an improved [working fluid] . . . characterized by its
efficiency, simplicity, [and] compactness . . . " . Minto did not
recognize, however, that the use of supercritical operating
conditions could further increase thermal efficiency of the process
and enhance simplicity by eliminating boiling and the attendant
boiler equipment.
[0013] Only one patent was discovered which mentioned the concept
of using ammonia as a working fluid with supercritical operating
conditions. In U.S. Pat. No. 3,986,362, Oct. 19, 1976, Baciu, a
process is described in which power is generated from a geothermal
heat source. The process employs a two step heat transfer
arrangement in which hot water from the geothermal heat source
warms an intermediate heat storage material, liquid sodium, which
in turn heats a pressurized working fluid stream of ammonia to
supercritical conditions. The working fluid ammonia is expanded and
then reheated by a second two step heat transfer arrangement like
that just described. This reheat method increases the efficiency of
the cycle but also makes it more complex in that the energy in the
heat source stream must be divided in some manner to provide heat
to both the primary, or high pressure, turbine's working fluid and
to the reheat, or low pressure, turbine's working fluid. The cycle
includes the generation of low temperature thermal energy as well
as electrical energy. The patent claims make no reference to
supercritical operation nor to ammonia as the working fluid,
although from the detailed description presumably, a supercritical
ammonia cycle is a preferred embodiment of the process.
[0014] Outside of patent literature, a 1993 paper describes a
supercritical ammonia cycle as a simpler and more efficient
alternative to both the multi-pressure steam cycle and to the
Kalina cycle, which uses variable composition mixtures of ammonia
and water in a subcritical cycle [Solomon D. Tetelbaum,
"Comparative Characteristics of the High Efficiency Supercritical
Bottoming Cycle," American Society of Mechanical Engineers,
IGTI-Vol. 8, 1993, p p. 445-452]. The main application for the
cycle described by Tetelbaum was as a bottoming cycle to capture
waste heat from the high temperature (>900.degree. F.) exhaust
leaving a combustion turbine. Because the working fluid enters the
expansion turbine at a high temperature, near that of the
combustion turbine's exhaust gas, the working fluid leaving the
cycle's expansion turbine is also relatively hot (about 400.degree.
F.) and thus contains a significant quantity of sensible heat. To
make effective use of this sensible heat, a regenerative method is
used in which the condensed ammonia from the recirculation pump is
pre-heated by indirect heat exchange with the hot ammonia exhaust.
Similar to the reheat method described above, the regenerative
method increases efficiency but adds complexity and cost to the
cycle.
SUMMARY OF THE INVENTION
[0015] The present invention is a thermodynamic power cycle system
and process to be used for the purpose of extracting a flow of heat
from a hot stream of gas, liquid, solid, or mixture of these,
hereafter referred to as a heat source stream, and converting a
portion of this extracted flow of heat to mechanical power. Heat is
extracted from the heat source stream to a working fluid in a heat
exchanger in which the two streams flow in opposite
(countercurrent) direction. The working fluid flows through the
heat exchanger at a pressure above its critical pressure and
emerges above its critical temperature. The use of supercritical
conditions permits the working fluid to transition from liquid to
gas at its critical temperature without boiling. (Or, stated
another way, the liquid and gas at the critical temperature have
identical properties and are indistinguishable.) This supercritical
step simplifies the equipment compared with Rankine cycle by
eliminating the boiler section. The boiler section of a Rankine
cycle normally contains added equipment of such size and complexity
that small heat source streams cannot be practically or
economically utilized.
[0016] The present invention in its simplest form is comprised of
four means to perform four process steps through which the working
fluid flows in a closed-loop cycle: (1) means for transferring heat
from the heat source stream to the working fluid, (2) means for
expanding the working fluid to generate mechanical power and the
means for expanding also providing a means for throttling the
working fluid to maintain a pressure within the means for
transferring heat that is greater than the critical pressure of the
working fluid, (3) means for cooling to condense and subcool the
working fluid after the means for expanding, and (4) means for
returning the working fluid to the means for transferring heat.
These are the only four means or steps in the system or process in
which energy is added to or removed from the working fluid in the
form of heat or work. Since the means used in the system of the
present invention so closely parallel the process steps of the
present invention, the same will often be treated
interchangeably.
[0017] Other steps typically used in Rankine cycle, which are often
needed to promote good thermal efficiency and which have the
drawback of increasing cost and operating complexity, such as
multiple stages of boiler pressures, reheat of turbine exhaust, and
deaeration, are not necessary in the present invention.
[0018] The present invention, by virtue of the properties of the
working fluid and the nature of the process flow scheme, is simple,
compact, and relatively efficient for heat sources of low to
moderate temperature (about 210.degree. F. and higher). This
combination permits heat to be economically converted to power from
heat source streams which are relatively low in heat content, i.e.,
either low in temperature or mass flow or a combination of the two.
One way to express this combination numerically is by using the
concept of availability of energy to do work based on the second
law of thermodynamics. By this definition, the present invention is
economically viable using heat source streams with as low as about
5 million Btu per hour of available heat content (per unit of
time). Such streams of low available heat content are not normally
economic when used as heat sources with Rankine cycle.
[0019] The working fluid of choice is ammonia for those
applications in which the heat source stream is in the temperature
range of about 400.degree. F. to about 700.degree. F. This is the
range contemplated for many applications of the present invention.
However, for heat source temperatures outside this range or for
unusual applications within this range, other fluid types may be
selected for use in the present invention for reasons of obtaining
greater efficiency, operability, or economics.
[0020] It is the principal object of this invention to provide a
low-cost, simple to operate, relatively efficient, and compact
system and process for the conversion of a flow of heat to
power.
DETAILED OBJECTS OF THE INVENTION
[0021] It is an object of this invention to provide a system and
process which are capable of economically capturing heat from more
than one heat source stream within the same facility.
[0022] Another object of this invention is to permit, if desired by
the user, all of the system and processing equipment (except for
the heat transfer means and the heat source stream) to be mounted
on one or more portable transportation means.
[0023] A related object of this invention is to permit, for a
system and process unit of about 4 MW or less in net power output,
all of the system and processing equipment (except for the heat
transfer means and the heat source stream) to be located on a
single portable transportation means.
[0024] A still further object of this invention is to permit
portable transportation means units to be designed and constructed
according to a standardized set of specifications.
[0025] It is a further object of this invention to provide a system
and process which are sufficiently simple in operation such that it
can be started and operated automatically under normal or routine
circumstances of operation with the use of conventional computer
controls, without benefit of human intervention.
[0026] Another object of this invention is to provide a system and
process with relatively fast startup capability.
[0027] Another object of this invention is to provide a system and
process for making power, which, by virtue of their inherent
simplicity, may be applied to existing heat source streams within
an existing facility without significant change to the operation of
the existing facility.
[0028] It is an object of this invention to provide a system and
process for converting the latent heat of condensing vapors within
a heat source stream into power.
[0029] A further object of this invention and a more specific
application is to utilize this invention to simplify the heat
recovery from a hot gas generated by a coal gasifier, and in
particular, the type of coal gasifier in which the hot coal gas is
quenched in water.
[0030] Another object of this invention is that it may be applied
as a bottoming cycle to convert waste heat from a topping cycle
into power.
[0031] It is a further object of this invention that it may be
applied to the hot exhaust gas from a combustion turbine system
(i.e., application as a bottoming cycle with a combustion turbine
as the topping cycle).
[0032] A further and more specific object of this invention is that
it may be applied to recuperated combustion turbine units, i.e.,
combustion turbine units in which the hot exhaust has been
partially cooled by heat exchange with combustion air.
[0033] A still further and more specific object of this invention
is that it may be applied to combustion turbines used for the
generation of peak loads, i.e., special purpose combustion
turbines, also called peaking units, which are employed to rapidly
provide electric power to a power transmission grid during
intermittent periods in which electric power demand is unusually
high.
[0034] Still further and m ore general objects and advantages oft
he present invention will appear from the detailed description set
forth below, it being understood, however, that this more detailed
description is given by way of illustration only, and not
necessarily by way of limitation since various changes therein may
be made by those skilled in the art without departing from the true
spirit and scope of the present invention.
[0035] Thus, in one aspect of the invention there is provided a
thermodynamic power cycle system for extracting a flow of heat from
a heat source stream and generating mechanical power from the flow
of heat by means of a working fluid flowing within a closed-loop
cycle comprising:
[0036] means for transferring heat from the heat source stream to
the working fluid such that the working fluid warms from a first
temperature to a second temperature that is more than 30.degree. F.
greater than the critical temperature of the working fluid wherein
the working fluid has a critical temperature more than 40.degree.
F. lower than the temperature of the heat source stream and has a
normal boiling point less than 32.degree. F.;
[0037] means for expanding the working fluid and converting work of
expansion of the working fluid to mechanical power; said means for
expanding and converting work of expansion also throttling the
working fluid such that a pressure of the working fluid exceeds the
critical pressure of the working fluid by an amount greater than 5%
of the critical pressure of the working fluid as the working fluid
emerges from the means for transferring heat;
[0038] means for cooling to condense and subcool the working fluid
after the means for expanding;
[0039] means for returning the working fluid to the means for
transferring heat;
[0040] the means for transferring heat, the means for expanding,
the means for cooling, and the means for returning being the only
four means in which energy is removed from or transferred into the
working fluid in the form of heat or work.
IN PREFERRED ASPECTS OF THE INVENTION
[0041] The heat source stream comprises a gas, liquid, solid or
mixture thereof.
[0042] There is further provided: an additional means for
throttling the working fluid after the means for transferring heat
and before the means for expanding; means for controlling the flow
rate of the working fluid; means for containing excess of the
working fluid in the liquid state after the means for cooling to
condense the working fluid; means for redirecting the flow of the
working fluid after the working fluid has exited the means for
transferring heat to bypass the means for expanding, the means for
redirecting the flow containing a means for throttling the working
fluid such that a pressure of the working fluid exceeds the
critical pressure of the working fluid by an amount greater than 5%
of the critical pressure of the working fluid as the working fluid
emerges from the means for transferring heat.
[0043] A means for increasing the pressure of the heat source
stream to restore the pressure lost by the heat source stream as it
flows through the means for transferring heat is provided.
[0044] Further preferred aspects of the invention include
embodiments where: there exists two or more heat source streams,
with the thermodynamic power cycle system comprising additional
means for transferring heat, each of which means for transferring
heat is dedicated to a single heat source stream; wherein the
working fluid is divided into separate streams, with each of the
separate streams being dedicated to a separate means for
transferring heat; and wherein the separate streams of working
fluid, after having been heated by transfer of heat from the heat
source streams, are combined into a single working fluid
stream;
[0045] the heat source stream is a gas, and the system further
comprises ducting means to transport the gas to the means for
transferring heat and wherein the means for increasing the pressure
is a fan or compressor;
[0046] the working fluid is ammonia, chlorodifluoromethane, sulfur
dioxide or bromotrifluoromethane;
[0047] the working fluid is ammonia;
[0048] the working fluid is chlorodifluoromethane;
[0049] the working fluid is sulfur dioxide.
[0050] Still further preferred aspects of the invention include
embodiments where: all means of the system except for the means for
transferring heat are mounted on one or more portable
transportation means;
[0051] the mechanical power is 4 MW or less and wherein there is
only one transportation means;
[0052] the working fluid is ammonia for the embodiment wherein
there is only one transportation means;
[0053] the working fluid is chlorodifluoromethane for the
embodiment wherein there is only one transportation means;
[0054] And still further aspects of the invention include
embodiments where: the heat source stream is a gas which contains a
condensable vapor;
[0055] the heat source stream is as tream of pressurized hot gas
which has been quenched in water;
[0056] the stream of pressurized hot gas has been produced by the
reaction of coal and oxygen in a coal gasifier;
[0057] the working fluid is ammonia for the embodiment wherein the
stream of pressurized hot gas has been produced by the reaction of
coal and oxygen in a coal gasifier:
[0058] the mechanical power is utilized to provide supplemental
drive power to an air compressor of an air separation unit, the air
separation unit being employed to provide oxygen to the coal
gasifier;
[0059] the working fluid is ammonia for the embodiment wherein the
mechanical power is utilized to provide supplemental drive power to
an air compressor of an air separation unit, the air separation
unit being employed to provide oxygen to the coal gasifier
[0060] And still further preferred aspects of the invention include
embodiments where: the heat source stream is generated by a topping
cycle;
[0061] the topping cycle a comprises a combustion turbine and
wherein the heat source stream is exhaust gas from the combustion
turbine;
[0062] the combustion turbine is a peaking unit;
[0063] the working fluid is ammonia for the embodiment wherein the
combustion turbine is a peaking unit;
[0064] the working fluid is sulfur dioxide for the embodiment
wherein the combustion turbine is a peaking unit;
[0065] the exhaust gas has been partially cooled by heat exchange
with compressed air;
[0066] the working fluid is ammonia for the embodiment wherein the
exhaust gas has been partially cooled by heat exchange with
compressed air.
BRIEF DESCRIPTION OF THE DRAWINGS
[0067] FIG. 1 is a flow diagram of the present invention.
[0068] FIG. 2 is a graph of the heat source stream temperature and
the working fluid stream temperature during heat exchange in
Example I.
[0069] FIG. 3 is a graph of the heat source stream temperature and
the working fluid stream temperature during heat exchange in
Example II.
[0070] FIG. 4 is a graph of the heat source stream temperature and
the working fluid stream temperature during heat exchange in
Example V.
DETAILED DESCRIPTION OF THE INVENTION
[0071] Although similar to the earlier discussed prior art
supercritical cycles, the present invention and the practice
thereof relates to a heretofore unrecognized, new, and novel
approach to the use of supercritical fluids for the generation of
power. The physical properties of some fluids are such that a
relatively large amount of power can be produced from a relatively
small volumetric flow, thus creating a process for generating power
that is relatively small and low in cost. With
chlorodifluoromethane as the working fluid, power can be obtained
from heat source streams with temperatures as low as about
290.degree. F. With ammonia as the working fluid, power can be
obtained from heat source streams across a broad range of moderate
temperatures, about 400.degree. F. to about 700.degree. F. That
which has not been recognized by other inventors is that the unique
properties of these fluids and others permit a process to be
designed that is simple, compact, and relatively low in cost
without sacrificing to any large degree the thermal efficiency of
the process. This characteristic simplicity and compactness permits
the process flow scheme, most of the equipment, and the process
operation to be standardized regardless of the application.
Standardizing the design, particularly if the design is simple and
compact as the present invention offers, has a distinct economic
advantage. Our studies show that a small, 2-4 MW generic version of
the present invention, in which most of the process equipment is
assembled in a shop and skid-mounted, is very economically
attractive as a source of electric power. A small skid-mounted
process such as this would permit many heretofore unusable heat
source streams that are low or moderate in temperature or low in
total heat content to be exploited for power production. Shop
assembly and skid-mounting of the equipment into an integrated,
transportable unit are techniques that provide significant cost
savings over custom-built, field-assembled processes.
[0072] Another benefit of a simple, standardized process is that,
with conventional computer-driven process controls, most functions
of the process can be automated. This reduces operating labor costs
and allows the present invention to be applied at sites where the
operating staff is unfamiliar with power generation technology.
[0073] As earlier indicated, the present invention in its simplest
form is comprised of four system means or process steps through
which the working fluid flows in a closed-loop cycle: (1) heat
transfer means, (2) expansion means, (3) cooling means, and (4)
means for returning the working fluid to the heat transfer means.
Also, optionally included in the present invention to facilitate
control and operation of the cycle are the following means and
corresponding process steps: (1) storage capacity means to hold an
excess of liquid working fluid, (2) throttling means (in addition
to that provided by the expanding means) to maintain supercritical
pressure in the heat exchanger, (3) bypass means and pressure
control means to allow the working fluid to bypass the expansion
step during startup and shutdown, (4) pressure elevating means
(e.g., a fan, compressor, or pump) to restore the pressure lost by
the heat source stream as it flows through the heat transfer means,
and (5) flow controlling means to control the circulation rate of
the working fluid.
[0074] The present invention, together with further objectives and
advantages thereof, will be better understood from a consideration
of the following description taken in connection with the
accompanying drawings and examples.
[0075] FIG. 1 is a process flowsheet generally illustrating the
principles of the present invention. The process operates in a
closed-loop cycle. Operation in a closed-loop cycle in the context
of the present invention simply means that the working fluid flows
in a continuous manner through the elements of the system when the
process is in operation and does not leave the process loop which
includes heat transfer means, throttling means, expanding means,
cooling means and return means. While theoretically it would be
possible to temporarily interrupt the process of the present
invention so that flow is not continuous, at present we see no
benefit to interrupting such continuous flow except when the
process is to be off-line. In similar fashion, working fluid could
be removed from the system during times of closed-loop cycle at
normal operation, but at present we see no benefit to such removal
of working fluid. However, the claims of the present application
should be construed in a manner which would not exclude such
temporary interruption of the closed-loop cycle or minor removal or
addition of working fluid from the closed-loop.
[0076] In heat exchanger 101, heat is transferred via indirect heat
exchange from the heat source stream, 102, to the working fluid
stream, 103. Stream 103 is at a temperature near ambient, having
been cooled at a previous point in the cycle by an external cooling
medium. Stream 103 enters 101 as a liquid at a pressure greater
than its critical pressure and emerges as a supercritical vapor
stream 104. Pressure control valve 105 maintains the pressure in
heat exchanger 101 above critical during periods of operation when
the flow rate of the working fluid is less than the flow prescribed
by design. During periods of operation when the flow rate is equal
to or greater than the flow rate prescribed by design, this valve
will be fully open. The working fluid enters expansion turbine 106
to produce work of expansion to drive electric generator 107 or,
alternatively, to drive some other device requiring power such as a
pump or a compressor. Stream 108 exits expansion turbine 106,
having been cooled by work of expansion, and is condensed
substantially completely, more preferably completely, to a liquid
in condenser 109. Condenser 109 also subcools the liquid by about
2-5.degree. F. below saturation to increase the net positive
suction head to the circulating pump 111. An external cooling
medium (not shown), such as water or air generally used by
industrial facilities, is used to condense and subcool the working
fluid. Vessel 110 provides storage capacity for an excess of the
working fluid beyond that which is the minimum volume needed for
the present invention to function. Pump 111 imparts flow and
pressure to the working fluid to return the working fluid to heat
exchanger 101. A flow control device, 112, such as a valve or a
speed controller for the pump sets the circulation flow rate of the
working fluid. A bypass line and pressure control valve, 113, is
included to permit expansion turbine 106 to be bypassed during
startup, shutdown or other times of major transition in the
operation of the present invention. The cooled heat source stream,
114, leaving heat exchanger 101, is directed back to the process
from which the heat source stream came, or, alternatively, stream
114 is directed to device 115, which is a fan, compressor, pump or
other means of elevating the pressure to restore the pressure lost
by the heat source stream as it flows through heat exchanger
101.
[0077] The present invention is applicable to heat source streams
containing a liquid, gas, solid, or a mixture of these. The term
gas is meant to also include vapor, which wholly or in part
condenses into a liquid as the heat source stream is cooled. Vapor
as a heat source is a particularly important application for the
present invention as is illustrated in Example II presented later.
Solids are preferably applicable in a mixture with gases or
liquids, because, by strict definition, a pure solid cannot flow as
a stream. However, a pure solid that generates a flow of heat by
means of internal reaction (such as a nuclear reactor core) or that
receives a flow of heat from an external heat source is a
potentially applicable form of heat source for the present
invention and such a solid is therefore included within the
definition of heat source stream.
[0078] Two important operating parameters for the present invention
are the pressure and temperature o f stream 104, the working fluid
as it leaves heat exchanger 101. This pressure and temperature must
exceed the working fluid's critical pressure and temperature,
respectively, during normal operation. (The critical temperature of
a gas is the temperature above which it is impossible to liquefy
the gas no matter how high the applied pressure. The critical
pressure is the minimum pressure needed to liquefy a gas at its
critical temperature. The term supercritical means above or higher
than critical; and the term "supercritical conditions" means that
the fluid is in a state in which both the critical pressure and
critical temperature are exceeded.) To avoid unstable operation,
which may occur as a result of large fluctuations in density of the
working fluid whenever stream 104 is too close to the critical
pressure, the pressure of stream 104 should be at least 5 percent
greater than the critical pressure, as measured in absolute
numbers, and the temperature of stream 104 should be at least
30.degree. F. greater than the critical temperature.
[0079] The working fluid may b e any fluid having a critical
temperature at least 40.degree. F. less than the temperature of the
heat source stream and having a normal boiling point of 32.degree.
F. or less. (Normal boiling point is defined as the temperature at
which the fluid boils when the fluid pressure is 14.696 psia,
standard atmospheric pressure.) The required 40.degree. F. minimum
difference between the critical temperature and the temperature of
the heat source stream is based on maintaining the required minimum
30.degree. F. temperature difference between the critical
temperature and the temperature of the working fluid, stream 104,
as described above, plus a minimum additional 10.degree. F.
difference for heat transfer to occur between stream 104 and the
heat source stream, stream 102. A maximum 32.degree. F. normal
boiling point assures that the condensing pressure is significantly
above atmospheric pressure (i.e., at least 5 psi above atmospheric
at 50.degree. F., the lowest contemplated temperature of
condensation), thus ensuring that the present invention will retain
certain important features that meet the objects of the invention,
namely, compactness of the equipment and the lack of a vacuum
anywhere within the working fluid during operation of the present
invention.
[0080] Within the constraints of the above limitations, the type of
working fluid to use in the present invention is a choice of the
plant designer. As those skilled in the art of plant design will
appreciate, such a choice will vary depending on parameters of the
application such as temperature of the heat source stream, mass
flow of the heat source stream, and temperature of the cooling
medium. To put this choice in the broadest perspective, the choice
of working fluid is simply one of many design options that a
designer will consider in seeking a plant that is the most economic
for the application.
[0081] For purposes of illustrating the effects of specific working
fluids on the present invention and its practice, all numerical
references regarding specific working fluids discussed herein are
for the theoretically pure fluid. These numerical references are
presented by way of illustration only and are not meant to imply
that the fluid must be pure in order to be acceptable for use as a
working fluid in the present invention. The working fluid used in
the present invention need not be at the extremely high purity
level of a chemical reagent, but generally speaking the process of
the present invention will be most practical with a working fluid
which is about as high or higher than the purity of a standard
commercial grade of the chemical or working fluid involved. The
primary reason for this is that this will permit the inevitable
loss of working fluid over time to be replenished without worrying
about constantly changing the nature of the basic chemical
composition of the working fluid and such a commercial grade of
chemical is easily available. However, the working fluid of the
present invention is selected primarily with regard to the physical
characteristics and properties that the working fluid must exhibit
in the closed-loop cycle of the present invention, and thus in
theory there is no reason why blends of various working fluids
could not be used to achieve the objects of the present
invention.
[0082] One working fluid that our computer simulations have shown
to produce excellent economic results across a broad range of heat
source temperatures (generally 400.degree. F. to 700.degree. F.)
and sizes (about 1 MW and larger) is ammonia. In order that those
skilled in the art may better understand how the properties of the
working fluid contribute to fulfilling the objects of the
invention, an analysis of ammonia as a working fluid is presented
below. As those skilled in the art will readily understand, the
numerical results of this analysis will differ for other types of
working fluids, but the general principles will still apply.
[0083] The present invention can, of course, be practiced without
the use of computer simulations by using principles well known in
the art. In fact, process simulation programs can be easily written
without undue experimentation, though commercially available
process simulation software such as ChemCAD.RTM. is readily
available and simple to use.
[0084] The low critical temperature of ammonia, 270.degree. F.,
allows the process to function at low temperature and to utilize
low or moderate temperature heat source streams. During heat
transfer from the heat source stream to the working fluid, by
maintaining the pressure of ammonia above its critical pressure,
1636 psia, no boiling occurs inside the heat exchanger. Instead the
working fluid's temperature rises continuously as heat is added.
The physical state of the fluid transitions from liquid to vapor as
the temperature of the fluid rises through its critical temperature
of 270.degree. F. This rise in temperature with heat input is close
to linear, unlike a boiling process (or Rankine cycle) in which the
fluid rises in temperature until it reaches the boiling point, then
remains at a constant temperature as heat is input until all of the
liquid has been converted to vapor, and then resumes rising in
temperature as the vapor is superheated. As those skilled in the
art of heat exchange are aware, a nearly linear rise of temperature
of the working fluid suppresses the formation of undesirable pinch
points, i.e., points of zero temperature difference between the two
fluids being exchanged. Generally, with boilers, effective and
efficient utilization of the available heat from the heat source
requires multiple stages of boiling to avoid encountering pinch
points in the design.
[0085] The pressure at which ammonia condenses is relatively high
compared with many other working fluids, which contributes to the
compactness of the process. For example, at a condensing
temperature of 85.degree. F., ammonia condenses at a pressure of
167 psia and steam condenses in a deep vacuum, about 0.6 psia. The
higher condensing pressure of ammonia suppresses the relative
volumetric flow of the working fluid, which, in turn, lowers the
relative size of the expansion turbine and condenser equipment.
[0086] By way of technical definition, the terms expansion or
expanding in the context of an expansion turbine or a means for
expanding refers to the expansion of the working fluid to produce
work and the conversion of that work to mechanical power. During
expansion, the working fluid cools as heat is converted to
mechanical power. (It should be noted that, strictly speaking, work
and heat are forms of energy, not power, which is a flow of energy.
However, those skilled i n the art understand that because the
working fluid flows, the work or heat derived from the working
fluid also flows and is therefore properly expressed as power.) The
term throttling, while similar in some aspects to expanding, has a
distinct and separate definition as used herein. Throttling is the
use of a narrowing or restriction in the flow path of the working
fluid which acts to increase or maintain the pressure upstream of
the restriction. For any given flow of the working fluid, an
increase in throttling, which is to say a greater restriction to
the flow, will increase the pressure upstream of the throttling
device. As those skilled in the art are familiar, any expansion
turbine will provide both work of expansion to produce power and a
throttling effect, which maintains pressure of the working fluid
upstream of the expansion turbine. However, a working fluid can
also be throttled without producing work of expansion by simply
restricting the flow with a valve, such as the use of valve 105 in
the present invention. Further discussion of valve 105 regarding
its purpose and relationship to expansion turbine 106 can be found
later in this section.
[0087] As with many steam cycle applications, the present invention
can be and typically would be designed and operated such that the
working fluid is cooled by expansion in the expansion turbine all
the way to the condensing temperature. In this case, as with steam,
essentially all of the heat loss from the cycle would be lost as
latent heat of condensation of the working fluid in the condenser.
In this respect, ammonia offers another advantage. The latent heat
of ammonia is relatively low, about half that of steam on an equal
molar basis.
[0088] A further advantage of ammonia is that there is no vacuum
pressure used anywhere within the working fluid during operation of
the present invention. With steam and other working fluids that
condense at a vacuum pressure at ambient temperatures, vacuum
conditions at the condenser cause air to be leaked into the working
fluid from the surrounding atmosphere through equipment seals. A
deaerator system is needed with these other working fluids to
remove the air during operation. With the present invention,
however, no deaerator is needed because ammonia condenses well
above atmospheric pressure. Furthermore, because of the very low
normal boiling point of ammonia (-28.degree. F.), the process
equipment remains pressurized when idle, even in subfreezing
winter-time conditions. Therefore, no vacuum pump is needed to
remove air from the condenser system prior to startup. This
facilitates rapid startup of the system.
[0089] The relatively high condensing pressure of ammonia has a
further advantage over working fluids that condense at a vacuum,
such as steam. With these latter fluids, the surface of the
condensed fluid must be elevated, often 30 feet or more, in order
to deliver sufficient net positive suction head to the circulating
pump (i.e., pump 111). This elevation adds greatly to the cost of
the equipment because of the need to include structural steel,
platforms, ladders, and other structural items to elevate the
condenser and associated piping and instrumentation. With ammonia,
however, subcooling of the liquid slightly below its condensing
temperature can achieve the same effect as elevation. For example,
subcooling the saturated liquid to a temperature of 83.degree. F.
from 85.degree. F., lowers the vapor pressure of the liquid by
about 5 psia, which is the equivalent of adding about 20 feet of
elevation. (In contrast, subcooling of saturated water to
83.degree. F. from 85.degree. F. adds about the equivalent of only
0.1 feet of elevation.) Thus the present invention allows the
liquid surface elevation to be low, i.e., under ten feet relative
to the pump elevation (which is normally at the lowest point in the
process structure). In turn, this low elevation profile keeps the
cost of structural components relatively low and permits the
present invention to retain the desired characteristic of
compactness. The terms subcool or subcooling, as used herein, mean
further cooling the condensed working fluid to a temperature which
allows the elevation of the surface of the working fluid above pump
111 to be lowered substantially without incurring a loss of net
positive suction head, a term well known in the art of pumping
fluids. The degree to which subcooling is used as a substitute for
elevation of the surface of the working fluid is a choice of the
system designer based on practical and economic considerations.
Subcooling thus permits pump 111, when the same is in operation, to
forward the working fluid to the heat transfer means without any
substantial degree of cavitation, i.e., cavitation which might harm
the process operation or equipment. Cavitation can result in a loss
of pump performance, and by the inherent nature of cavitation can
cause rapid changes in flow which could result in shutdown of the
process of the present invention or even physical damage to the
pump 111. As is well known in the art, cavitation is vaporization
of a liquid within a pump, and while is generally impossible to
avoid all possible occurrences of cavitation, one object of
subcooling is to provide sufficient net positive suction head in
order to reduce cavitation to a point where it has no noticeable
impact on the process or equipment used in the practice of the
present invention.
[0090] Essentially, when idle, the system of the present invention
is ready to start as soon as the heat source stream becomes
available. Warming of the working fluid to supercritical
temperature may occur in as little time as it takes to make one
pass through the heat exchanger. Thus the present invention can be
rapidly started in those cases in which immediate power is
needed.
[0091] With ammonia as the working fluid, the minimum temperature
of heat source streams to which the present invention is applicable
is about 310.degree. F., 40.degree. F. above the critical
temperature of ammonia (the minimum needed to meet the basic
requirements of a working fluid for the present invention, as
described above). In practice, however, the lowest temperature
applicable to the present invention is determined as a matter of
economics. The larger the heat source stream, the lower the
acceptable temperature can be because of the economics of scale. In
general, for most applications of the present invention, the heat
source temperature should be at least 400.degree. F.
[0092] The maximum temperature of heat source streams is unlimited
from the standpoint of theoretical process operation. However, as
the temperature of the heat source increases, the present invention
is subject to a mechanical limitation in that very high pressure
operation is needed to make full use of the available heat. In
general, as either temperature or total heat content or a
combination of the two increase, other types of thermodynamic power
cycles would begin to exhibit superior thermal efficiency and
better overall economics. Roughly, 700.degree. F. is a practical
upper limit for most heat source streams. However, if the heat
source stream is very small or some other special condition of
operation applies, such as the need for rapid startup, the present
invention may still prove to be more economic with higher
temperature heat source streams than other types of power
cycles.
[0093] With some applications of the present invention, other
working fluids will be preferred over ammonia, particularly in
those applications where the heat source stream is at a temperature
outside the range of preferred application for ammonia (400.degree.
F.-700.degree. F.).
[0094] Among the possible fluids, two have been identified as good
alternatives to ammonia: chlorodifluoromethane and sulfur dioxide.
Table I lists the key properties of these two compounds and
compares them with the properties of ammonia.
1TABLE I Condensing Normal Chemical Critical Critical pressure
boiling Fluid formula pressure temperature @ 85.degree. F. point
Ammonia NH.sub.3 1636 psia 270.degree. F. 167 psia -28.degree. F.
Chloro- CHClF.sub.2 721 psia 205.degree. F. 172 psia -41.degree. F.
difluoro- methane Sulfur SO.sub.2 1143 psia 316.degree. F. 65 psia
14.degree. F. dioxide
[0095] Because of its relatively low critical temperature,
chlorodifluromethane should be considered as the working fluid in
applications in which the heat source temperature is less than
400.degree. F. and above 290.degree. F. Moreover, the low critical
pressure of chlorodifluoromethane provides a practical design
benefit in that it makes it possible to select a design pressure
that is sufficiently low to prevent condensation of liquid in the
expansion turbine during expansion. Condensation can be a practical
problem for the mechanical design of the expansion turbine because
of the tendency for liquid droplets to erode the turbine blades.
Above 400.degree. F. for the heat source stream, ammonia would
normally be preferred simply for the reason that it produces more
net power, mainly because the parasitic power required to pump
chlorodifluoromethane is significantly higher than that required
for ammonia.
[0096] The above cited temperature of 290.degree. F. is roughly the
coldest temperature that is still within the preferred range of
application for the present invention. However, colder temperatures
are theoretically possible with proper selection of working fluid.
For instance, bromotrifluoromethane, also known as Freon 13B1, with
a critical temperature of 152.degree. F., permits operation of the
present invention with heat source streams as cold as about
210.degree. F. However, at this temperature, the thermal efficiency
of the present invention becomes very low--in the range of 4 to 5
percent. Economic operation would be possible only in unusual cases
in which the heat source stream is very large or in which the
wholesale price of electricity is unusually high.
[0097] Sulfur dioxide should be considered as the working fluid in
applications in which the heat source stream is roughly 600.degree.
F. or higher. Sulfur dioxide is excellent as a working fluid in
that it will, in theory, produce more net power than ammonia under
similar conditions of design and heat source application. Below
600.degree. F., this superior performance can only be obtained by
allowing condensation to occur in the expansion turbine, which is
problematic for the mechanical design of the turbine. Furthermore,
sulfur dioxide, as shown in Table I, condenses at a much lower
pressure than ammonia. This makes the process less compact and the
expansion turbine more expensive, in that more stages of expansion
are required in the expansion turbine. As those skilled in the art
of design optimization will appreciate, the determination of
whether to select sulfur dioxide or ammonia as the working fluid
can only be determined by comparing the results of detailed designs
to determine if the power advantage of sulfur dioxide outweighs its
disadvantages with respect to the cost of the equipment. As a
general rule, however, as both heat source temperature and plant
scale increase, sulfur dioxide will begin to exhibit superior
economics.
[0098] Although the choice of working fluid is important in
achieving the objects of the invention, the mechanical design and
arrangement of the process equipment is also important to achieving
these objects as described below.
[0099] Heat exchanger 101 is a heat exchange device of proper
engineering design and material selection for the two fluids being
exchanged. The exchange of heat takes place indirectly across a
solid barrier between the two streams, generally such barrier being
a metal tube wall. Flow of the two streams is in opposite, or
countercurrent, direction through the heat exchanger such that the
working fluid is heated to a temperature near to that of the heat
source stream, and simultaneously, the heat source stream is cooled
to a temperature near to that of the working fluid as it enters the
heat exchanger. In most cases of heat exchanger design, the working
fluid would be directed to flow inside multiple, small-diameter
tubes because this is the least expensive way to confine a high
pressure fluid. The heat source, which will usually be lower in
pressure than the working fluid, would be directed to flow outside
the tubes within an outer shell of the heat exchanger. The design
and layout of the heat exchanger's tubes will require no special
consideration for boiling since no boiling of the working fluid
occurs. However, design considerations may dictate that the tube
layout and tube diameter change at some intermediate point within
the heat exchanger because of the large change in volumetric flow
which occurs as the working fluid is heated and transitions from
liquid to vapor.
[0100] Multiple units of heat exchanger 101 may also be used if
more than one heat source is available in the same general
facility. In this case, the working fluid is split into streams of
parallel flow and each stream routed to a separate heat source with
separate heat exchanger, from each of which the separate streams
are recombined into a single stream for conversion into power. In
most cases, the flow of each separate stream of the working fluid
must be controlled independently in order to obtain the desired
temperature of each stream as it leaves each respective heat
exchanger. The heat sources can be dissimilar in temperature,
pressure, or physical state because each heat source is exchanged
separately. For example, the present invention could be applied
simultaneously to streams as different as a flue gas and a liquid
chemical stream under high pressure.
[0101] In most cases, to minimize disruption to the operation of an
existing facility, good engineering practice would dictate that the
heat exchanger(s) be located at the existing locations of the heat
source stream(s), i.e., the heat source streams are not rerouted.
The remaining process units of the present invention may be placed
anywhere else within the facility. Indeed, this is an important
feature of the present invention, in that most of the power
production equipment may be located at a significant distance from
the existing facility if safety reasons or other reasons warrant.
The high pressure of the working fluid is the property which
permits long distances to be practical, because any pressure losses
in the piping runs are a minor proportion of the total pressure
and, as such, have a relatively small negative impact on power
production. The present invention also permits, by virtue of its
simplicity of operation and design, for the design of the piping
runs to and from the heat exchanger(s), and the heat exchanger(s)
themselves, to be welded construction throughout all of the
sections in contact with ammonia without need for valves, fittings,
instrumentation, or other screwed or mechanically sealed connectors
to be present which may otherwise leak the working fluid.
[0102] The equipment used for expansion turbine 106 is not limited
to any particular type; it may be any mechanical device of proper
design and size which converts the work of expansion of the working
fluid to mechanical energy. This includes, but is not limited to,
centrifugal expanders, which are generally used in small
applications, and axial-flow machines, which are most often used
for large applications. For applications of the present invention
in which very high pressures of the working fluid are used, a
non-continuous expansion device such as a device that contains
expansion chambers with piston drivers may prove to be the most
practical and economic.
[0103] The pressure of the working fluid in heat exchanger 101 must
exceed its critical pressure to suppress boiling and to operate in
a manner consistent with the purpose of the present invention.
Otherwise, the operating pressure has no maximum limit from the
standpoint of theoretical process operation. A fluid above its
critical temperature remains as a vapor and thus is theoretically
suitable for expansion in a turbine regardless of pressure. In
practice, however, the choice of operating pressure is subject to
the practical design limitations of the process equipment and
overall economic considerations. In particular, one limitation on
the use of higher pressure is the potential for the working fluid
to condense at the low pressure, cold end of the expansion turbine.
Higher pressure means that the expansion turbine operates with a
greater pressure ratio of expansion, resulting in more cooling of
the working fluid and thus more condensation. Some condensation is
acceptable in a properly designed expansion turbine, but excessive
condensation can erode the turbine blades or produce other forms of
mechanical damage.
[0104] As a general rule, the higher the temperature of the working
fluid as it enters the expansion turbine, the higher the pressure
can be without causing excessive condensation. To illustrate this
point, consider an expansion turbine operating on ammonia with a
0.85 adiabatic efficiency and which discharges to a condenser
operating at a pressure of 167 psia and 85.degree. F. If the
working fluid ammonia enters the turbine at a temperature of
450.degree. F., then, according to the physical properties of
ammonia and its behavioral characteristics when undergoing work of
expansion, the maximum pressure that this stream can be without
causing more than 5 percent of the fluid at the cold end of the
turbine to condense is about 2261 psia. By way of comparison, if
the fluid temperature is 550.degree. F., 100.degree. F. higher, the
maximum pressure for 5 percent condensation is about 3792 psia. If
no condensation were allowed, then these maximum pressures would be
about 1788 psia and 3016 psia, respectively.
[0105] The bypass line and valve, 113, which permit bypass of the
working fluid around the expansion turbine 106, provide an
important operational benefit, particularly whenever rapid startup
is desired and/or when the heat source stream flows intermittently.
(An example of this is an application in which the present
invention is employed as the bottoming cycle for a peak load power
generating combustion turbine, as described in Example IV,
presented later.) For instance, when the heat source stream is not
available but is expected to begin flowing soon, the bypass line
permits full flow circulation of the working fluid in the liquid
state. The valve maintains pressure of the liquid in the heat
exchanger at or above the designated normal operating pressure. In
this mode, the system is ready for the heat source stream to begin
flowing and then normal operation can commence as rapidly as
possible. After the heat source stream flow has begun and as soon
as the working fluid reaches an operationally acceptable
supercritical temperature as it emerges from heat exchanger 101,
valve 105 in the normal flow path is opened slowly to initiate flow
through the expansion turbine 106. Simultaneously, valve 113 is
closed slowly as flow is transferred to the turbine.
[0106] It is possible for the present invention to function without
the bypass line and valve 113. In this case, startup is initiated
with the heat source stream already flowing through heat exchanger
101. When pump 111 is started, liquid working fluid enters heat
exchanger 101 and begins raising the pressure inside the heat
exchanger. Valve 105 is closed, which permits the pressure to
increase behind it. Because the working fluid will boil during this
phase of pressure increase during startup, heat exchanger 101 must
be of a design and construction which can withstand the internal
vibration and forces of boiling during the length of time needed
for startup. Furthermore, expander 106 must be of a design and
construction which can withstand a rapid introduction of hot,
vaporized working fluid passing through valve 105 when valve 105
begins opening to maintain the desired pressure in heat exchanger
101. Because of these special process equipment requirements, it is
expected that applications of the present invention in this form
(i.e., no bypass line and valve 113) will be limited. However, this
form of the present invention does offer an advantage: startup and
transition to normal operation c an be effected by simply starting
pump 111.
[0107] Valve 105 provides a means of throttling the working fluid
to maintain the desired supercritical pressure in heat exchanger
101 during periods of operation in which the flow rate of the
working fluid is significantly less than that prescribed by design.
Generally, such periods of operation would only be necessary
whenever the flow or temperature of the heat source stream itself
is significantly less than that prescribed by design. Whenever the
flow of the working fluid is near or higher than the design flow,
expansion turbine 106 provides, by virtue of its resistance to flow
through its various internal components, the means of throttling.
In other words, the expansion turbine provides a dual role, both as
means of expansion to produce power and as means of throttling. In
this case, valve 105 is opened to its fullest extent possible and
provides little or no throttling of the working fluid. A variation
of the present invention can be envisioned in which valve 105 is
eliminated from the design. This variation would be possible for
particular applications in which the flow of the working fluid can
always be maintained near to or above the design flow because there
is always sufficient heat available from the heat source stream.
Further variations can be envisioned in which valve 105 is
eliminated and replaced with an adjustable form of flow throttling
that is part of the expansion turbine itself (such as guide vanes
commonly used with commercial turbines).
[0108] An advantage of the present invention is that all of the
process equipment (except for the heat exchanger) may be mounted on
one or more portable transportation means. One form of portable
transportation means commonly known to those skilled in the art is
a portable skid. A portable skid is essentially a free-standing
platform on which the process equipment of the present invention is
mounted. For practical purposes of transportation, the physical
dimensions of a portable skid are limited to that which can be
transported by railroad car or by truck over most highways of the
United States as defined by the U.S. Department of Transportation
regulations. The use of portable skids has an economic advantage in
that most of the process hardware can be assembled in a shop rather
than at its operational site in the field. Two features of the
present invention make skid-mounting a practical alternative to
field construction. First is the invention's simplicity, in that
very few large items of process equipment are needed. Second is the
invention's compactness. The relatively high working pressure and
the high condensing pressure of the working fluid keeps the
volumetric flow of the working fluid relatively low through the
expansion turbine and condenser, and therefore the size of the
equipment and interconnecting piping is relatively small.
Subcooling of the condensed working fluid, discussed earlier, also
contributes to the compactness of the present invention. Subcooling
permits the elevation of the liquid surface above pump 105 to be
relatively small (under ten feet, if desired, by the designer) such
that the height of the process equipment above the surface of the
portable skid is relatively low (compared with steam cycles). A
related object of this invention is to permit, for a process unit
of about 4 MW or less in net power output, all of the processing
equipment (except for heat exchange with the heat source stream) to
be located on a single portable skid.
[0109] Another advantage of the present invention is that the
process equipment (also called means or elements in the claims)
that is mounted on the portable skid may be standardized, i.e.,
designed and constructed according to a fixed set of specifications
without regard to a specific application to a heat source stream.
Only the heat exchanger, which is separate from the portable skid,
need be designed for any particular application. The design of the
heat exchanger must be such that working fluid can be processed by
the heat exchanger and returned to the portable skid at a flow
rate, pressure, and temperature within an acceptable range of
normal operation for the standardized equipment. Of course, the
heat source stream must be sufficiently large in heat content and
high in temperature to meet these requirements. This, in effect, is
a limitation of standardization, because it is unlikely that a heat
source stream will exactly match the requirements of a standardized
unit. Therefore, some heat will not be converted to power which may
otherwise be used if all of the equipment were custom designed.
However, the use of standardized equipment eliminates many costs
associated with custom design and thus standardization is expected
to produce an economic advantage with many applications.
[0110] Another advantage of the present invention is that operation
of the process may be automated, i.e., made capable of operation
under normal or routine circumstances without benefit of human
intervention. An automated version of the present invention is one
in which human intervention is limited to preparing the equipment
for normal operation and then authorizing the unit to operate by
activating an electronic switch. All remaining startup and normal
operation activities are then carried out by instruction to the
operating equipment via electronic signal from a process control
computer. No claim is made with respect to the process control
computer itself. Rather it is claimed that the simplicity of the
present invention, having only a single circulating loop of the
working fluid, permits an automated version of the process to
function with state-of-the-art computer controls already in
existence. Example I, which contains an explanation of a typical
startup sequence, illustrates the simplicity of the startup
procedure.
[0111] A specific application of this invention is to produce power
from a flowing stream of hot pressurized gas which has been
generated in a coal gasifier. A coal gasifier is a vessel in which
coal or some other carbon containing solid material is converted to
a gas by reacting with oxygen. Of particular importance is the type
of coal gasifier in which the hot gas is quenched in water. The
resulting water vapor-laden gas is the heat source stream to which
this invention i s applied (as further detailed in Example II
below). A related object is to utilize the power generated from
this application as supplemental drive power for the main air
compressor in an air separation unit, which supplies oxygen to the
coal gasifier. An air separation unit is a process in which oxygen
is extracted from air as a stream of high purity (usually over 90
percent by volume) oxygen. The main air compressor delivers air to
the air separation unit at a pressure sufficient for the extraction
process to take place.
[0112] The present invention may be applied as a bottoming cycle to
convert waste heat from a topping cycle into power. A topping cycle
is defined as any power cycle which utilizes a fuel as an energy
source to produce power. A bottoming cycle is defined as any cycle
which converts waste heat from the topping cycle into additional
power. The two cycles taken together are often referred to as a
combined cycle. Examples of topping cycles as defined here include
combustion turbines, internal combustion engines, and fuel cells.
Of particular interest is that the present invention may be applied
to the hot exhaust gas from a combustion turbine (i.e., it is
applied as a bottoming cycle with a combustion turbine as the
topping cycle). A combustion turbine is a power generating system
of equipment in which air from the atmosphere is compressed, a fuel
is burned directly in the compressed air stream to produce a hot
stream of combustion gas, and then the hot combustion gas is
expanded to produce power. Of particular importance are those
applications where the present invention is likely to be more
economic or practical than the traditional bottoming cycle methods
(such as the steam-based Rankine cycle). Examples of important
applications for the present invention are: (1) small applications,
in which the bottoming cycle produces less than about 10 MW of
power; (2) recuperated combustion turbines, in which the hot
exhaust after expansion to produce power has been partially cooled
by heat exchange with compressed air, a method of heat recovery
known as recuperation; and (3) peaking units, combustion turbines
that operate intermittently to produce electric power during
unusual periods of high electricity demand called peak loads.
Typically, peak loads have a duration of less than a day. (The
McGraw-Hill Encyclopedia of Science & Technolog, 7.sup.th
edition, discusses the art of servicing peak loads with peaking
units and defines these terms.) If a bottoming cycle is employed to
meet the demands of peak loads, it must start quickly and
automatically to be viable as an economic source of power. As
illustrated in Example IV below, the present invention has these
characteristics.
[0113] The present invention is also characterized by the process
steps or methods (and attendant equipment items) that it does not
require in order to function or to meet the objects of the
invention. Such steps, which are employed in many Rankine cycle
power generating plants, particularly steam plants, are obviated by
the inherent simplicity of the present invention and the beneficial
effect of the physical properties of the working fluids. The
process steps or methods that are not used by the present invention
include:
[0114] Boiler equipment, which is defined as equipment needed to
facilitate the boiling operation other than the heat exchanger,
such as a vessel to separate vapor from liquid, forced or natural
circulation piping, forced circulation pumps, mist eliminators, and
deionization equipment.
[0115] Vaporization of the working fluid at more than one discrete
pressure level ("discrete pressure level" being defined as any
pressure within a range of plus or minus 5 percent of an exact
numerical pressure).
[0116] Deaeration (i.e., the removal of non-condensable gases that
have entered the working fluid from the atmosphere during normal
operation).
[0117] Reheat (i.e., warming of the working fluid between expansion
turbine stages by adding heat to the working fluid from an outside
heat source).
[0118] Recuperation (i.e., transferring heat from the warm working
fluid exhaust leaving the expansion turbine, stream 108, to the
relatively cool condensed working fluid, stream 103).
[0119] Elevation of the liquid surface of the condensed working
fluid to a height more than 10 feet above the intake of pump
111.
[0120] Vacuum pressure operation and the attendant vacuum producing
equipment.
EXAMPLES
[0121] In order that those skilled in the art may better understand
how the present invention can be practiced, the following examples
are given by way of illustration only and not necessarily by way of
limitation, since numerous variations thereof will occur and will
undoubtedly be made by those skilled in the art without
substantially departing from the true and intended scope and spirit
of the instant invention herein taught and disclosed.
[0122] Five examples are presented below, each illustrating a
different application of the present invention.
[0123] The following design and operating parameters, except as
noted, were used in all three examples, and are typical values for
commercial applications:
[0124] Adiabatic efficiency of the expansion turbine:
Turb.sub.eff=85 percent
[0125] Total mechanical and generator losses: M&G.sub.loss=2
percent
[0126] Condensing temperature: T.sub.cond=85.degree. F.
(105.degree. F. in Example IV)
[0127] Adiabatic efficiency of the circulating pump:
Pump.sub.eff=80 percent
[0128] Electric motor efficiency: Motor.sub.eff=90 percent
[0129] For each example, the gross electric power output of the
expansion turbine and the gross electric power consumption by the
circulating pump were calculated by the following formulas,
respectively.
Gross electric power output=(Theoretical expansion turbine power)
(Turb.sub.eff/100) (100-M&G.sub.loss)/100)
Gross pumping power used=(Theoretical pumping
power)/(Pumper.sub.eff/100)/- (Motor.sub.eff/100)
[0130] where,
[0131] The theoretical expansion turbine power and the theoretical
pumping power were both derived by determining the power derived
from the reversible isentropic change in pressure of the working
fluid. Properties of the working fluid were estimated by
ChemCAD.RTM. process simulation software using equations of state.
ChemCAD.RTM. also calculated all temperatures, pressures, flows,
and other properties of streams cited in the examples. NOTE: Any
references made herein to materials and/or apparatus which are
identified by means of trademarks, trade names, etc., are included
solely for the convenience of the reader and are not intended as,
or to be construed, an endorsement of the materials and/or
apparatus.
[0132] Total net power was calculated by the formula:
Net power=(Gross expansion turbine power)-(Gross pumping power)
[0133] Thermal efficiency figures cited in the examples were
calculated by expressing the net power as a percentage of the
available heat in the heat source stream (as a flow of heat),
converted to units of electric power. Available heat is defined as
all of the sensible and latent heat which could be extracted from
the heat source stream if it were cooled to lowest temperature in
the system, i.e., the same temperature at which the condenser
operates, 85.degree. F. (105.degree. F. in Example IV).
Example I
[0134] Example I illustrates the general case of extracting heat
from a typical industrial heat source stream. A pressurized stream
of hot water in the liquid state at 1000 psia, 500.degree. F. and
flowing at a rate of 200 gallons per minute represents the heat
source. The working fluid of choice is ammonia.
[0135] Only three process design parameters need be selected for
the cycle in Example I: (1) the temperature of the working fluid as
it leaves heat exchange (stream 104 in FIG. 1), (2) the maximum
pressure of the working fluid, and (3) the circulation rate of the
working fluid. A temperature of 450.degree. F. is selected for the
working fluid, which, given the 500.degree. F. temperature of the
heat source, provides an adequate temperature difference of
50.degree. F. for heat transfer at the hot end of the heat
exchanger. A working fluid pressure of 1800 psia is selected,
which, upon expansion in the expansion turbine 106 at an initial
450.degree. F., yields a modest and acceptable level of 0.1 percent
liquid in the expansion turbine exhaust. Finally, the circulation
rate of the working fluid is selected to provide, with benefit of
properly sized heat exchange equipment, a 50.degree. F. temperature
difference at the cold end of the heat exchanger. This figure is
calculated to be about 59,500 pounds per hour.
[0136] During normal and steady state operation of the present
invention, the operating conditions at points throughout the cycle
are as follows (see FIG. 1 for references to line and equipment
designation numbers). Stream 104 emerges from heat exchanger 101 at
a temperature of 450.degree. F. having been heated by the
500.degree. F. heat source stream. Valve 113 is fully closed and
valve 105 is fully open. At the calculated design flow rate of
59,500 pounds per hour for stream 104, the expansion turbine 106 by
virtue of its design provides resistance to flow (i.e., throttling)
such that the pressure in stream 104 is maintained at about 1800
psia. Upon expansion of stream 104 through expansion turbine 106 to
a pressure of 167 psia, the working fluid, having been cooled by
work of expansion, emerges saturated at 85.degree. F. and contains
about 0.1 percent liquid of condensation (stream 108). These
conditions are calculated on the basis of obtaining 85 percent of
the theoretical power from the isentropic expansion of ammonia.
Gross power generated at the expansion turbine 106 shaft is 2397
kW. After deducting 2 percent for losses in the generator, gross
electric power output from generator 107 is 2349 kW. Stream 108 is
condensed in condenser 109 at a temperature of 85.degree. F. by an
external cooling source (such as cooling water at 60.degree. F.).
The pressure at which ammonia condenses at 85.degree. F. is 167
psia, which sets the pressure at the discharge of expansion turbine
106. The working fluid liquid is also subcooled in condenser 109 to
a temperature of 83.degree. F. No accumulation or loss of fluid
occurs in surge vessel 110 at steady state. Pump 111 increases the
pressure of the working fluid to 1805 psia, which is 5 psi above
the pressure at the entrance to expansion turbine 106. This 5 psi
difference is to allow for a pressure drop due to flow of the
working fluid through heat exchanger 101 and the attendant piping.
Pump 111 is a type that provides a fixed volume of liquid per
revolution of its drive shaft (i.e., a positive displacement pump).
The prescribed circulation rate of 59,500 pounds per hour is set by
controlling the rotational speed of the drive shaft of pump 111.
Electric power consumed by the drive motor of pump 111 is 198 kW.
The mechanical act of pumping the working fluid warms it slightly
to about 90.degree. F. from 83.degree. F. Thus stream 103 has
conditions of 90.degree. F. and 1805 psia as it enters heat
exchanger 101, completing the cycle as described.
[0137] In the above described cycle, the flow of energy into and
out of the working fluid occurs only at four steps in the cycle:
(1) heat exchanger, (2) expansion turbine, (3) condenser, and (4)
pump. In the heat exchanger, energy flows into the working fluid
from the heat source stream at the rate of 38.36 million Btu per
hour. In the expansion turbine, energy flows out of the working
fluid and is converted to mechanical energy at the rate of 8.18
million Btu per hour. The condenser, which also includes subcooling
of the fluid after it has condensed, removes the bulk of the heat
by rejecting it to an external cooling medium at the rate of 30.79
million Btu per hour. Finally, the pump, using an electric motor as
the energy source, adds 0.61 million Btu per hour to the working
fluid in the form of heat and pressure rise. As with all continuous
cycles, there is no net accumulation or loss of energy by the
working fluid as it makes a cycle during steady state operation.
The total energy added to the fluid in steps 1 and 4, which is
38.97 million Btu per hour, is exactly matched by that removed in
steps 2 and 3.
[0138] The present invention is also capable of operating under
conditions which are significantly different from the above
described normal conditions, a state of operation known as
"off-design." Off-design operation is required whenever external
conditions deviate from normal. There are only two points in the
process where external changes can occur: (1) the heat source
stream as it enters heat exchanger 101 and (2) the cooling stream
for condenser 109. The heat source stream, 102, may vary in the
quantity of heat it carries. This change in quantity of heat is
manifested by the heat source stream as either a change in flow
rate or a change in temperature or some combination of flow and
temperature. If the quantity of heat falls, the circulation rate of
the working fluid must be reduced by an amount necessary to
maintain the maximum power output which can be achieved under the
new conditions. This reduction in circulation rate will cause the
pressure entering the expansion turbine 106 to be reduced below the
normal operating pressure of 1800 psia. In this case, valve 105 is
closed partially to assure that supercritical conditions are
maintained in heat exchanger 101 and that boiling inside the heat
exchanger is avoided. In similar manner, if the quantity of heat
rises, the circulation rate of the working fluid is increased by an
amount necessary to maintain the maximum power output which can be
achieved under the new conditions. However, this increase in
circulation rate is subject to limitations of the system design to
withstand the increase in pressure at the expansion turbine 106
entrance or the capability of the generator to generate the
increased electric power. A change in cooling capacity in condenser
109, which is usually manifested as a change in the temperature of
the cooling stream, will result in either higher or lower
condensing pressure. Normally, no change in operation is required;
the decrease or the increase in power output, respectively, is
simply accepted. However, if the condensing pressure falls to a
point at which the power output limitation is exceeded, or to a
point at which there is more liquid condensate in the expansion
turbine 106 outlet than acceptable by mechanical design of the
turbine, the condensing pressure is raised as needed by limiting
the cooling stream flow.
[0139] To demonstrate off-design operation in this example,
consider the case in which the flow rate of the heat source stream
is reduced by 20 percent to 160 gallons per minute from 200 gallons
per minute. Condensing and subcooling temperatures in condenser 109
remain the same as above. The pressure in heat exchanger 101 is
maintained at 1800 psia. Our process simulation estimates that the
maximum power output under this new condition is achieved by
reducing the working fluid circulation rate to 48,500 pounds per
hour from 59,500 pound per hour. The pressure entering the
expansion turbine 106 falls to about 1455 psia as a result of the
reduced flow rate. Valve 105 is partially closed to maintain the
desired 1800 psia pressure in heat exchanger 101. The temperature
of the working fluid leaving the heat exchanger is estimated to be
456.degree. F., and the temperature entering the expansion turbine
106 is 435.degree. F., a 21.degree. F. difference which is a result
of constant enthalpy expansion across valve 105. (Although the
working fluid is cooled as it passes through valve 105, no energy
enters or leaves the working fluid at this step. This cooling is
simply a phenomenon of the expansion of a non-ideal gas through a
throttling valve in which no work is done.) Gross power generated
at the expansion turbine 106 shaft is 1837 kW, about 23 percent
lower than for normal operation described above.
[0140] An important objective of the present invention is to have
the capability of being started quickly and simply. A typical
startup of the present invention proceeds as follows: The heat
source stream 102 has not begun flowing through heat exchanger 101,
but it is available to be routed to the heat exchanger at the
direction of the process operator. Pump 111 is energized to begin
circulation of the working fluid. Flow passes through heat
exchanger 101 and then through valve 113, bypassing expansion
turbine 106 for now. Valve 105 is fully closed. Valve 113 is
partially open to produce a restriction on the flow of the working
fluid such that the normal operating pressure of 1800 psia or
higher is maintained in the fluid as it exits heat exchanger 101.
Flow continues through the normal flowpath to condenser 109, surge
vessel 110, and back to pump 111 to complete the circuit. All of
the working fluid is in the liquid state at this point in the
startup sequence. Coolant flow (not shown) is started in
preparation for heat input from the heat source stream.
[0141] On direction from the process operator, flow of the heat
source stream 102 is initiated. The working fluid, stream 104, is
expected to heat rapidly to normal operating temperature. The time
required for this rapid warming of the fluid is only the time
needed for one pass of the fluid through the heat exchanger plus
some additional time to heat the structural materials (tubing,
etc.) within the heat exchanger. As stream 104 rises in
temperature, its density decreases, eventually making a transition
from liquid to vapor. The opening in valve 113 is increased
gradually to accommodate this change in density so as to maintain
the desired operating pressure. As stream 104 nears the normal
operating temperature of 450.degree. F., valve 105 is opened
partially to initiate flow to expansion turbine 106.
Simultaneously, valve 113 is closed partially to maintain the
desired operating pressure for stream 104. The process of opening
valve 105 and closing valve 113 continues until all of the flow has
been transferred to expansion turbine 106 and valve 113 is fully
closed. This completes the startup sequence. As this startup
description shows, startup is simple and rapid, an important
feature of the present invention. After starting circulation of the
fluid and initiating flow of the heat source stream, the entire
startup sequence is controlled by setting the position of just two
valves (valves 105 and 113). With benefit of modern process
controls, this startup sequence can be automated and, if desired,
initiated from a remote location.
[0142] FIG. 2 shows the internal temperature profile along the
length of heat exchanger 101 in this example. Temperature
differences between the two streams remain acceptable throughout
the exchanger. The smallest temperature difference between the two
streams is about 30.degree. F., and the largest is about
110.degree. F. The log mean temperature difference (LMTD) weighted
over the length of the exchanger is about 51.degree. F., about the
same as the LMTD of 50.degree. F. based on the endpoints. Overall,
FIG. 2 demonstrates the general acceptability of countercurrent
exchange used by the present invention. FIG. 2 also demonstrates an
important and desirable feature of the process, in that nearly all
of the available heat is extracted from the heat source stream by
virtue of it being cooled to a temperature near ambient. About 87
percent of the heat in the heat source stream is removed.
[0143] The calculated performance numbers for Example I are
summarized as follows:
Gross electric power output=2349 kW
Gross pumping power used=198 kW
Net power=2151 kW
Thermal efficiency=16.7 percent
[0144] An important and much desired feature of the present
invention is that it should be compact in size. One measure of this
feature is the volumetric flow rate of the working fluid. The
volumetric flow rate is roughly indicative of the physical
dimensions of the process equipment. In Example I the flow rate of
the working fluid upstream of expansion turbine 106 is calculated
to be 15,700 cubic feet per hour (ft.sup.3/h). Downstream, after
expansion, this volume increases to 110,500 ft.sup.3/h.
[0145] To put these figures and the performance numbers above into
perspective, they are compared with the numbers from a similar
model of the single-pressure Rankine steam cycle in which
water/steam is the working fluid. The heat source stream was kept
the same. Also, as with the present invention, the processing steps
of the single-pressure Rankine steam cycle are limited to the same
simple and basic configuration of four process steps in which the
working fluid changes state: heat exchange, expansion,
condensation, and pumping. For the heat exchange step of the steam
cycle, three separate steps are necessary: (1) preheating of liquid
water to the boiling point, (2) boiling to produce steam, and (3)
superheating of the steam. In this respect, the heat exchange for
the water/steam cycle case is more complex than that for the
present invention, but for purposes of this exercise so that a
comparison can be made, these three steps are considered to be one
step of heat exchange.
[0146] To the extent possible, other design factors were kept the
same, including adiabatic efficiency of the expansion turbine,
total mechanical and generator losses, condensing temperature,
adiabatic efficiency of the circulating pump, and electric motor
efficiency. The working fluid was superheated to 450.degree. F. in
both cases. The average log mean temperature difference for heat
exchange was kept the same at a value of 51.degree. F.
[0147] Steam pressure entering the expansion turbine was set at 38
psia, which is the optimum pressure for the cycle. Higher and lower
pressures produced less power. Expansion turbine exhaust pressure
was set at 0.6 psia, the condensing pressure for steam at
85.degree. F.
[0148] The results showed that the steam cycle produced
significantly less power than the present invention: 1712 kW versus
2151 kW, net, respectively. Corresponding efficiency was also
lower: 13.3 percent versus 16.7 percent, net, respectively.
[0149] Those skilled in the art will readily recognize that the
steam cycle could be altered to produce more power by adding more
stages of boiling. However, in so doing the steam cycle would lose
the desired characteristic of simplicity exhibited by the present
invention, a characteristic needed to make possible the objects of
the invention described previously.
[0150] With respect to the volumetric flow of the working fluid,
the difference between the present invention and the steam cycle is
very large and very significant. The volumetric flow rates of the
working fluid in the steam cycle, upstream and downstream of the
expansion turbine, respectively, are 347,000 ft.sup.3/h and
12,239,000 ft.sup.3/h. These volumes are more than two orders of
magnitude larger than those described earlier for the present
invention in this example. The relative sizes of the piping, the
condenser, the expansion turbine, and all parts of the steam cycle
involved in handling the working fluid in the vapor phase will have
physical sizes similarly larger by about two orders of magnitude.
For example, the vapor-carrying pipelines for the steam cycle will
have cross-sectional flow areas of more than 100 times those of the
present invention, which means that the respective diameters of the
pipelines will be more than 10 times larger. The cross-sectional
flow areas at points along the flow path inside the expansion
turbine in the steam cycle will be similarly about 100 times larger
than those points in the expansion turbine of the present
invention.
[0151] In summary, Example I illustrates the simple, compact, and
relatively efficient characteristics of the present invention.
These three characteristics, with Example I providing the basis,
may be expressed as follows: The present invention is (1) simple,
in that there are only four basic process steps in which energy is
added to or removed from the working fluid in the form of heat or
work, (2) compact, in that the maximum volumetric flow of the
working fluid is less than one percent of the maximum flow in a
comparable steam cycle, and (3) relatively efficient, in that the
thermal efficiency is greater than 16 percent, a reasonably high
efficiency for a 500.degree. F. heat source stream and greater than
the efficiency of a comparable steam cycle.
EXAMPLE II
[0152] Example II illustrates the application of the present
invention to extracting heat from a pressurized hot gas stream
which has been quenched in water. More specifically, the intended
application is a stream of quenched gas leaving an oxygen-fired
coal gasifier. Typically such a gasifier operates at a temperature
over 2000.degree. F. in the firing zone where oxygen and coal are
reacted. The gas is then quenched in a bath of water to bring the
temperature down to manageable levels, typically 400.degree. F. to
500.degree. F., and to clean coal slag particles out of the gas.
Most of the heat released during oxidation is thus retained by the
quenched gas in the form of latent heat of water vapor mixed with
the gas. This quenched gas must be cooled to near ambient to
facilitate removal of sulfur-bearing compounds later in the
process. Thus the recovery and re-use of this large quantity of
latent heat during gas cooling is important for obtaining good
thermal efficiency.
[0153] Normally, recovery of this latent heat requires a series of
steps including two or more stages of medium- or low-pressure steam
generation for subsequent power generation plus, in some cases,
resaturation of the clean, sulfur-free gas. The present invention,
however, when applied to the raw gas as a heat source, eliminates
this series of steps and replaces them with a single power
generating expansion turbine. Moreover, because of the compact size
of the expansion turbine and the simplicity of the present
invention, the power from the turbine can be used to provide
supplemental drive power for the main air compressor in the
gasification plant's air separation unit, the largest single user
of electric power in the plant. Thus the inherent inefficiency of
first generating electricity then and subsequently consuming it in
an electric motor is eliminated. (The term supplemental drive power
refers to power that offsets but does not completely eliminate the
power provided by the main air compressor's electric drive motor.
The air separation unit is a process which separates the oxygen and
nitrogen in air into separate streams, each stream having a higher
proportion of oxygen or nitrogen, respectively, than normally found
in air. The main air compressor provides the necessary flow and
pressure for the air to be separated within the air separation
unit.)
[0154] Table II shows typical flowing conditions for a quenched gas
stream from a coal gasifier operating at high pressure, near 1000
psia. Although the gas contains many compounds in small or trace
quantities, for purposes of simplification, only the major
components are included. The gas is saturated, such that any
removal of heat is accompanied by both cooling in temperature and
condensation of water. The quantity of water in the gas was
estimated for the true, non-ideal gas case, and thus the gas
composition estimate below contains more water vapor than would be
predicted using steam tables.
2 TABLE II Pressure 955 psia Temperature 471.degree. F. Flow
852,000 lb/h Composition H.sub.2O 61 mole % CO 20 mole % H.sub.2 14
mole % CO.sub.2 4 mole % N.sub.2 0.5 mole % H.sub.2S 0.5 mole %
[0155] Quenched gas, when cooled, does not exhibit the same linear
change in temperature as the hot water stream used as the heat
source in Example I. Instead, the removal of heat produces very
little change in temperature at first, followed by a rapid
acceleration of temperature loss as the quantity of water vapor in
the gas is depleted. FIG. 3 shows the internal temperature profile
along the length of the heat exchanger in this example. The
smoothly curving line for the heat source illustrates this property
of the quenched gas. Because the heat source's line and the working
fluid's line curve away from each other in the middle of the heat
exchanger, it is possible to specify an exchanger design in which
the endpoints are very close to each other in temperature
difference. For this example, a temperature difference of
10.degree. F. was chosen for both the cold end and the hot end of
the exchanger. Even with this narrow difference at the endpoints,
the average log mean temperature difference over the whole
exchanger is a quite acceptable value of 67.degree. F.
[0156] The advantage of quenched gas as a heat source is twofold.
The working fluid can be heated to a higher temperature, resulting
in higher expansion turbine efficiency and output, and the heat
source can be cooled to a lower temperature, resulting in more heat
being recovered. The overall effect is a greater power output and a
higher thermal efficiency than would be possible with heat source
streams of equal temperature in the forms of liquids or
non-condensing gases. Or, expressed in another way, the quenched
gas from a gasifier is an excellent match as a heat source for the
present invention, and thus represents an important specific
application for which the present invention is intended.
[0157] The working fluid of choice for this example is ammonia. A
working fluid pressure of 1900 psia is selected, which, upon
expansion in expansion turbine 106 at an initial 461.degree. F.,
yields a modest and acceptable level of 0.1 percent liquid in the
turbine exhaust. The circulation rate of the working fluid is
selected to provide, with benefit of properly sized heat exchange
equipment, a 10.degree. F. temperature difference at the cold end
of the heat exchanger. This is calculated to be about 933,400
pounds per hour. The cooling of the heat source stream to this
10.degree. F. difference resulted in over 95 percent of the
available heat being removed.
[0158] During normal and steady state operation of the present
invention, the operating conditions at points throughout the cycle
are as follows (see FIG. 1 for references to line and equipment
designation numbers). Stream 104 emerges from heat exchanger 101 at
a temperature of 461.degree. F. having been heated by the
471.degree. F. heat source stream. Simultaneously, the heat source
stream is cooled to about 100.degree. F., and more than 99 percent
of the water vapor in the stream is condensed. Valve 113 is fully
closed and valve 105 is fully open. At the calculated design flow
rate of 933,400 pounds per hour for stream 104, expansion turbine
106 by virtue of its design provides resistance to flow such that
the pressure in stream 104 is maintained at about 1900 psia. Upon
expansion of stream 104 through expansion turbine 106 to a pressure
of 167 psia, the working fluid, having been cooled by work of
expansion, emerges saturated at 85.degree. F. and contains about
0.1 percent liquid of condensation (stream 108). These conditions
are calculated on the basis of obtaining 85 percent of the
theoretical power from the isentropic expansion of ammonia. Gross
power generated at the shaft of expansion turbine 106 is 38,628 kW.
No power is deducted for generator losses because no generator is
used. Instead the power output is used to assist the electric motor
drive for the main air compressor in the air separation plant.
Stream 108 is condensed in condenser 109 at a temperature of
85.degree. F. by an external cooling source (such as cooling water
at 60.degree. F.). The pressure at which ammonia condenses at
85.degree. F. is 167 psia, which sets the pressure at the discharge
of expansion turbine 106. The working fluid liquid is also
subcooled in condenser 109 to a temperature of 83.degree. F. No
accumulation or loss of fluid occurs in surge vessel 110 at steady
state. Pump 111 is a centrifugal type that provides a flow rate and
discharge pressure as prescribed by the pump's performance curve. A
centrifugal type is selected because it is practical for the
relatively large flow rate (about 3000 gallons per minute) found in
this example. A flow control valve 112 downstream of the pump is
partially closed to limit the discharge flow from pump 111 to the
desired flow rate of 933,400 pounds per hour. Pump 111 increases
the pressure of the working fluid to 1930 psia, which is 30 psi
above the pressure at the entrance to expansion turbine 106. This
30 psi difference is to allow for three separate pressure drops
between pump 111 and expansion turbine 106: (1) 20 psi for the loss
through the flow control valve 112 downstream of pump 111, (2) 5
psi for losses in heat exchanger 101, and (3) 5 psi for the losses
in the piping between the heat exchanger and the expansion turbine.
This additional 5 psi loss in the piping, which was not used in
Example I, is to account for the expected long run of piping
between the gasification area and the air separation unit. Electric
power consumed by the drive motor of pump 111 is 3330 kW. The
mechanical act of pumping the working fluid warms it slightly to
about 90.degree. F. from 83.degree. F. Thus stream 103 has
conditions of 90.degree. F. and 1910 psia as it enters heat
exchanger 101, completing the cycle as described.
[0159] The calculated performance numbers for Example II are as
follows:
Gross expansion turbine power output=38,628 kW (unadjusted power
output; no generator loss is included since no generator is
used)
Gross pumping power used=3330 kW
Net power=35,298 kW
Thermal efficiency=19.6 percent
Example III
[0160] Example III illustrates the case of extracting heat from the
hot exhaust stream of a combustion turbine unit. In other words,
the present invention is applied as the bottoming cycle of a
combined cycle unit. The type of combustion turbine selected for
this example is one in which the compressed air stream feeding the
burner and combustion turbine is recuperated, i.e., the air is
preheated by heat exchange with the hot exhaust coming directly
from the combustion turbine. This exchange of heat leaves the final
exhaust stream significantly cooler than exhaust streams from
non-recuperated cycles, which typically have exhaust temperatures
above 900.degree. F.
[0161] A commercial gas turbine, the Solar.RTM. Mecury.TM. 50, was
selected as the basis for this example. At design rated conditions,
the Mercury 50 has a net output of 4180 kW. Exhaust flow is 129,190
lb/h at a temperature of 694.degree. F. This combination of
moderate exhaust temperature and relatively small size are a good
match for the capabilities of the present invention.
[0162] The working fluid of choice for this example is ammonia. The
design parameters are selected as follows. A temperature of
644.degree. F. is selected for the working fluid, which, given the
694.degree. F. temperature of the heat source, provides an adequate
temperature difference of 50.degree. F. for heat transfer at the
hot end of the exchanger. Because of the high temperature of the
working fluid, it is not considered practical to select a pressure
sufficiently high to expand the working fluid to condensing
conditions. Such a pressure would exceed 4500 psia and may approach
or be beyond the state-of-the-art for mechanical design. A pressure
of 3500 psia is used instead, which is typical for commercial high
pressure steam turbines. Finally, the circulation rate of the
working fluid is selected to provide, with benefit of properly
sized heat exchange equipment, a 50.degree. F. temperature
difference at the cold end of the heat exchanger. This rate is
calculated to be about 24,300 pounds per hour. Overall, the
weighted log mean temperature difference between the two fluids
being exchanged is about 66.degree. F., very adequate for
commercial design. About 84 percent of the heat in the heat source
stream is removed.
[0163] During normal and steady state operation of the present
invention, the operating conditions at points throughout the cycle
are as follows (see FIG. 1 for references to line and equipment
designation numbers). Stream 104 emerges from heat exchanger 101 at
a temperature of 644.degree. F. having been heated by the
694.degree. F. heat source stream. Valve 113 is fully closed and
valve 105 is fully open. At the calculated design flow rate of
24,300 pounds per hour for stream 104, the expansion turbine 106 by
virtue of its design provides resistance to flow such that the
pressure in stream 104 is maintained at about 3500 psia. Upon
expansion of stream 104 through expansion turbine 106 to a pressure
of 167 psia, the working fluid, having been cooled by work of
expansion, emerges at a temperature of 144.degree. F. (stream 108).
These conditions are calculated on the basis of obtaining 85
percent oft he theoretical power from the isentropic expansion of
ammonia. Gross power generated at the shaft of expansion turbine
106 is 1495 kW. After deducting 2 percent for losses in the
generator, gross electric power output from generator 107 is 1465
kW. Stream 108 is cooled from 144.degree. F. to its saturation
temperature of 85.degree. F. and condensed in condenser 109 by an
external cooling source (such as cooling water at 60.degree. F.).
The pressure at which ammonia condenses at 85.degree. F. is 167
psia, which sets the pressure at the discharge of expansion turbine
106. The working fluid liquid is also subcooled in condenser 109 to
a temperature of 83.degree. F. No accumulation or loss of fluid
occurs in surge vessel 110 at steady state. Pump 111 increases the
pressure of the working fluid to 3505 psia, which is 5 psi above
the pressure at the entrance to expansion turbine 106. This 5 psi
difference is to allow for a pressure drop due to flow of the
working fluid through heat exchanger 101 and the attendant piping.
Pump 111 is a type that provides a fixed volume of liquid per
revolution of its drive shaft (i.e., a positive displacement pump).
The prescribed circulation rate of 24,300 pounds per hour is set by
controlling the rotational speed of the drive shaft of pump 111.
Electric power consumed by the drive motor of pump 111 is about 164
kW. The mechanical act of pumping the working fluid warms it to
about 96.degree. F. from 83.degree. F. Thus stream 103 has
conditions of 96.degree. F. and 3505 psia as it enters heat
exchanger 101, completing the cycle as described.
[0164] The calculated performance numbers for Example III are
summarized as follows:
Gross electric power output=1465 kW
Gross pumping power used=164 kW
Net power=1301 kW
Thermal efficiency=20.9 percent
Example IV
[0165] As in Example III, Example IV illustrates the case of
extracting heat from the hot exhaust stream of a combustion turbine
unit. In this example, however, the combustion turbine is employed
to rapidly provide supplemental electric power during intermittent
periods in which electric power demand is unusually high. This
supplemental electric power is commonly referred to in the electric
utility industry as peak load. The combustion turbines used to
generate power during peak load are called peaking units. As those
persons skilled in the art of power generation by an electric
utility will appreciate, the present invention has operational
characteristics which are advantageous to the generation of peak
load power, including (1) rapid and simple startup and (2)
non-interference with the operation or power output of the peaking
unit.
[0166] This example also illustrates the use of sulfur dioxide as a
working fluid and compares its performance and operational
characteristics with ammonia as a working fluid.
[0167] The combustion turbine in this example is a General Electric
Model PG7121(EA), a model typical of peaking units used by large
electric utilities. It is assumed for this example that the
PG7121(EA) is operating under hot summertime conditions in which
the ambient temperature is 95.degree. F. Hot weather is one of the
most common times when peak load power production is needed by a
utility. Hot weather is also the most difficult of conditions under
which to generate peak load power because the combustion turbine,
which relies on ambient air for its working fluid and as a source
of oxygen for combustion, produces less power than it would
operating under moderate or cold ambient temperatures.
[0168] Hot summertime conditions are also a more difficult
application for the present invention because the external coolant
used in the condenser will also be hotter. To reflect this
difference, it is assumed for this example that condensation of the
working fluid takes place at 105.degree. F. rather than the
85.degree. F. temperature used in the previous three examples.
[0169] At 95.degree. F. ambient temperature, our estimate shows
that the PG7121(EA) produces an exhaust stream of about 2,215,000
pounds per hour at a temperature of 994.degree. F. Those skilled in
the art will readily recognize that for a heat source stream of
such large mass flow and high temperature, a bottoming cycle with
steam as the working fluid (and utilizing two or three boiler
pressures with reheat of the steam in between boiler stages) would
normally be more economic and efficient than that offered by the
present invention. However, such as team cycle lacks the
characteristics needed for peak load power generation. Startup of a
steam cycle is slow and labor intensive. Under peak load power
demand conditions, in which combustion turbines are started from a
remote control room as needed to fulfill power demand, a bottoming
cycle is needed which, similar to the combustion turbines, can be
started quickly and automatically by sending a start signal from a
distant control room.
[0170] The design parameters in this example are as follows. The
operating pressure and temperature of the working fluid as it
leaves heat exchange with the heat source stream are set at 2800
psia and 700.degree. F. These figures are based on practical and
economic considerations of the materials of construction, that is,
these figures are roughly the highest pressure and temperature
which can be safely confined in a 16-inch, schedule 80 pipe (the
approximate size for this application) made of Type 304 stainless
steel, a commonly used and economic material for handling corrosive
chemicals. Of course, this pressure and temperature selection is a
rough approximation by way of illustration only. A detailed design
and cost analysis will be needed to ascertain the most economic
design.
[0171] Because of the high temperature difference between the heat
source stream and the working fluid at the hot end of the heat
exchanger (700.degree. F. versus 994.degree. F.), a relatively
small temperature difference at the cold end of the exchanger is
possible. A temperature difference of 10.degree. F. is used. This
small difference permits more heat to be extracted from the heat
source stream and also provides an important additional benefit
discussed in the next paragraph. With ammonia as the working fluid,
the overall weighted log mean temperature difference between the
two fluids being exchanged is a very acceptable 96.degree. F., and
no internal temperature difference is less than the 10.degree. F.
design figure at the cold end. Similarly, with sulfur dioxide as
the working fluid, the overall weighted log mean temperature
difference in the heat exchanger is a very acceptable 84.degree.
F., and no internal temperature difference is less than the
10.degree. F. design figure at the cold end.
[0172] An important design consideration for a peak load power
application is that the bottoming cycle should not interfere with
the operation or performance of the combustion turbine. If heat
exchange tubing were to be permanently installed in the flowpath of
the combustion turbine exhaust gas as is normally done with a
bottoming cycle, then a fixed pressure drop across the tubing is
introduced into the design. Such a pressure drop would reduce the
power output and efficiency of the combustion turbine whether the
bottoming cycle was operated or not. This is unacceptable because
the often short duration need to generate peak load power may in
some cases not allow time to start the bottoming cycle.
Nevertheless, the power loss experienced by the combustion turbine
would always be present if the heat exchanger were present. The
present invention provides a way to avoid this problem. A design
configuration is used in which most of the combustion turbine
exhaust is drawn by an induced draft fan (device 115 in FIG. 1)
into a separate duct system which contains the heat exchange unit.
The exhaust gas that is not drawn into the duct, if any, is
discharged along its normal flowpath to the atmosphere. The induced
draft fan is located downstream of the heat exchange unit where the
exhaust gas has been cooled. As those skilled in the art will
appreciate, the very low temperature difference of 10.degree. F. at
the cold end of the heat exchanger with the present invention
causes the exhaust gas to be as cold as practicable and thus, in
turn, lowers the power usage by the fan to as low as practicable.
From the perspective of the overall system, the power consumed by
the fan is slightly less than the power that would have been lost
(i.e., not generated) in a system without a fan.
[0173] Two working fluids, sulfur dioxide and ammonia, are
presented and compared in this example. For any given set of
operating conditions of the present invention, sulfur dioxide
always produces more net power than ammonia. However, for low or
moderate temperature applications such as those presented in
Examples I and II, where the heat source temperature is less than
600.degree. F., sulfur dioxide is problematic from a mechanical
design standpoint because of condensation that occurs inside the
expansion turbine. However, for this example, where the heat source
temperature is high, the expansion turbine exhaust temperature is
raised, thereby avoiding condensation.
[0174] Table III compares the properties of the two working fluids
and illustrates the difference in the two fluids under different
operating conditions.
3 TABLE III Sulfur Dioxide Ammonia Critical temperature, .degree.
F. 316 270 Critical pressure, psia 1143 1636 Condensing pressure @
85.degree. F., psia 65 167 Condensing pressure @ 105.degree. F.,
psia 93 231 Percent liquid in expansion turbine 13.7 <0.1
outlet, after expansion from 1800 psia, 450.degree. F. to the
condensing pressure at 85.degree. F. Percent liquid in expansion
turbine None (169.degree. F. None (270.degree. F. outlet, after
expansion from 2800 psia, outlet outlet 700.degree. F. to the
condensing temperature) temperature) pressure at 105.degree. F.
[0175] The last two rows of figures in Table III illustrate the
advantage of sulfur dioxide over ammonia in a high temperature
application such as that presented in this example. Sulfur dioxide
cools more than ammonia during expansion, hence more energy is
converted to power with sulfur dioxide. At a relatively cool inlet
temperature for the expansion turbine (450.degree. F.), this
property of sulfur dioxide produces a large quantity of liquid
(13.7 percent) in the expansion turbine exhaust. In contrast,
ammonia produces almost no condensation in the expansion turbine.
At a higher initial temperature, however, this advantage for
ammonia becomes a disadvantage. As shown in the last row of the
table, ammonia only cools to 270.degree. F., which leaves a
significant quantity of heat unconverted to power. This unused heat
simply adds to the burden of heat rejection to the environment
which occurs in the condenser.
[0176] During normal and steady state operation of the present
invention with sulfur dioxide as the working fluid, the operating
conditions at points throughout the cycle are as follows (see FIG.
1 for references to line and equipment designation numbers). Stream
104 emerges from heat exchanger 101 at a temperature of 700.degree.
F. having been heated by the 994.degree. F. heat source stream.
Valve 113 is fully closed and valve 105 is fully open. At the
calculated design flow rate of 2,285,000 pounds per hour for stream
104, the expansion turbine by virtue of its design provides
resistance to flow such that the pressure in stream 104 is
maintained at about 2800 psia. Upon expansion of stream 104 through
expansion turbine 106 to a pressure of 93 psia, the working fluid,
having been cooled by work of expansion, emerges at a temperature
of 169.degree. F. (stream 108). These conditions are calculated on
the basis of obtaining 85 percent of the theoretical power from the
isentropic expansion of sulfur dioxide. Gross power generated at
the expansion turbine shaft is 43,388 kW. After deducting 2 percent
for losses in the generator, gross electric power output from
generator 107 is 42,520 kW. Stream 108 is cooled from 169.degree.
F. to its saturation temperature of 105.degree. F. and condensed in
condenser 109 by an external cooling source (such as cooling water
at 80.degree. F.). The pressure at which sulfur dioxide condenses
at 105.degree. F. is 93 psia, which sets the pressure at the
discharge of expansion turbine 106. The working fluid liquid is
also subcooled in condenser 109 to a temperature of 103.degree. F.
No accumulation or loss of fluid occurs in surge vessel 110 at
steady state. Pump 111 is a centrifugal type that provides a flow
rate and discharge pressure as prescribed by the pump's performance
curve. A centrifugal type is selected because it is practical for
the relatively large flow rate (about 3200 gallons per minute)
found in this example. A flow control valve 112 downstream of the
pump is partially closed to limit the discharge flow from pump 111
to the desired flow rate of 2,285,000 pounds per hour. Pump 111
increases the pressure of the working fluid to 2825 psia, which is
25 psi above the pressure at the entrance to expansion turbine 106.
This 25 psi difference is to allow for two separate pressure losses
between pump 111 and expansion turbine 106: (1) 20 psi for the loss
through the flow control valve downstream of pump 111, and (2) 5
psi for losses in heat exchanger 101. Electric power consumed by
the drive motor of pump 111 is about 5697 kW. The mechanical act of
pumping the working fluid warms it to about 120.degree. F. from
103.degree. F. The heat source stream leaving heat exchanger 101,
stream 114, is cooled to within 10.degree. F. of the working fluid,
or about 130.degree. F. Using an estimated 0.6 psi pressure drop
for the heat source stream as it passes through heat exchanger 101,
the power required by the induced draft fan, equipment item 115, is
estimated to be 1552 kW. Stream 103 has conditions of 120.degree.
F. and 2805 psia as it enters heat exchanger 101, completing the
cycle as described.
[0177] During normal and steady state operation of the present
invention with ammonia as the working fluid, the operating
conditions at points throughout the cycle are as follows (see FIG.
1 for references to line and equipment designation numbers). Stream
104 emerges from heat exchanger 101 at a temperature of 700.degree.
F. having been heated by the 994.degree. F. heat source stream.
Valve 113 is fully closed and valve 105 is fully open. At the
calculated design flow rate of 643,000 pounds per hour for stream
104, the expansion turbine by virtue of its design provides
resistance to flow such that the pressure in stream 104 is
maintained at aboutly 2800 psia. Upon expansion of stream 104
through expansion turbine 106 to a pressure of 231 psia, the
working fluid, having been cooled by work of expansion, emerges at
a temperature of 270.degree. F. (stream 108). These conditions are
calculated on the basis of obtaining 85 percent of the theoretical
power from the isentropic expansion of ammonia. Gross power
generated at the expansion turbine shaft is 38,535 kW. After
deducting 2 percent for losses in the generator, gross electric
power output from generator 107 is 37,765 kW. Stream 108 is cooled
from 270.degree. F. to its saturation temperature of 105.degree. F.
and condensed in condenser 109 by an external cooling source (such
as cooling water at 80.degree. F.). The pressure at which ammonia
condenses at 105.degree. F. is 231 psia, which sets the pressure at
the discharge of expansion turbine 106. The working fluid liquid is
also subcooled in condenser 109 to a temperature of 103.degree. F.
No accumulation or loss of fluid occurs in surge vessel 110 at
steady state. Pump 111 is a centrifugal type that provides a flow
rate and discharge pressure as prescribed by the pump's performance
curve. A centrifugal type is selected because it is practical for
the relatively large flow rate (about 2100 gallons per minute)
found in this example. A flow control valve 112 downstream of the
pump is partially closed to limit the discharge flow from pump 111
to the desired flow rate of 643,000 pounds per hour. Pump 111
increases the pressure of the working fluid to 2825 psia, which is
25 psi above the pressure at the entrance to expansion turbine 106.
This 25 psi difference is to allow for two separate pressure losses
between pump 111 and expansion turbine 106: (1) 20 psi for the loss
through the flow control valve 112 downstream of pump 111, and (2)
5 psi for losses in heat exchanger 101. Electric power consumed by
the drive motor of pump 111 is about 3483 kW. The mechanical act of
pumping the working fluid warms it to about 115.degree. F. from
103.degree. F. The heat source stream leaving heat exchanger 101,
stream 114, is cooled to within 10.degree. F. of the working fluid,
or about 125.degree. F. Using an estimated 0.6 psi pressure drop
for the heat source stream as it passes through heat exchanger 101,
the power required by the induced draft fan, equipment item 115, is
estimated to be 1540 kW. Stream 103 has conditions of 115.degree.
F. and 2805 psia as it enters heat exchanger 101, completing the
cycle as described.
[0178] The calculated performance numbers for Example IV with
sulfur dioxide as the working fluid are summarized as follows:
Gross electric power output=42,520 kW
Gross pumping power used=5697 kW
Gross power used by the induced draft fan=1552 kW
Net power=35,271 kW
Thermal efficiency=21.4 percent
[0179] With ammonia as the working fluid:
Gross electric power output=37,765 kW
Gross pumping power used=3483 kW
Gross power used by the induced draft fan=1540 kW
Net power=32,742 kW
Thermal efficiency=19.8 percent
Example V
[0180] This example illustrates a general case in which the heat
source stream temperature is too low for practical operation with
ammonia as the working fluid. A pressurized stream of hot water in
the liquid state at 200 psia, 300.degree. F. and flowing at a rate
of 1000 gallons per minute was selected to represent the heat
source. Although ammonia is theoretically applicable as a working
fluid because the heat source temperature of 300.degree. F. exceeds
ammonia's critical temperature (270.degree. F.), ammonia is highly
impractical because the initially cold temperature results in a
large amount of condensation of ammonia (>30 percent liquid)
inside the expansion turbine. Instead of ammonia, the working fluid
of choice for this example is chlorodifluoromethane.
[0181] Chlorodifluoromethane, which has the chemical formula
CHClF.sub.2, is commonly used in industrial and residential cooling
applications as a refrigerant and has the refrigerant designation
R-22. Chlorodifluoromethane has a relatively low critical
temperature (205.degree. F.) and a relatively low critical pressure
(721 psia), which make it ideal for relatively cold heat source
streams. At a condensing temperature of 85.degree. F.,
Chlorodifluoromethane condenses at a pressure of 172 psia, which is
similar to that for ammonia (167 psia).
[0182] As with Example I, only three process design parameters need
be selected for the cycle: (1) the temperature of the working fluid
as it leaves heat exchange (stream 104 in FIG. 1), (2) the maximum
pressure of the working fluid, and (3) the circulation rate of the
working fluid. A temperature of 270.degree. F. is selected for the
working fluid, which, given the 300.degree. F. temperature of the
heat source, provides an adequate temperature difference of
30.degree. F. for heat transfer at the hot end of the exchanger. A
working fluid pressure of 865 psia is selected, which, upon
expansion in the expansion turbine at an initial 270.degree. F.,
yields a modest and acceptable level of 0.1 percent liquid in the
expansion turbine exhaust. Finally, the circulation rate of the
working fluid was selected to provide, with benefit of properly
sized heat exchange equipment, a 50.degree. F. temperature
difference at the cold end of the heat exchanger. This rate is
calculated to be about 888,200 pounds per hour.
[0183] During normal and steady state operation of the present
invention, the operating conditions at points throughout the cycle
are as follows (see FIG. 1 for references to line and equipment
designation numbers). Stream 104 emerges from heat exchanger 101 at
a temperature of 270.degree. F. having been heated by the
300.degree. F. heat source stream. Valve 113 is fully closed and
valve 105 is fully open. At the calculated design flow rate of
888,200 pounds per hour for stream 104, the expansion turbine by
virtue of its design provides resistance to flow such that the
pressure in stream 104 is maintained at about 865 psia. Upon
expansion of stream 104 through expansion turbine 106 to a pressure
of 172 psia, the working fluid, having been cooled by work of
expansion, emerges saturated at 85.degree. F. and contains about
0.1 percent liquid of condensation (stream 108). These conditions
are calculated on the basis of obtaining 85 percent of the
theoretical power from the isentropic expansion of
chlorodifluromethane. Gross power generated at the shaft of
expansion turbine 106 is 3646 kW. After deducting 2 percent for
losses in the generator, gross electric power output from generator
107 is 3573 kW. Stream 108 is condensed in condenser 109 at a
temperature of 85.degree. F. by an external cooling source (such as
cooling water at 60.degree. F.). The pressure at which
chlorodifluromethane condenses at 85.degree. F. is 172 psia, which
sets the pressure at the discharge of expansion turbine 106. The
working fluid liquid is also subcooled in condenser 109 to a
temperature of 83.degree. F. No accumulation or loss of fluid
occurs in surge vessel 110 at steady state. Pump 111 increases the
pressure of the working fluid to 870 psia, which is 5 psi above the
pressure at the entrance to expansion turbine 106. This 5 psi
difference is to allow for a pressure drop due to flow of the
working fluid through heat exchanger 101 and the attendant piping.
Pump 111 is a type that provides a fixed volume of liquid per
revolution of its drive shaft (i.e., a positive displacement pump).
The prescribed circulation rate of 888,200 pounds per hour is set
by controlling the rotational speed of the drive shaft of pump 111.
Electric power consumed by the drive motor of pump 111 is about 634
kW. The mechanical act of pumping the working fluid warms it
slightly to about 92.degree. F. from 83.degree. F. Thus stream 103
has conditions of 92.degree. F. and 870 psia as it enters heat
exchanger 101, completing the cycle as described.
[0184] FIG. 4 shows the internal temperature profile along the
length of the heat exchanger in this example. Temperature
differences between the two streams remain acceptable throughout
the exchanger. The smallest temperature difference between the two
streams is about 17.degree. F., and the largest is about 50.degree.
F. The log mean temperature difference (LMTD) weighted over the
length of the exchanger is about 26.degree. F. Overall, FIG. 4
demonstrates the general acceptability of countercurrent exchange
used by the present invention. About 73 percent of the heat in the
heat source stream is removed.
[0185] The calculated performance numbers for Example V are as
follows:
Gross electric power output=3573 kW
Gross pumping power used=634 kW
Net power=2939 kW
Thermal efficiency=9.2 percent
[0186] Invention Parameters
[0187] After sifting and winnowing through the data from our
process simulations of the present invention herein presented, as
well as other process simulations and economic studies of the
instant, new, novel, and improved process, including methods and
means for effecting thereof, the operating variables, including the
acceptable and preferred conditions for carrying out the instant,
new, and novel invention, are summarized in Table IV below.
4TABLE IV MOST OPERATING PREFERRED PREFERRED VARIABLES LIMITS
LIMITS LIMITS* Heat source >210.degree. F.** 290-1100.degree. F.
500-700.degree. F. temperature, stream 102 Working fluid
200-1000.degree. F. 270-700.degree. F. 450-550.degree. F.
temperature, stream 104 Working fluid 450-3500 psia 760-3500 psia
1800- psia pressure, 2400 stream 104 Condensing 60-140.degree. F.
70-105.degree. F. 75-85.degree. F. temperature Percent liquid
<20 <0.1 0 in expansion turbine outlet Net power >0.5 MW
1-50 MW 2-4 MW output *The most preferred limits refer specifically
to those applications in which ammonia is used as the working fluid
and in which the equipment of the present invention (except for the
heat exchanger) is mounted on a single portable skid. **The
temperature of the heat source is limited by way of the practical
and economic consideration of the cost and availability of the
materials of construction and not limited by way of the
process.
[0188] While we have shown and described particular embodiments of
this invention, modifications and variations thereof will occur to
those skilled in the art. It is to be understood therefore that the
appended claims are intended to cover such modifications and
variations which are within the true scope and spirit of this
invention.
* * * * *