U.S. patent application number 15/936277 was filed with the patent office on 2018-08-02 for heat energy distribution systems and methods for power recovery.
The applicant listed for this patent is ElectraTherm, Inc.. Invention is credited to David C. Williams.
Application Number | 20180216500 15/936277 |
Document ID | / |
Family ID | 49993528 |
Filed Date | 2018-08-02 |
United States Patent
Application |
20180216500 |
Kind Code |
A1 |
Williams; David C. |
August 2, 2018 |
HEAT ENERGY DISTRIBUTION SYSTEMS AND METHODS FOR POWER RECOVERY
Abstract
Systems and methods are provided for the recovery of mechanical
power from heat energy sources via multiple heat exchangers and
expanders receiving at least a portion of heat energy from a
source. The distribution of heat energy from the source may be
portioned, distributed, and communicated to the input of each of
the heat exchangers so as to permit utilization of up to all
available heat energy. In some embodiments, the system receives
heat energy from more than one source at one or more temperatures.
Mechanical energy from expansion of working fluid in the expanders
may be communicated to other devices to perform useful work or
operatively coupled to one or more generators to convert the
mechanical energy into electrical energy.
Inventors: |
Williams; David C.; (Carson
City, NV) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
ElectraTherm, Inc. |
Flowery Branch |
GA |
US |
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|
Family ID: |
49993528 |
Appl. No.: |
15/936277 |
Filed: |
March 26, 2018 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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14955064 |
Dec 1, 2015 |
9926813 |
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15936277 |
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14816045 |
Aug 2, 2015 |
9840940 |
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14955064 |
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13836442 |
Mar 15, 2013 |
9115603 |
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14816045 |
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61674868 |
Jul 24, 2012 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F01K 7/18 20130101; F01K
7/16 20130101; F01K 25/08 20130101; F01K 23/00 20130101; F01K 7/20
20130101; F01K 13/006 20130101; F01K 23/065 20130101 |
International
Class: |
F01K 25/08 20060101
F01K025/08; F01K 23/06 20060101 F01K023/06; F01K 23/00 20060101
F01K023/00; F01K 13/00 20060101 F01K013/00 |
Claims
1. A system for recovering energy from an input flow of fresh air
heated by compression, the system comprising: A. more than one flow
control valve in heated air flow receiving communication with said
input flow of heated fresh air; and B. more than one primary heat
exchanger, each said primary heat exchanger in heated fresh air
flow receiving communication with at least one of said more than
one flow control valve; wherein each of said more than one flow
control valves are operative to portion, distribute, and
communicate a controllable portion of air flow comprising heat
energy from the input flow of heated fresh air to at least one of
the more than one primary heat exchangers.
2. The system of claim 1 wherein each of said controllable portions
of air flow may comprise all, some, or none of the heat energy
available from the input flow of heated fresh air.
3. The system of claim 2 further comprising one or more
intermediate heat exchanger(s) disposed between input flow of
heated fresh air and said more than one primary heat exchangers
such that air flow comprising heat energy is communicated from said
input flow of heated fresh air to each of said more than one
primary heat exchangers via at least one of said one or more
intermediate heat exchangers.
4. The system of claim 2 further comprising at least one organic
Rankine cycle (ORC) system comprising an ORC working fluid, at
least one expander, at least one condenser, and at least one
working fluid pump, wherein at least one of said more than one
primary heat exchangers is configured to communicate heat energy to
said ORC working fluid for expansion in said at least one expander
to generate mechanical power.
5. The system of claim 4 further comprising at least one electrical
power generator and wherein at least a portion of said generated
mechanical power is communicated to said at least one electrical
power generator.
6. The system of claim 4 comprising more than one ORC system and
wherein at least one of said more than one primary heat exchangers
is configured to communicate heat energy from the input flow of
heated fresh air to two or more of said more than one ORC
systems.
7. The system of claim 4 wherein at least one controllable portion
of air flow comprising heat energy is communicated from the input
flow of heated fresh air to at least one of said more than one
primary heat exchangers not configured to communicate heat energy
to any of the at least one ORC systems.
8. The system of claim 1 wherein all of said more than one air flow
control valves are in direct air flow receiving communication with
said input flow of heated fresh air.
9. A method of recovering energy from a flow of fresh air heated by
compression, the method comprising: A. receiving a input flow of
fresh air heated by compression; and B. portioning, distributing,
and communicating a controllable portion of heated air flow from
said input flow of heated fresh air via more than one air flow
control valve, thereby creating and providing more than one
controllable portion of heat energy.
10. The method of claim 9 wherein each of said more than one
controllable portion of air flow comprising heat energy portioned,
distributed, and communicated by each of said more than one air
flow control valve may comprise all, some, or none of the heat
energy available from said input flow of heated fresh air.
11. The method of claim 10 further comprising more than one heat
exchanger wherein said more than one controllable portions of heat
energy are communicated to at least one of said more than one heat
exchanger.
12. The method of claim 11 wherein at least one of said more than
one controllable portion of heat energy communicated to at least
one of said more than one heat exchanger is subsequently
communicated to another of said more than one heat exchangers.
13. The method of claim 11 further comprising the additional steps
of: A. using one or more organic Rankine cycle (ORC) system(s) each
comprising an ORC working fluid, at least one expander, at least
one condenser, and at least one working fluid pump, creating heated
ORC working fluid by communicating heat energy from at least one of
said more than one heat exchangers to said ORC working fluid of at
least one of said one or more ORC system(s); and B. expanding said
heated ORC working fluid in said at least one expander of said at
least one of said one or more ORC system(s) to generate mechanical
power.
14. The method of claim 13 wherein at least one of said one or more
ORC system(s) further comprise(s) an electrical power generator in
mechanical power receiving communication with said expander, and
said method further comprises an additional step of utilizing at
least a portion of said mechanical power to generate electrical
power via said electrical power generator.
15. The method of claim 13 wherein at least one of said more than
one controllable portions of heat energy is communicated to at
least one of said more than one heat exchangers not configured to
communicate heat energy to any of said one or more ORC
system(s).
16. The method of claim 9 wherein all of said more than one heat
transfer flow control valves in heat energy receiving communication
with said input flow of heated fresh air are in direct heat energy
receiving communication with said input flow of heated fresh
air.
17. A method for generating power from a flow of fresh air heated
by compression, the method comprising: A. receiving a input flow of
fresh air heated by compression; B. using more than one air flow
control valves in heated air flow receiving communication with said
input flow of heated fresh air to portion, distribute, and
communicate more than one controllable portions of said heated
fresh air to more than one heat exchanger via adjustment of any of
said more than one air flow control valves; C. directly or
indirectly communicating heat energy from at least one of said more
than one heat exchanger to at least one organic Rankine cycle (ORC)
system comprising an ORC working fluid, at least one expander, at
least one condenser, and at least one working fluid pump; and D.
communicating said heat energy to said ORC working fluid and
expanding said heated ORC working fluid in said at least one
expander to generate mechanical power.
18. The method of claim 17 wherein any of said more than one
controllable portions of heated compressed air portioned,
distributed, and communicated by each of said more than one air
flow control valve may comprise all, some, or none of the heated
input flow of fresh air.
19. The method of claim 18 wherein said at least one ORC system
further comprises an electrical generator in mechanical power
receiving communication with at least one of said at least one
expander and said method further comprises an additional step of
utilizing at least a portion of said mechanical power to generate
electrical power.
20. The method of claim 18 wherein at least one of said more than
one controllable portion of heated compressed air is communicated
to at least one of said more than one heat exchangers not
communicating heat energy to said ORC working fluid.
21. The method of claim 17 wherein all of said more than one heat
transfer flow control valves are in direct heated air flow
receiving communication with said input flow of heated fresh air.
Description
RELATED APPLICATIONS
[0001] This application is a Continuation and claims domestic
benefit of Applicants' pending U.S. Nonprovisional Utility patent
application Ser. No. 14/955,064, entitled "Heat Energy Distribution
Systems and Methods for Power Recovery", filed Dec. 1, 2015, which
claimed domestic benefit of U.S. Nonprovisional Utility patent
application Ser. No. 14/816,045 entitled "Multiple Organic Ranking
Cycle Systems and Methods", filed Aug. 2, 2015, now U.S. Pat. No.
9,840,940, which claimed domestic benefit of U.S. Nonprovisional
Utility patent application Ser. No. 13/836,442, entitled "Multiple
Organic Ranking Cycle Systems and Method" filed on Mar. 15, 2013,
now U.S. Pat. No. 9,115,603, which in turn claimed domestic benefit
of U.S. Provisional Patent Application 61/674,868, filed on Jul.
24, 2012. All four of said applications (14/955,064, 14/816,045,
13/836,442, and 61/674,868) are hereby incorporated in this
application by reference in their entirety and for all useful
purposes. Further, Applicant also incorporates herein by reference
in their entirety and for all useful purposes, co-owned U.S.
Nonprovisional Utility patent application Ser. Nos. 13/949,843,
14/816,046, and 15/898,648. In the event of inconsistency between
anything stated in this specification and anything incorporated by
reference in this specification, this specification shall
govern.
FIELD OF INVENTION
[0002] The present invention relates to apparatus, system, and
methods of utilizing organic Rankine cycle ("ORC") systems for the
generation of power with multiple expanders and, in some
embodiments, a common working fluid.
BACKGROUND
[0003] Many physical processes are inherently exothermic, meaning
that some energy previously present in another form is converted to
heat by the process. While the creation of heat energy may be the
desired outcome of such a process, as with a boiler installed to
Provide radiant heat to a building using a network of conduits
which circulate hot water to radiators or a furnace used for the
smelting of metals, in many other instances unwanted heat is
produced as a byproduct of the primary process. One such example is
an automobile internal combustion engine, which provides motive
force as well as significant unwanted heat. Even in those processes
in which the generation of heat energy is desired, some degree of
residual heat typically escapes or remains that can be managed
and/or dissipated. Whether generated intentionally or incidentally,
this residual or waste heat represents that portion of the input
energy which was not successfully applied to the primary function
of the process in question. This wasted energy detracts from the
performance, efficiency, and cost effectiveness of the system.
[0004] With respect to the internal combustion engine ("ICE"),
considerable waste heat energy is generated by the combustion of
fuel and the friction of moving parts within the engine. ICE
efficiency is generally less than 40%; 60% or more of the engine
fuel's energy is therefore converted to waste heat energy that is
commonly dissipated to the ICE's surroundings.
[0005] Automobiles are usually equipped with extensive systems that
transfer the heat energy away from the source locations and
distribute that energy throughout a closed-loop recirculating
system. This recirculating system usually employs a water-based
coolant medium flowing under pressure through jackets within the
engine coupled to a radiator across which the imposition of forced
air dissipates a portion of the undesired heat energy into the
environment. This cooling system is managed to permit the engine to
operate at the desired temperature, removing some but not all of
the heat energy generated by the engine.
[0006] As a secondary function, a portion of the heat energy
captured by the engine cooling system may be used to indirectly
provide warm air as desired to the passenger compartment for the
operator's comfort. This recaptured and re-tasked portion of the
waste heat energy generated as a byproduct of the engine's primary
function represents one familiar example of the beneficial use of
waste heat.
[0007] Considerable additional waste heat is expelled from the ICE
via the engine exhaust system. The byproducts of the combustion,
including gasses containing some particulate matter, exit the
engine as a result of the pressure differential between the
engine's internal pressure and the lower ambient pressure.
Considerable heat is also removed from the system in this process.
For most ICE applications, however, it is uncommon to use the heat
of the engine exhaust system for a secondary purpose. The
temperature of the exhaust flow usually exceeds that of the cooling
jacket water. However, the proportion of heat energy removed from
the engine and/or available for conversion to other purposes via
may not be similarly distributed. For example, the total available
heat energy in the jacket water may be less than, equal to, or
greater than the total heat energy contained in the exhaust gas
flow.
[0008] In addition to the cooling of ICEs, jacket water cooling
systems have been utilized in a number of other industrial
applications, including but not limited to compressor heads or
other components in which an increase in pressure, internal
friction, or other physical phenomena causes an increase in
temperature that must be removed from the system for proper
operation. In such systems, exhaust gasses may simultaneously be
generated by the same device or by an interconnected device or
system, such as the source of power for a gas compressor system. In
the case of systems that capture radiated energy including but not
limited to solar-based systems, jacket water may be used to cool
the apparatus. In some cases, this jacket cooling may be in
addition to any primary flow of media inside the system that
constitutes the primary conversion function of the system, and the
heat energy captured by the secondary cooling system may be
considered waste heat energy if it is of no use to the primary
solar-based system.
[0009] Characteristics of the heat sources that affect quality may
include but are not limited to its temperature (sufficiency and
stability), form (gaseous, liquid, radiant, etc.), the presence of
corrosive elements associated with the heat source, accessibility
for use, and the duty cycle of availability. Waste heat energy
sources are classified by grade according to these characteristics.
Prior art ORC systems prefer higher grade sources of heat that are
readily accessible, of generally high and stable temperature, are
free of contaminants, and are available without interruption. Lower
grade sources of heat, particularly those at lower temperatures,
are not as desirable and have not been fully utilized by the prior
art.
[0010] Large internal combustion engines, as another example, are
widely used in heavy industry in numerous applications. For
example, General Electric's Jenbacher gas engine division produces
a full range of engines with output power capabilities ranging from
250 kW to over 8,000 kW. By comparison, a typical mid-class
automobile engine produces about 150 kW of usable output power. The
Jenbacher engines may be powered by a variety of fuels, including
but not limited to diesel, gasoline, natural gas, biogas, and other
combustible gasses including but not limited to those produced from
landfills, sewage, and coal mines. These engines are frequently
employed to drive electric power generators, thereby converting the
mechanical energy produced from the energy of combustion into
electrical energy.
[0011] In operation, these Jenbacher engines generate tremendous
amounts of waste heat energy that has historically been dissipated
into the environment. In the case of the combined Jenbacher model
J316 engine and generator system with a rated electric power output
of approximately 835 kW, approximately 460 kW of heat energy is
lost (dissipated) in the exhaust gas at an approximate temperature
of 950.degree. F. and approximately another 570 kW is lost in the
internal cooling system with a typical jacket water coolant
temperature of approximately 200.degree. F. Of that 570 kW,
approximately 463 kW is suitable for waste heat recovery at
sufficient temperature with the remainder of such low grade as to
not be practicable for direct conversion. From this data, less than
half of the system's energy output is in the desired form (in this
case, electric power output from the system generator). In many
prior art systems, a substantial portion of the input energy
converted to heat will be lost The heat from exhaust gas generally
escapes into the atmosphere, and the recirculating jacket water is
cooled by an outboard apparatus (such as by large external
condensing radiators driven by forced air sources), which consume
additional electric power to function and further reduce the
efficiency of the system.
[0012] Additionally, the dissipation of this waste heat energy into
the environment can have deleterious effects. Localized heating may
adversely affect local fauna and flora and can require additional
power, either generated locally or purchased commercially, to
provide additional or specialized cooling. Further, the noise
generated by forced air cooling of the jacket water heat radiators
can have undesirable secondary effects.
[0013] Waste heat energy systems employing the organic Rankine
cycle (ORC) system have been developed and employed to recapture
waste heat from sources such as the Jenbacher 312 and 316
combustion engines. One typical prior art ORC system for electric
power generation from waste heat is depicted in FIG. 1. Heat
exchanger 101 receives a flow of a heat exchange medium in a closed
loop system heated by energy from a large internal combustion
engine at port 106.
[0014] For example, this heat energy may be directly supplied from
the combustion engine via the jacket water heated when cooling the
combustion engine, or it may be coupled to the ORC system via an
intermediate heat exchanger system installed proximate to the
source of exhaust gas of one or more combustion engines. In either
event, heated matter from the combustion engine or heat exchanger
is pumped to port 106 or its dedicated equivalent. The heated
matter flows through heat exchanger 101 and exits at port 107 after
transferring a portion of its latent heat energy to the separate
but thermally coupled closed loop ORC system which typically
employs an organic refrigerant as a working fluid. Under pressure
from the system pump 105, the heated working fluid, predominantly
in a gaseous state, is applied to the input port of expander 102,
which may be a positive displacement machine of various
configurations, including but not limited to a twin screw expander
or a turbine. Here, the heated and pressurized working fluid is
allowed to expand within the device, and such expansion produces
rotational kinetic energy that is operatively coupled to drive
electrical generator 103 and produce electric power which then may
be delivered to a local, isolated power grid or the commercial
power grid. The expanded working fluid at the output port of the
expander, which typically is a mixture of liquid and gaseous
working fluid, is then delivered to condenser subsystem 104 where
it is cooled until it has returned to a sufficiently liquid state
for repressurization by system pump 105.
[0015] The condenser subsystem sometimes includes an array of
air-cooled radiators or another system of equivalent performance
through which the working fluid is circulated until it reaches the
desired temperature and state, at which point it is applied to the
input of system pump 105. System pump 105 provides the motive force
to pressurize the entire system and supply the liquid working fluid
to heat exchanger 101, where it once again is heated by the energy
supplied by the combustion engine waste heat and experiences a
phase change to its gaseous state as the organic Rankine cycle
repeats. The presence of working fluid throughout the closed loop
system ensures that the process is continuous as long as sufficient
heat energy is present at input port 106 to provide the requisite
energy to heat the working fluid to the necessary temperature. See,
for example, Langson U.S. Pat. No. 7,637,108 ("Power Compounder")
which is hereby incorporated by reference.
[0016] As a result of the transfer of waste heat energy from the
combustion engine to the ORC system, these types of prior art ORC
systems serve two functions. They convert this waste heat energy,
which would otherwise be lost, into productive power; and they
simultaneously provide a beneficial, and sometimes a necessary,
cooling or condensation function for the combustion engine. In
turn, the ORC system's shaft output power has been used in a
variety of ways, such as to drive an electric power generator or to
provide mechanical power to the combustion engine, a pump, or some
other mechanical apparatus.
[0017] ORC systems can extract as much useful heat energy as they
can utilize from one or more waste heat sources (often referred to
as the "prime mover"), but owing to various physical limitations
they cannot convert all available waste heat to mechanical or
electric power via the expansion process discussed above. Similar
in some respects to the cooling requirements of the prime mover,
the ORC system requires post-expansion cooling (condensation) of
its working fluid prior to repressurization of the working fluid by
the system pump and delivery of the working fluid to the heat
exchanger. The heat energy lost in this condensation process,
however, represents wasted energy which detracts from the overall
efficiency of the system.
[0018] Prior art ORC systems capture a portion of the waste heat
energy from either the exhaust gas flow or jacket cooling water, or
a combination of both, from a prime mover but must discard a
portion of the waste heat energy that might otherwise be captured
and converted into useful mechanical and/or electrical energy. Some
heat energy is distributed within the internal processes of the
prior ORC systems, and this heat energy must be recaptured or it
will be lost, thereby decreasing efficiency. For example, the prior
art includes systems that utilize superheated fluids, including
water, and the recuperation process to increase efficiency (see,
for example, Kaplan, US 2010/0071368). This approach recaptures
heat energy that would otherwise be lost in the post-expansion
fluid during condensation and redirects that energy back to the
energy transfer components (vaporizers), which heat the system's
working fluid.
[0019] The prior art also includes, for example, the use of
multiple expanders with multiple heat sources (Biederman,
US2010/0263380), cascaded expanders (Stinger, U.S. Pat. No.
6,857,268), and other ORC system configurations with multiple
working fluids (Ast, 2010/0242476). These types of systems,
however, each add structure and processing to the basic ORC cycle
in a fashion that consumes or wastes heat energy that could
otherwise be utilized in an ORC cycle. These additional structures
also add cost to the systems.
[0020] Exacerbating the situation is the fact that these and other
prior art systems require the use of high grade waste heat. For
example, the expanders typically used in these systems require
superheated (other than wet) working fluid. As a result, their
input temperature requirements are such that high temperature waste
heat is required to properly drive the systems.
[0021] Further, these and other references teach the use of
additional components, including intermediate heat exchangers to
transfer heat energy from one portion of the system to another,
including between ORC processes that use separate working fluids of
possibly different compositions. Such intermediate components add
cost and cause the system to operate at reduced efficiency compared
to what can be attained without them.
[0022] Further, the use of cascaded heat transfer subsystems
necessary to accommodate multiple working fluids decrease the
exergy, or the heat energy, recovered from the prime mover that is
available for use by the ORC. These types of heat transfer
subsystems also increase the cost, complexity, and size of the ORC
waste heat recovery system while decreasing reliability and
requiring greater maintenance.
[0023] Some prior art combined prime mover/ORC engine applications
have utilized heat generated by the ORC condensation process in a
conventional ORC system condenser while simultaneously providing
power (electrical and/or mechanical) for various purposes. Combined
heat and power ("CHP") ORC systems have typically fulfilled a
secondary purpose by using a portion of the heat energy from the
prime mover and/or heat energy remaining in the post-expansion
working fluid. FIG. 5A depicts a prior art ORC system including
combustion engine heat energy output port 501 and condenser heat
energy output port 502.
[0024] In one prior art ORC application, residual heat extracted
from a dedicated ORC condenser during the cooling of post-expansion
ORC working fluid at condenser heat energy output port 502 is used
to provide domestic hot water, radiant heating, or both. This
process uses a conventional ORC condenser system well known in the
art. The energy flow of one such application is depicted in the
block diagram of FIG. 5A. In this application, a heat generating
engine 601 is operatively coupled to electric generator 602 and
provides waste heat energy 603 to the ORC system 604. In turn, the
ORC system 604 is operatively coupled to drive electric generator
605. Heat energy from the prime mover 601 is delivered to heat
energy output port 501 and, in some prior art systems, is extracted
to a first heat energy input port 606 (such as for radiant
heating); in addition, heat energy from the ORC condenser is
delivered to a second heat energy input port 607 (such as for hot
water heating). In those ORC systems known by the Applicant,
utilization of residual heat from the post-expansion working fluid
is intentionally extracted from the system but is not utilized for
further system optimization of the prime mover or, for example, for
heating a production material such as microorganisms to generate
biofuel.
[0025] As noted above, screw and twin screw expanders have long
been utilized in many applications in the prior art. Certain of
these types of expanders have long been capable of operating with
wet (i.e., non-superheated) working fluid. As a result, these types
of expanders have also long been utilized with heat sources and
working fluid temperatures well below the comparable temperatures
provided by high temperature heat sources and the superheated
working fluid developed in the associated ORC and its expander as a
result.
BRIEF SUMMARY OF SOME ASPECTS OF DISCLOSURE
[0026] Applicants have invented apparatus, systems and methods that
generate mechanical and/or electrical power from one or more waste
heat flows using a system of multiple heat exchangers and working
fluid expanders. In some embodiments, the expanders operate at
multiple temperatures and/or multiple pressures ("MP") and utilize
a common working fluid. In other embodiments, the expanders each
operate with separate working fluid circuits.
[0027] In certain embodiments of the system, two expanders are
utilized. This two-expander MP ORC system is a dual-pressure, or
two-pressure ("2P"), configuration. In certain embodiments of a 2P
system, one expander operates in a high-pressure ("HP") ORC cycle
and the second expander operates in a low-pressure ("LP") ORC
cycle. Both ORC cycles utilize a common working fluid comprising an
organic refrigerant or other suitable substance.
[0028] In some applications, multiple heat sources can provide
input energy and may originate from a single prime mover, such as,
for example, the jacket cooling water and exhaust flow from an
internal combustion engine. The ORC heat input may also be provided
by two or more prime movers, such as multiple ICEs and/or any other
suitable sources.
[0029] In certain embodiments, one or more heat sources provide
heat energy to more than one heat exchanger in working fluid
communication with one or more expanders. Heat from one or more
particular source(s) may be apportioned among and provided to more
than one heat exchanger to optimize overall system performance,
including but not limited to generating maximum output power
(mechanical or electrical), consuming all available waste heat
available from one or more source(s), balancing the heat
distribution to or the output power delivered by any or all of said
expanders, and the like. Attainment of any useful purpose achieved
by distribution of heat energy from one or more source(s) is
envisioned by this disclosure.
[0030] In some applications, differing heat sources can supply heat
energy to a closed loop ORC system including multiple ORC's
utilizing a wet working fluid, including as the input to and
through one or more expanders in the closed loop system. In some
systems, this can allow use of the closed loop ORC system to
recover energy from one or more heat sources that will not
superheat the ORC working fluid in one or more expanders. In turn,
this allows the ORC to avoid use of at least one superheater or
recuperator, with the associated cost and heat energy loss of such
systems.
[0031] In some embodiments, at least one of the expanders is screw
expander capable of being driven by wet working fluid. Some
instances of the screw expander constitute a twin screw expander.
In some instances, the closed loop ORC system includes at least two
ORC's, each of which have a screw expander operable with wet
working fluid. In some of these embodiments, the screw expander is
a twin screw expander.
[0032] In some embodiments, the MP ORC system accepts waste heat
energy at different temperatures. In certain embodiments, the MP
ORC system utilizes a single closed-loop cycle of organic
refrigerant flowing through up to all expanders in the system. In
some instances, the distribution of heat energy to each of the
expanders is allocated and controlled to utilize more, and, when
desired, up to and including all available heat energy and increase
or maximize the power output of the waste energy recovery process.
One or more of the expanders may be operatively coupled to one or
more generators that convert the mechanical energy of the expansion
process into electrical energy.
[0033] In some embodiments, heat energy may be coupled from one
heat exchanger to one expander on an exclusive basis. In some
embodiments, one heat exchanger my provide heat energy to more than
one expander via a common working fluid or via separate working
fluids flows within the same heat exchanger. In some embodiments,
more than one heat exchanger may provide heat energy to a
particular expander, either via a common working fluid passing
through said more than one heat exchanger(s) sequentially or via
separate working fluid flows combined at or prior to the working
fluid inlet of said particular expander. In some embodiments, more
than one heat exchanger may provide heat energy to more than one
expander via a common working fluid flow or via separate working
flows advantageously combined at or prior to the inlet(s) of said
expander(s).
[0034] The prime mover of some embodiments can be any system,
apparatus, or combination of apparatus that converts some or all of
its input energy into heat energy or waste heat energy in a form
and quantity sufficient for use by one or more MP ORC system(s). In
some embodiments, the principal or only purpose of the prime mover
can be to generate heat for the MP ORC system(s). Any heat energy
sources co-located, compatible for use with, and utilizable by one
or more MP ORC system(s), fall within the scope of the term "waste
heat" for the purpose of this application.
[0035] In some systems, a prime mover can generate and deliver
mechanical power to an electric or other power generator in
addition to providing waste heat energy for the MP ORC system(s).
In certain embodiments, a prime mover can simultaneously generate
more than one form of waste heat, such as, for example, cooling
water, hot exhaust gas, or radiated heat.
[0036] In some embodiments, a suitable prime mover may be a gas
compression system in which one or both of the compressor and a
system that cools a compressed gas line or reservoir may serve as
sources of waste heat energy for the MP ORC.
[0037] In some embodiments, a suitable prime mover may be an air
compression system such as a turbocharger for a large industrial
engine. Heat from the charge air flow may be extracted via one or
more heat exchangers and transferred to working fluid therein for
expansion in one or more ORC expanders. Further, additional heat
energy may be extracted from any compressor cooling system, such as
a forced-air cooling system or an oil, refrigerant, or other
recirculating liquid cooling system, with said extracted heat
energy provided to one or more ORC systems to heat working fluid
for expansion and energy recovery.
[0038] In some systems, the waste heat recovery system(s) include
one or more power generating system, which can be MP ORC system(s),
and one or more power receiving components, which can be but are
not limited to electric power generator(s), prime mover(s),
pump(s), combustion engine(s), fan(s), turbine(s), compressor(s),
and the like. The rotational mechanical power generated by the
power generating system(s) can also be delivered to the power
receiving components.
[0039] Waste heat energy may be captured and provided to the MP ORC
system in any practicable manner, either directly or via one or
more intermediate heat exchanger systems.
[0040] In some embodiments, the prime mover can include one or more
devices used in an industrial application, such as, for example,
electrical power generation, industrial manufacturing, gas
compression, inlet air compression for turbochargers, gas or fluid
pumping, and the like.
[0041] In some embodiments, one or more prime movers provide waste
heat energy to one or more MP ORC systems, each of which include
multiple ORC cycle operating at different pressures, different
temperatures, or different temperatures and pressures. The heat
energy is transferred from the prime mover(s) to the MP ORC
system(s) via one or more heat exchanger subsystem(s). The heat
exchanger subsystem(s) can utilize any practicable method of heat
transfer and/or media, such as, for example, water, oil,
refrigerant, air, radiation, convection, direct contact, and the
like.
[0042] In certain embodiments, a single heat exchanger subsystem
may be employed for an MP ORC system, a prime mover, a source of
heat energy from each prime mover, or for more than one MP ORC
system, prime mover, or heat energy source. Such heat exchanger
subsystems can have separate inlets and separate outlets for the
energy source(s) or a single inlet and/or outlet may be utilized
for more than one source.
[0043] In certain embodiments, one or more MP ORC systems has a
closed loop cycle to prevent intermixture of working fluid between
MP ORC systems. In some instances, one more prime movers operates
with a separate closed loop jacket water cooling system to prevent
any intermixture of jacket water between the prime mover(s) and
another system such as an MP ORC system.
[0044] In some embodiments, an exhaust gas heat recovery subsystem
may be employed to recover waste heat energy from more than one
prime mover and convey such heat energy to more than one associated
MP ORC system. In some embodiments, a heat recovery subsystem may
receive heat energy input from one or more sources and/or provide
heat energy to more than one MP ORC system.
[0045] In some embodiments, an internal combustion engine
generating sufficient waste heat energy in the form of jacket
cooling water and exhaust gas provides the energy to separate heat
exchanger subsystems coupled to a 2P ORC system. The heat energy
can be applied in prescribed amounts to one or both of the two ORC
cycles within the 2P ORC system, with the two ORC cycles operating
at different pressures. In some such embodiments, up to all of the
available waste heat energy may be utilized to the fullest extent
possible for conversion to mechanical energy by an expander and/or,
by operative connection to a generator, into electrical energy.
[0046] In some embodiments, the heat energy from more than one
prime mover may be coupled to a single MP ORC system. This can be
particularly advantageous when a plurality of prime movers are
co-located and the available heat energy from a single ICE is
insufficient to fully utilize the energy conversion capability of a
single MP ORC system.
[0047] In some systems, the heat energy from more than one prime
mover may be coupled to a plurality of MP ORC systems.
[0048] In some applications, one or more MP ORC systems constitute
the entire jacket water cooling system for the prime mover(s). In
such cases, the MP ORC systems can replace alternative prime mover
cooling systems, which consume, rather than generate, power during
operation and therefore usually have a significant cost of
operation in addition to their cost of installation. Such
power-consuming, dedicated prime mover cooling systems can have a
significantly larger footprint than an ORC system, and therefore
they may require additional physical space at the generation
facility. They may also generate noise and unwanted environmental
heat pollution as a consequence of operation. Employing one or more
ORC systems in lieu of power consuming dedicated prime mover
cooling systems, which are net consumers of power under such
circumstances, can be economically, physically, and/or
environmentally beneficial.
[0049] In some embodiments, the MP ORC system(s) provide a portion
of the cooling system for the prime mover(s) and operate in
conjunction with additional cooling systems. Electric or other
power generated by some MP ORC systems can be applied to the
operation of said additional cooling systems for the prime mover as
well as provide electric or other power for other purposes at the
site or elsewhere. This can be particularly advantageous if, for
example, the prime mover is configured to solely provide mechanical
power output and a commercial source of electric power is not
readily available.
[0050] In some embodiments, the residual heat energy remaining in
the MP ORC system after all recoverable energy has been converted
into mechanical and/or electrical energy may be employed for a
further purpose, such as, for example, building heating, domestic
and/or industrial hot water applications, the heating of bacterial
cultures for anaerobic digestion of biodegradable waste materials,
or other purpose(s).
[0051] In certain systems, the MP ORC system utilizes all or nearly
all of the available and recoverable waste heat energy available
from the prime mover(s) and converts that waste heat energy into
mechanical and/or electrical energy.
[0052] Instances of the MP ORC configuration can provide the
opportunity to couple additional heat energy input to the system so
that higher sustained power output may be realized while
simultaneously increasing system efficiency and/or fully utilizing
all available waste heat energy.
[0053] One advantage of certain disclosed MP ORC systems are their
ability to utilize waste heat energy from multiple sources, such
as, for example (meaning herein, without limitation), from sources
of different temperatures and of differing quality.
[0054] The flexibility afforded by the use of certain multiple ORC
cycles and some methods of calculating the required distribution of
heat energy from multiple sources of varying grades between the ORC
cycles can permit some systems to be optimized for a specific
application within a wide range of possibilities.
[0055] An additional advantage of some disclosed MP ORC systems is
that they can permit up to all or nearly all of the available and
recoverable waste heat energy available from one or more sources to
be utilized to a greater and, in some embodiments, the fullest
extent possible within the physical limitations of the ORC process
described in detail below. By more fully utilizing more or up to
all available and recoverable waste heat energy, the MP ORC system
provides improved, and in some instances, the greatest possible
conversion efficiency and economic return.
[0056] An additional advantage of certain MP ORC systems is that,
by more fully utilizing the waste heat energy from one or more
sources, such as for example but not limited to the jacket cooling
water from an ICE, the need for additional cooling systems can be
significantly reduced or even eliminated. In the prior art known to
the applicants, it has been necessary to dissipate remaining
available heat energy from sources that cannot be fully utilized by
the ORC; that is, available heat energy not captured and converted
by the ORC system has been be cooled via secondary means, such as,
for example, via use of radiators. These systems not only require
considerable space and expense, but they typically consume
significant electric power to drive the fans that provide the
necessary cooling. As at least some MP ORC systems can fully
extract all or nearly all available and recoverable heat energy
from its sources, such systems can provide the dual function of
generating electric power while obviating the need to consume,
e.g., electric power as required in the present art to provide the
necessary cooling.
[0057] The foregoing is a brief summary of only some of the novel
features, problem solutions, and advantages variously provided by
the various embodiments. It is to be understood that the scope of
an issued claim is to be determined by the claim as issued and not
by whether the claim addresses an issue noted in the Background or
provide a feature, solution, or advantage set forth in this Brief
Summary. Further, there are other novel features, solutions, and
advantages disclosed in this specification; they will become
apparent as this specification proceeds.
BRIEF DESCRIPTION OF THE DRAWINGS
[0058] Without limiting the invention to the features and
embodiments depicted, certain aspects this disclosure, including
the preferred embodiment, are described in association with the
appended figures in which:
[0059] FIG. 1 is a block diagram of a prior art ORC system used to
convert waste heat energy into electric power;
[0060] FIG. 2 is a block diagram of an embodiment of a 2P
multi-pressure ORC system with two expanders;
[0061] FIG. 3 is a flow chart describing the method in one
embodiment of determining the operating parameters for a 2P ORC
system;
[0062] FIG. 4 depicts the temperature versus heat energy of the
source and a hypothetical working fluid during the heat energy
transfer process from the source to the ORC working fluid in the
low pressure cycle of a 2P multi-pressure ORC system;
[0063] FIG. 5A is a block diagram of a prior art ORC system used to
convert waste heat energy into electric power including heat
extraction ports that can be used to provide heat for other
applications; and
[0064] FIG. 5B is a block diagram of the energy flow in a prior art
system including a prime mover, an ORC system used to convert waste
heat energy into electric power, and heat extraction ports for
other non-system applications.
[0065] FIG. 6A is a block diagram of one embodiment of the
invention depicting the flow of a charge air heat stream
apportioned via passage through multiple heat exchangers.
[0066] FIG. 6B is a block diagram of one embodiment of the
invention depicting the flow of a charge air heat stream
apportioned via passage through multiple intermediate heat
exchangers in heat transfer communication with ORC heat
exchangers.
[0067] FIG. 7A is a block diagram of the embodiment of the
invention depicted in FIG. 6A further comprising multiple flow
control valves to permit controlled apportionment and distribution
of the heat energy from the flow of a charge air heat stream.
[0068] FIG. 7B is a block diagram of the embodiment of the
invention depicted in FIG. 6B further comprising multiple flow
control valves to permit controlled apportionment and distribution
of the heat energy from the flow of a charge air heat stream.
[0069] FIG. 8A is a block diagram of the embodiment of the
invention depicted in FIG. 7A further comprising an alternate flow
control valve scheme to permit controlled apportionment and
distribution of the heat energy from the flow of a charge air heat
stream.
[0070] FIG. 8B is a block diagram of the embodiment of the
invention depicted in FIG. 7B further comprising an alternate flow
control valve scheme to permit controlled apportionment and
distribution of the heat energy from the flow of a charge air heat
stream.
[0071] FIG. 9 is a block diagram of one embodiment of the invention
comprising a single intermediate heat exchanger with multiple ORC
systems and flow control valves operative to permit controlled
apportionment and distribution of the heat energy from the flow of
a charge air heat stream between ORC systems and the individual
heat exchangers therein.
[0072] FIG. 10 is a block diagram of one embodiment of the
invention comprising dual intermediate heat exchangers with
multiple ORC systems and flow control valves operative to permit
controlled apportionment and distribution of the heat energy from
the flow of a charge air heat stream between ORC systems and the
individual heat exchangers therein.
DETAILED DESCRIPTION
[0073] FIG. 2 depicts a multi-pressure ORC system 200 that utilizes
two expanders 224, 242 operating at different pressures. This
configuration is an embodiment of a dual-pressure or 2P ORC
system.
[0074] By way of example and not limitation, this embodiment as
described is suitable for use with a J316 ICE engine, as specified
and manufactured by the Jenbacher Gas Engine division of General
Electric Energy, as the prime mover. Those skilled in the art will
recognize that different configurations suitable for other
applications are clearly envisioned by this invention, such as the
use of prime movers including but not limited to ICEs with power
outputs ranging from 250 kW to 8,000 kW. In this embodiment, the
J316 serves a single prime mover for the 2P ORC system and supplies
heat energy from both exhaust gas flow and jacket cooling
water.
[0075] Heat energy contained in the exhaust gas flow of the prime
mover is supplied at 201 to a thermal oil heat transfer subsystem
203 operatively coupled to first high pressure cycle evaporator 205
via a recirculating flow of oil through conduits 204 and 206.
Thermal oil heat transfer subsystem 203 may include an exhaust gas
heat exchanger such as those manufactured and sold by E. J. Bowman
Ltd. of Birmingham, UK. The oil flow through this intermediate heat
transfer system is facilitated by a pump 207. Following extraction
of up to all of the useful heat energy from the exhaust gas flow,
at least to the degree of a desired working fluid temperature
increase through the first high pressure cycle evaporator 205, the
reduced temperature exhaust gas exits the thermal oil heater
subsystem at 202. The first high pressure cycle evaporator 205 may
be a brazed plate heat exchanger such as those supplied by GEA Heat
Exchangers GmbH of Bochum, Germany.
[0076] In this particular embodiment, the temperature of the
exhaust gas at 201 is approximately 950.degree. F. and
approximately 350.degree. F. at 202. Extracting additional heat
energy from the exhaust gas flow would further reduce the
temperature at 202, resulting in the condensation and precipitation
of certain corrosive agents from the exhaust gas flow that would
damage and adversely affect the performance of the system.
So-called "bad actor" corrosive agents include residual and largely
non-combustible elements and compounds present in the fuel supplied
to the prime mover ICE, particularly those found in biogas produced
by decomposition of unknown biological and/or other materials.
Sulfur is one particularly notorious bad actor, as it may combine
to form hydrogen sulfide gas (H.sub.2S) or sulfuric acid
(H.sub.2SO.sub.4). Both are extremely corrosive and toxic and, if
allowed to precipitate within the exhaust gas heat exchanger
portion of thermal oil heat transfer subsystem 203, would
significantly degrade the performance and reduce the operating life
of that subsystem. For optimum system performance, it is desirable
that these bad actors remain in the vapor state until expelled from
the system's exhaust stack.
[0077] In one embodiment, the working fluid may be heated by any
different form of intermediate heat transfer system. In one
embodiment, the working fluid may be heated directly by the exhaust
gas without the use of an intermediate heat transfer system such as
thermal oil heat transfer subsystem 203. For example, the working
fluid may be directed through conduits and manifolds directly
exposed to the high temperature exhaust gasses, thereby heating the
working fluid directly without the use of intermediate media such
as oil.
[0078] In one embodiment, the temperature of working fluid as
heated by high pressure cycle evaporator 205 does not exceed the
saturation temperature of the working fluid vapor. One common type
of working fluid, (Genetron R-245fa), has a saturation temperature
of approximately 280.degree. F. at a pressure of 390 psia. High
pressure cycle evaporator 205, such as the GBS series of brazed
plate heat exchangers manufactured and sold by GEA Heat Exchangers
GmbH of Bochum, Germany, can be used in this embodiment to heat
this particular working fluid to 280.degree. F. at a pressure of
390 psia. As the amount of heat energy transferred to the working
fluid increases to a point, the enthalpy of the working fluid will
increase and the proportion of vaporized working fluid to liquid
working fluid will increase, but the temperature will not exceed
280.degree. F. at a pressure of 390 psia. If the system pressure is
increased without adding any additional heat energy, the working
fluid temperature will increase but the fluid maintains a constant
enthalpy. Similarly, if the system pressure is decreased
adiabatically, the working fluid temperature will decrease but the
fluid will maintain a constant enthalpy. Were a superheater to be
employed to transfer sufficient additional heat energy to the
working fluid, the enthalpy of the heated working fluid would
continue to increase until the working fluid in this example would
eventually be completely vaporized and its temperature would then
begin to exceed 280.degree. F. at the pressure of 390 psia. This
process of increasing the enthalpy of the working fluid to a point
such that the temperature of the heated working fluid exceeds its
temperature of vaporization at the operative pressure is referred
to as superheating. However, the 2P ORC system of this embodiment
utilizes a wet working fluid throughout and does not require or
utilize a superheater or superheated working fluid. Superheating
typically requires recuperation to prevent loss of heat energy in
the post-expansion working fluid and the elimination of superheated
working fluid and the recuperation process represents an
improvement over the prior art. The proportion of liquid state
working fluid to vapor state working fluid at any point in the
system may vary from completely liquid to completely vaporized
depending upon the enthalpy and pressure of the working fluid at
that point.
[0079] Heat energy contained in the jacket cooling water from the
prime mover is supplied at inlet 208 to a jacket water distribution
subsystem 210, which consists of a series flow control valves such
as the D08 series of proportional control valves available from
Continental Hydraulics of Savage, Minn. Under the control of
microprocessor-based control subsystem 219 such as the DirectLogic
series of programmable logic controllers (PLCs) available from
Automation Direct of Cumming, Ga., the control valves in the jacket
water distribution system outlet 211 provide the requisite amount
of heated jacket water to the high pressure cycle preheater 212 at
inlet 213 and to the low pressure cycle preheater and evaporator
215 at inlet 214. These preheaters and evaporators may also be
those such as the GBS series of brazed plate heat exchangers
manufactured and sold by GEA Heat Exchangers GmbH of Bochum.
Germany.
[0080] In one embodiment, the low pressure cycle preheater and
evaporator 215 described above is a single unit. In one embodiment,
the low pressure cycle preheater and evaporator 215 comprises two
separate units of similar origin and functionality. In one
embodiment, one or more separate preheaters and/or evaporators may
be used. All of the heated jacket water received at inlet 208 is
provided to either inlet 213 or inlet 214. After passing through
the high pressure cycle preheater 212 and the low pressure cycle
preheater and evaporator 215, the reduced-temperature jacket water
is returned via outlets 216 and 217, respectively, to inlet 218 of
jacket water distribution subsystem 210 where it is returned to the
prime mover via outlet 209 for recirculation. In this embodiment,
the temperature of the jacket water at outlet 211 is approximately
195.degree. F. Subsequent to the transfer of heat within the high
pressure cycle evaporator 205 and low pressure cycle preheater and
evaporator 215, the temperature of the jacket water at inlet 218 is
approximately 160.degree. F. The temperature of the jacket water
returned to the prime mover at outlet 209 is maintained within the
manufacturer's specified range for proper operation of the prime
mover. For the Jenbacher 316 ICE, this range is nominally
50.degree. C. (122.degree. F.) to 90.degree. C. (194.degree.
F.).
[0081] In one embodiment, high pressure cycle preheater 212 heats
the working fluid to the saturation temperature of the working
fluid at the operating pressure. In one embodiment, high pressure
cycle preheater 212 heats the working fluid to a temperature less
than the saturation temperature of the working fluid. For example,
high pressure cycle preheater 212 may heat the working fluid to a
temperature of 280.degree. F. at a pressure of 390 psia or any
other temperature between the working fluid temperature at inlet
221 (nominally 90.degree. F.) and 280.degree. F. However, the high
pressure cycle preheater 212 can only heat the working fluid to a
maximum temperature that, owing to limitations of the heat transfer
apparatus and laws of thermodynamics, approaches but may never
exceed the maximum temperature of the input flow of heated jacket
water at inlet 213, which in the preferred embodiment is
approximately 195.degree. F. A further discussion of the difference
between the temperature of input heat energy and the maximum
temperature of the heated working fluid output (known as the
"pinch") is provided below. Heating the working fluid to a greater
temperature will necessitate a higher grade of waste heat energy
input to jacket water distribution subsystem 210.
[0082] Control subsystem 219 is also operatively coupled to a
plurality of sensors, control valves, and other control and
monitoring devices throughout the 2P ORC system. To maintain
clarity of the Figures, these operative couplings are not depicted
in FIG. 2 but are well known to those of ordinary skill in the art.
The correct allocation of jacket water heat energy is essential for
optimization of 2P ORC operation, and the method for determining
and accomplishing this distribution as implemented by control
subsystem 219 is described more fully below.
[0083] In one embodiment, 2P ORC system 200 utilizes a single
closed loop of working fluid typically comprising a mixture of
lubrication oil and organic refrigerant suitable for heating and
expansion within the range of temperatures provided by the prime
mover. By way of example and not limitation, the refrigerant may be
R-245fa, commercially known as Genetron.degree. and manufactured by
Honeywell. The performance of the working fluid described in
association with FIG. 4 is similar but not identical to R-245fa.
However, any organic refrigerant including but not limited to R123,
R134A, R22, and the like as well as any other suitable hydrocarbons
or other fluids may be employed in other embodiments. In some
embodiments, a small percentage of lubrication oil by volume is
mixed with the refrigerant for lubrication purposes.
[0084] The working fluid is pressurized by centrifugal fluid pumps
and variable frequency drive ("VFD") motors 220 and 239
collectively referred to as VFD pumps, operatively monitored and
controlled by control subsystem 219. In one embodiment, a single
VFD pump may be utilized with suitable valves and controls to serve
both ORC cycles. Within the high pressure ORC cycle, VFD pump 220
pressurizes the working fluid to a nominal pressure of 400 psia to
cause the working fluid to flow directly through high pressure
cycle preheater 212 where it receives heat energy from a portion of
the heated jacket water, and then directly to high pressure cycle
evaporator 205 where it receives additional heat energy from the
exhaust gas flow. The combined heat energy transferred to the
working fluid as it passes through these two evaporators causes the
working fluid to change state from a heated liquid to a saturated
heated vapor. In some embodiments, the heated working fluid may be
partially in a liquid state and partially in a vaporized state. The
heated and vaporized working fluid is applied to the input of the
high pressure cycle expander 224 at an approximate pressure of
390.+-.100 psia and a temperature of 280.+-.25.degree. F. Following
expansion, the working fluid flows directly from the expander
outlet via 226 at an approximate pressure of 90.+-.30 psia and an
approximate temperature of 185.+-.20.degree. F. to a pressurized
tank serving as a high pressure cycle separator 227 where any
liquid phase portion of the working fluid in equilibrium with the
vapor phase portion of the working fluid within the separator may
be removed at the bottom. The remaining working fluid in its vapor
phase leaves the separator at or near the top and is retained for
use in the low pressure ORC cycle, described below, while the
liquid working fluid is conveyed directly via 229 to a pressurized
tank serving as a low pressure cycle separator 230. In another
embodiment, low pressure cycle separator 230 is optional and may be
omitted. In such embodiment, low pressure cycle expander outlet 244
may be directly coupled to inlet 231 of condenser subsystem 232
such as the fin fan air cooled condensers available from Guntner
U.S. LLC of Schaumburg, Ill., and outlet 229 may be directly
coupled via a throttle valve to inlet 231 of condenser subsystem
232.
[0085] In some embodiments, condenser subsystem may be a water
cooled condenser where cold water input is supplied at inlet 233
and subsequently outlet at 234. In some embodiments, condenser
subsystem 232 may be an air-cooled condenser. In some embodiments,
condenser subsystem 232 may be utilized to provide heat energy for
a desirable secondary purpose, including but not limited to the
heating of buildings, domestic or industrial hot water, heating
bacterial cultures used for anaerobic digestion of biodegradable
waste materials, and the like.
[0086] In one embodiment, condenser subsystem 232 may be cooled by
any suitable alternative means, including but not limited to those
utilizing natural environmental resources to dissipate the residual
heat energy in the working fluid. The condensed working fluid, now
in its liquid state at an approximate temperature of 84.degree. F.,
is conveyed via outlet 235 directly to working fluid receiver 237
and conveyed via 238 directly to low pressure cycle VFD pump 239.
Low pressure cycle VFD pump 239 provides the motive force
(nominally 95 psia in this embodiment) necessary to pressurize the
low pressure ORC cycle and also provides a portion of the motive
force necessary to pressurize the high pressure ORC cycle, the
balance of which is provided by high pressure cycle VFD pump 220.
In one embodiment, a single VFD pump may provide sufficient motive
force for both cycles.
[0087] Low pressure cycle VFD pump 239 provides liquid state
working fluid via 240 directly to the input of low pressure cycle
preheater and evaporator 215, which transfers heat energy from a
portion of the jacket water to the working fluid to heat and effect
a change of state of the working fluid from liquid to partially or
fully vaporized state. The fully or partially vaporized working
fluid, at approximate pressure of 90 psia and approximate
temperature of 160.degree. F., is then directly conveyed to high
pressure cycle separator 227 where it is combined with the
partially or fully vaporized working fluid previously expanded in
the high pressure cycle expander 224. The partially or fully
vaporized working fluid from both sources is applied directly to
the inlet 228 of low pressure cycle expander 242 at an approximate
pressure of 90.+-.15 psia and approximate temperature of
160.degree..+-.10.degree. F. Within the expander, the partially or
fully vaporized working fluid is expanded, removed at outlet 244 at
an approximate pressure of 27 psia and approximate temperature of
113.degree. F., directly conveyed to low pressure cycle separator
230, condenser subsystem 232, and then to VFD pump 239 for
repressurization as previously described.
[0088] High pressure and low pressure cycle expanders 224 and 242
may be any devices capable of translating a decrease in pressure
into mechanical energy, including but not limited to screw-type
expanders, other positive displacement machines such as scroll
expanders or turbines, and the like. In multi-pressure systems
including the 2P ORC system, the expanders may be of similar or
different types. In some embodiments, the expanders will be
identical machines of the twin screw configuration as taught by
Stosic in U.S. Pat. No. 6,296,461. These expanders can be of
identical characteristics or may be different.
[0089] Such units are available, for example, in the XRV series
from Howden Compressors of Glasgow, Scotland. Such expanders
utilized in association with the specific temperatures discussed in
association with FIGS. 204 herein are twin screw expanders and
operable with wet (i.e., non-superheated) working fluid from the
input through to the output of these expanders. They can thus be
operated at much lower temperatures than expanders that require
superheated working fluid. They can also be utilized with lower
temperature heat sources than those that will superheat typical
working fluids such as disclosed herein if the ORC system seeks to
utilize up to all of the available heat energy from such a
source.
[0090] High pressure cycle expander 224 is operatively coupled to
electric generator 225, such as the Magnaplus series available from
Marathon Electric of Wausau, Wis., so that the mechanical energy
produced by expansion of the working fluid may be converted into
electric power. Similarly, low pressure cycle expander 242 is
operatively coupled to electric generator 243 of similar make and
origin. Either or both generators may be coupled to the local power
grid for the purpose of delivering electrical energy to the
grid.
[0091] In some embodiments, either or both of these generators may
be used to provide power for local use, particularly when
commercial electric power is not available at the location of the
prime mover and 2P ORC system. This power may be used for the
parasitic loads of the ORC and prime mover, including the numerous
pumps and condenser systems often used to support system
operation.
[0092] The generators may be of the synchronous or asynchronous
type, depending upon the particular requirements of the system. In
one embodiment, the generators are asynchronous induction machines
with their stators operatively coupled to the commercial power grid
so that the mechanical energy imparted by the expander to the rotor
of the induction machine causes alternating current electric power
to be generated and delivered to the commercial power grid.
[0093] In one embodiment, the mechanical power from the expander
shafts may be coupled to one or more other device or system,
including but not limited to the prime mover, a pump, fans, and
other power utilizing structure or systems in lieu of being coupled
to an electric generator.
[0094] From the foregoing, it can be seen that the decrease in
pressure of the single working fluid in the 2P ORC system that
results from its expansion occurs partially in the high pressure
cycle expander 224 and partially in the low pressure cycle expander
242. This distribution and proportion of pressure reduction between
the two expanders is one substantial benefit of this invention. As
with all physical components, certain operating limitations are
imposed on the expanders due to the constraints of fabrication
materials, size, and geometry. The prior art does not allow the
capture and use of all available heat energy from the prime mover,
as is taught in the detailed embodiment described herein, or the
heat energy from other prime movers in different applications, for
conversion using a single expander and single working fluid or
multiple expanders and a shared single working fluid. Attempting to
do so would result in the dissipation of wasted heat energy in the
ORC system condenser subsystem. By dividing the expansion of highly
pressurized working fluid between two expanders, arranged in what
can be essentially a series configuration with a precise allocation
of the available input heat energy between the two interconnected
ORC cycles with a single shared working fluid, better, and in some
embodiments the most efficient, operation and output of recovered
energy is realized. Additionally, this may also be characterized as
an induction configuration with two sources of fully or partially
vaporized working fluid supplied to the low pressure cycle expander
242.
[0095] ORC waste heat recovery systems can be inherently
inefficient due to a number of factors. Notably, the physical
characteristics of the chosen working fluid can limit the range of
temperatures within which the ORC system can effectively convert
heat energy via the expansion of pressurized working fluid vapor.
Effective heat energy transfer through the heat exchange
subsystems, including the thermal oil heat transfer subsystem 203,
high pressure cycle evaporator 205, and low pressure cycle
preheater and evaporator 215 may each approach 80% only under ideal
conditions and may actually yield lower performance than 80%. When
cascaded, these sub-unity efficiencies are multiplied and yield an
even lower total effective transfer (80% of 80% is 64%). Further,
the use of recuperation processes within an ORC system constitute
an attempt to recover a portion of excess heat energy that has
previously be applied to the system but is not useful for
conversion to electrical or mechanical energy and is therefore
potentially wasted. As with any thermal process, recuperation is
not fully efficient so heat energy is inevitably lost. As a result,
in these types of prior art systems much of the available waste
heat energy produced by the prime mover is not actually being
recovered and transferred to the working fluid. Further, there are
significant heat losses within the system due in large measure to
the considerable residual heat energy that remains in the
post-expansion working fluid and which must be dissipated by the
condenser system prior to repressurization by the VFD pump(s). The
combined effect of these various losses applied to a prior art ORC
system depicted in FIG. 1 that utilize a single twin screw
expander, evaporator, and condenser as generally described above
along with the same working fluid (R-245fa) can achieve a nominal
efficiency of approximately 7% in sustained operation when supplied
with the waste heat energy available from a suitable prime mover,
such as the Jenbacher J316 in one embodiment taught herein.
[0096] Embodiments of 2P ORC specified in FIGS. 2-4 and associated
text above can improve, and in some embodiments dramatically
improve, upon this performance. When supplied with the waste heat
energy available from a Jenbacher J316 as the specified prime mover
to the particular system identified above, approximately 921 kW of
recoverable waste heat energy from exhaust gas above 356.degree. F.
and jacket cooling water heat is available for recovery and use by
the 2P ORC system. Approximately 458 kW is available from the
exhaust gas flow and the remaining 463 kW is present in the jacket
water. When all of the available 458 kW of waste heat energy from
the exhaust gas flow is provided to the high pressure cycle
evaporator 205 via thermal oil heat transfer subsystem 203, 216 kW
of available waste heat energy from the jacket cooling water is
applied to high pressure cycle preheater 212, and the remaining 247
kW of available waste heat energy from the jacket cooling water is
applied low pressure cycle preheater and evaporator 215, the 2P ORC
system can produce at least approximately 45 kW of electric power
from high pressure cycle generator 225 and another 58 kW of
electric power will be produced by low pressure cycle generator
243. The combined 103 kW of electric power generated by the 2P ORC
system constitutes an overall conversion efficiency of 11.2% of the
waste heat energy of 920 kW available from the prime mover.
Accordingly, the 2P ORC system provides an increase of 58% compared
to the nominal 7% conversion efficiency of the present art system.
This represents a very significant improvement by industry
standards.
[0097] Additionally, the prior art multiple ORC superheating
systems inherently allocate available heat energy in a fashion that
cannot be converted and therefore, in some embodiments, is
recovered by the recuperation process to salvage some efficiency.
Since, however, the superheating/recuperation process itself
imposes substantial energy loss to drive the process, the 2P ORC
system specified in association with FIGS. 2-4 is substantially
more efficient than these types of processes because all or in any
event more available heat is allocated to generating power from the
specified closed wet working fluid multiple ORC system.
[0098] Another significant advantage of the specified 2P ORC system
is its ability to fully utilize up to all of the recoverable waste
heat energy available in the jacket water of a suitably-matched
prime mover. In prior art systems known to the applicants, only a
portion of the heat energy in the jacket water can be utilized and
the remainder is cooled through the use of conventional radiators
that require additional electric power to operate the cooling fans.
In the specified embodiment of this specification, however, the 2P
ORC system is combined with waste heat generated by, for example, a
widely-used prime mover (such as the Jenbacher J316 internal
combustion engine) so that up to all of the available heat energy
in the jacket water flow may be fed to the 2P ORC system for waste
heat energy conversion into electric power. This can obviate the
need for a traditional radiator system to support the prime mover
that would consume rather than generate electric power. In
addition, a substantial portion of the waste heat in the exhaust
gas flow can be captured and converted by the specified 2P ORC
system and others disclosed herein. Embodiments of these systems
also can reduce and, in some embodiments, minimize thermal
pollution of the environment.
[0099] The distribution of waste heat energy from each source to
each of the two ORC cycles in the 2P ORC system is an operating
condition that can be calculated and maintained in order to achieve
desired, and in some embodiments, optimal performance. The method
of determining the distribution of heat energy between the high and
low pressure cycles also overcomes the limitations of the prior art
which require heat recuperation from the working fluid to minimize
losses and therefore constitutes a significant improvement over the
prior art. The method may also be utilized to determine and
maintain any desired lesser degree of utilization of available
waste heat available from the prime mover at the most efficient
point of system operation. In addition the following description,
the method of determining the 2P ORC system control and set points
is provided as a flow chart in FIG. 3.
[0100] The first steps in the iterative method of determining the
control and set points for 2P ORC system operation require the
computation of the available heat energies in the exhaust gas flow
and the jacket cooling water (301, 302). For the exhaust gas, the
temperature differential T(ex) between the exhaust gas flow T(ex_1)
at the input 201 and T(ex_2) at the output 202 to the thermal oil
heat transfer subsystem 203 may be measured if such apparatus is
available for measurement under operating conditions. If said
apparatus is not available, the available heat energy from the
exhaust gas flow may be determined from the manufacturer's
specification data for the prime mover. If neither is available,
the values may be estimated based on best available information,
recognizing that errors may be introduced by inaccurate estimations
and that further refinement and parameter adjustment will likely be
required to compensate for difference between estimated and actual
values later realized in practice.
[0101] For the jacket water, the same temperature differential
between T(jw_1) at the input 208 and T(jw _2) at the output 209 of
the jacket water distribution subsystem 210 may be measured,
calculated, or estimated using best available resources (303).
[0102] The mass flow rates M(ex) of the exhaust gas flow and M(jw)
of the jacket water flow of the prime mover may be measured,
calculated, or estimated based on best available information
(304).
[0103] The heat energy Q(ex) contained in the exhaust gas is
defined as
Q ( ex ) = M ( ex ) .intg. T ( ex_ 2 ) T ( ex_ 1 ) CpdT
##EQU00001##
where Cp is the specific heat of the exhaust gas mixture, which is
generally calculated based on the composition of the exhaust gas
and dT is the variable of integration. Assuming that the
temperature differential is sufficiently low so that Cp may be
considered to be constant at its mean value, Q(ex) may be
calculated (305) via
Q(ex)-M*Cp*.DELTA.T(ex)
where T(ex)=T(ex _1)-T(ex 2). The minimum final temperature of the
exhaust gas, T(ex_2), is normally set by the engine manufacturer at
some safe level above the acid dew point temperature of the gas
depending on the fuel used. As previously described, cooling the
exhaust gas below the acid dew point will likely cause damage,
including corrosion to the engine exhaust system and waste heat
recovery heat exchanger.
[0104] The temperature of the heated working fluid may approach
that of the waste heat source but never be able to reach it due to
the limitations imposed by the Second Law of Thermodynamics and the
physical limitations of heat exchangers used to transfer the heat
from the source to the working fluid. As a principal consequence,
the final temperature of the working fluid being heated can never
reach the highest temperature of the source being cooled.
[0105] FIG. 4 is a general depiction of the heat energy versus
temperature of the source heat and working fluid during a heat
transfer process at a pressure similar to that which may occur in
the low pressure ORC cycle. The data depicted in this figure is
illustrative of the performance of some embodiments but is not
meant to be an accurate numerical representation of any particular
embodiment. However, the properties of the example working fluid
closely resemble those of R-245fa Genetron refrigerant which
exhibits a saturation temperature of 70.degree. C. at a nominal
pressure of 90 psia as may exist at inlet 228 to low pressure cycle
expander 242. Line segment 401 represents the source heat and
segment 402 represents the working fluid. Point 404 depicts the
state of the jacket water at inlet 214 and point 403 represents the
state of the jacket water at outlet 217 of low pressure cycle
preheater and evaporator 215. In this example, the jacket water
experiences a decrease in temperature of approximately 35.degree.
C. (from 100.degree. C. to 65.degree. C.). In a similar manner,
point 405 represents the state of the working fluid at inlet 240
and point 406 represents the state of the working fluid at outlet
241 of low pressure cycle preheater and evaporator 215. Along this
path, it can be seen that the temperature of the working fluid
increases from 30.degree. C. to 70.degree. C., which in this
example is the temperature at which the working fluid begins to
vaporize at the liquid saturation temperature. Although the
temperature does not increase beyond this vaporization temperature
in this example, the heat energy content of the working fluid
continues to increase as it receives additional heat energy from
the jacket water and the working fluid is increasingly
vaporized.
[0106] During this heat transfer process, the paths representing
the working fluid heating and jacket water cooling processes do not
intersect, lest there be no additional heat transfer between the
source and working fluid, in accordance with the Second Law of
Thermodynamics. That is, the temperature of the working fluid can
never equal that of the waste heat energy input and will always be
lower by a certain amount. The temperature at the closest distance
between these two paths, point 407, is normally referred to as the
"pinch point". It is the minimum temperature difference between the
source and working fluid at any point in the heat exchanger. In the
design of ORC power plant evaporators, condensers, heat exchangers,
and the like, the pinch point is used to determine the pressure,
temperature and mass flow of the working fluid leaving the heat
exchanger.
[0107] In some embodiments, the pinch may be selected to be as low
as 3.degree. C. and as high as 10.degree. C. However, the pinch is
usually selected by ORC design engineers to be approximately
5.degree. to 10.degree. C. depending on the absolute temperature of
the source. The pinch value depicted in the example of FIG. 4 is
approximately 5.degree. C. Selection of a larger pinch value
reduces system efficiency while selection of a pinch value that is
too small increases surface requirements of the heat exchanger and
corresponding cost. Since the temperature of the waste heat energy
flow decreases as it passes through the evaporator, in the
preferred embodiment the working fluid output is in closest contact
with the waste heat energy input and the working fluid input in
closest contact with the waste heat energy output
(counterflow).
[0108] In one embodiment, the heat contained in the prime mover's
exhaust gas is applied to high pressure cycle heat exchanger 205
either directly or via thermal oil heat transfer subsystem 203, and
the design conditions of the high pressure ORC cycle are generally
set by the temperature and pressure specifications and limitations
of the expander. Those limits are imposed by the heat exchanger's
pinch point. In particular, the temperature and pressure of the
working fluid heated by the exhaust gas flow may not exceed the
rated values for the expander's inlet.
[0109] Having determined the heat energy of the exhaust gas and
assuming that all of this heat is transferred to the working fluid,
the mass flow rate of the working fluid M(wf) may be computed (306)
via
M(wf)=Q(ex)/.DELTA.H(wf hpe)
where H(wf_hpe) represents the difference in the enthalpy, or total
energy, of the working fluid between the high pressure cycle
evaporator 205 outlet 223 and inlet 222 which corresponds to a
temperature approximately 5.degree. C. below the maximum
temperature of the low temperature source. In other words, the
working fluid mass flow rate can be determined by the amount of
exhaust heat used and by the minimum and maximum enthalpy of the
working fluid heated either directly or indirectly (via thermal oil
loop) by the exhaust gas.
[0110] The total heat energy available from all jacket cooling
water is typically provided by the engine manufacturer and also may
be calculated (307) via
Q(jw_tot)=M(jw)*Cp*.DELTA.T(jw)
where T(jw) represents the difference in the temperature of the
jacket cooling water between the inlet 208 and the outlet 209 of
the jacket water distribution subsystem 210.
[0111] As previously described, waste heat energy from the jacket
cooling water may be provided to the high pressure ORC cycle via
the high pressure cycle preheater 212 that receives a portion of
the jacket cooling water from jacket water distribution subsystem
210, depending on the maximum temperature of the jacket water. The
amount of jacket water heat energy required for the high pressure
cycle may be calculated (308) via
[0112] Q(jw hp)=M(wf)*.DELTA.H(wf_hpp)
where H(wf_hpp) represents the difference in the enthalpy of the
working fluid between the outlet 222 and the inlet 221 to high
pressure cycle preheater 212.
[0113] The quantity of jacket water provided to the high pressure
cycle by jacket water distribution subsystem 210 and control
subsystem 219 is determined by the temperature difference of the
jacket water circuit as specified by the manufacturer of the prime
mover. That mass flow rate may be calculated at the outlet 222 of
high pressure cycle preheater 212 (309):
M(jw hp)=(Q(jw hp)/(.DELTA.T(jw)*Cp)
[0114] VFD pump 220 controls the pressure at the input to high
pressure cycle expander 224, and via control subsystem 219, the
mass flow rate of the working fluid in the high pressure cycle is
set to achieve the desired temperature and pressure at the inlet of
high pressure cycle expander 224.
[0115] The total waste heat energy contained in the jacket water
available for the low pressure cycle is the difference between the
total jacket water heat available and that already applied to the
high pressure cycle preheater 212 as calculated above:
Q(jw_lp)-Q(jw_tot)-Q(jw _hp)
[0116] The temperature and pressure at low pressure cycle expander
inlet 228 for optimal system performance may now be determined
iteratively via the following method:
[0117] 1) Assume that the temperature of the vaporized working
fluid T(wf_v) is equal to the minimum temperature of the jacket
water T(jw_pinch) in the low pressure cycle. This is equivalent to
setting the initial value of the pinch in the cycle to zero
(310).
[0118] 2) Calculate the mass flow rate of the working fluid in the
low pressure cycle (311) via
M(wf_lp)=Q(jw_)/.DELTA.H(wf_lpe)
where H(wf_lpe) represents the difference in enthalpy of the
working fluid leaving the low pressure cycle preheater and
evaporator 215 at 241 (where its enthalpy is maximum) and at the
entry to the low pressure cycle preheater and evaporator 215 at
240.
[0119] 3) Using the working fluid property tables, determine the
enthalpies (312): a) H(wf_cond) of the working fluid in the low
pressure cycle at the outlet 235 of condenser subsystem 232, b)
H(wf_v) at the point of initial vaporization (saturated liquid),
and c) H(wf_hps) at high pressure cycle separator 227 inlet flow
241.
[0120] 4) Calculate heat addition at the pinch point Qp (313):
Qp-[(H(wf_v)-H(wf_cond))/(H(wf_hps)-H(wf_cond))]*Q(jw_lp)
[0121] 5) Because
Qp=M(jw_lp)*Cp* (T(jw_pinch)-T(jw_o))
we may calculate (314)
T(jw_pinch)-(Qp/(M(jw_lp)*Cp))-T(jw_o)
where T(jw_pinch) is the temperature of the jacket water at the
pinch point and T(jw_o) is the temperature of the jacket water at
the outlet 217 of low pressure cycle preheater and evaporator
215.
[0122] 6) Compare (315) T(jw_pinch) to T(wf_v). If the difference
is less than 5.degree. C. (316) (the desired pinch value), reduce
T(wf_v) by 2.degree. C. (317) and repeat the iteration. If the
difference between T(jw_pinch) and T(wf_v) is greater than
5.degree. C. (318), increase T(wf_v) by 2.degree. C. (319) and
reiterate.
[0123] 7) Continue the iteration until the pinch
(T(jw_pinch)-T(wf_v)) is 5.degree. C. plus or minus 1.degree.
C.
[0124] Finally, once the parameters of the low pressure cycle have
been determined in this manner, the pressure at the high pressure
cycle expander outlet 226 may be set to the pressure of the low
pressure cycle expander inlet 228 (320). In one embodiment, one or
more control valves or other means of controlling the pressure may
be incorporated in the system.
[0125] With respect to the depiction of heated extraction ports in
the prior art systems depicted in FIGS. 5 and 6, the same
possibilities exist for MP ORC systems. The condenser subsystem 232
may be replaced, in whole or in part, by an alternate subsystem
that utilizes the residual heat energy present in the
post-expansion working fluid for any other useful purpose.
[0126] In addition to the use of jacket water distribution
subsystem 210 disclosed in detail above, other distribution systems
and apparatus may be used to portion, distribute, and communicate
heat energy from one or more sources other than jacket cooling
water to more than one heat exchanger in a myriad of other
applications. In embodiments where heat energy may be communicated
via a recirculating liquid medium such as water, oil, an organic
compound, or an inorganic compound, a distribution system highly
similar to that of the jacket water distribution subsystem 210 with
both an inlet 208 and an outlet 209 operative to provide
closed-loop circulation. In other embodiments where there is no
requirement for or advantage in recirculating a heat transfer
media, including but not limited to heat transfer from exhaust
gasses, heat transfer media may be expelled from the outlet of said
distribution system once the available heat energy has been
extracted.
[0127] Embodiments that consume up to all of the available heat
energy from jacket water of internal combustion engines provide a
number of concurrent advantages and solve several problems at once.
Heat must be extracted from the recirculating jacket water and
dissipated to maintain engine operation within the manufacturer's
specified range. This is typically accomplished in the known art
via the use of one or more conventional radiators, often cooled via
forced air flow provided by the consumption of additional energy in
the form of electricity to drive large fans, rotational energy
generated by the engine via fuel consumption, or the like. The
recovery of rotational power via expansion of a working fluid
heated by said jacket water performs the identical function as does
a conventional radiator but also generates mechanical or electrical
power rather than consuming power as does a conventional
radiator.
[0128] A similar situation exists in the use of turbochargers,
particularly those associated with large industrial engines. In a
turbocharged engine, the intake airflow is intentionally increased
over that provided to a normally-aspirated engine via the use of
compression means, often comprising one or more turbines, to
increase the density of the charge air supplied to the engine for
combustion. The increased density of the charge air, and
particularly the presence of additional oxygen necessary for fuel
combustion, enables a turbocharged engine to deliver greater output
power than a comparable engine with a normally-aspirated combustion
system.
[0129] However, the compression of charge air in a turbocharger
dramatically increases its temperature, which in turn creates other
problems. The density of a gas is inversely proportional to its
temperature, and at least a portion of the increased density
provided by the compression of turbocharging is offset by the
reduction in density due to increased temperature. Further,
increasing the temperature of the intake airflow also increases the
combustion temperature within the engine as well as the temperature
of the resulting exhaust gas. Finally, increased temperatures
within engine combustion chambers generally results in the
production of increased levels of undesired exhaust gas products,
including mono-nitrogen oxides such as nitric oxide and nitrogen
dioxide generally referred to as NO.sub.x.
[0130] For the purposes of this disclosure, the term "charge air"
is intended to refer to any stream of compressed air at an elevated
temperature, thereby encompassing comparable systems, techniques,
and processes also associated with heated streams of compressed air
or gasses including but not limited to "scavenge air",
"intercooling", and the like.
[0131] Industrial turbocharging applications that require
significant power generate significant quantities of heat at
temperatures well-suited for recovery of mechanical or electrical
power. For example, a single internal combustion engine such as the
K80MC-S9 dual fuel 2-stroke engine manufactured and sold by MAN
Diesel & Turbo SE generates in excess of 40 MW of output power,
and in doing so requires potential charge air cooling at a rate on
the order of 17 MW, depending in part on the temperature of the
ambient intake air. What is needed is a solution that consumes heat
from the compressed charge air stream, thereby reducing its
temperature to avoid the adverse consequences identified above.
Applicant's system of input heat distribution discussed previously
with respect to heated jacket water distribution provides a novel
solution to this problem while providing the advantages of
mechanical and electrical power described elsewhere herein.
[0132] The disclosure that follows draws heavily upon the basic
principles of mechanical and electrical power recovery from heat
source(s) disclosed above in great detail. While the differences
between the previously-presented embodiments and those disclosed
below will be identified to the greatest extent possible, common
elements of said embodiments will be presumed to be understood from
the entire scope of this specification. All material presented
elsewhere herein is intended to apply to the following embodiments
as well unless specifically disclaimed as not being pertinent to
any particular embodiment.
[0133] Whenever discussed with respect to the following
embodiments, the interchangeable terms "working fluid" and "ORC
working fluid" are intended to apply to any medium suitable for use
in an ORC system for heating and expansion to generate mechanical
or electrical power via an expander of any known configuration,
including but not limited to a twin screw expander capable of
operating with a partially-vaporized "wet" working fluid. In some
embodiments, said working fluid is an organic refrigerant disclosed
in detail elsewhere herein. Additionally, the term "heat transfer
medium" refers to any medium suitable for use in transferring heat
energy from a source to a destination, particularly via the use of
one or more heat exchanger(s). Any gas, liquid, or combination
thereof suitable for this purpose is envisioned by the scope of
this disclosure, including but not limited to air, other gasses,
water, oil, organic compounds, inorganic compounds, other liquids,
or any combination of one or more of the foregoing. Whenever any
heat transfer medium is referred to as "heated transfer medium" or
simply a "heated medium", it is intended to denote that said medium
has previously received heat energy from a source in its capacity
as an agent for the transfer of heat energy from a source to a
destination, usually between the source of heat such as a heat
stream or between a first heat exchanger serving as a source of
heat energy and either a second heat exchanger or a condenser
serving as the destination of said heat energy. When a "separate
heat transfer medium" is disclosed, it should be appreciated that
this refers only to the fact that the flow of said separate medium
is not commingled with the flow of any other heat transfer medium.
Separate heat transfer media may be of identical or different
composition independent from that of any other heat transfer media
without limitation, and the only connotation that should be drawn
from a disclosure of "separate" medium is that said medium is a
quantity distinct from any other medium.
[0134] With reference to FIG. 6A, an input heat stream 601 is
applied to heat source input 615 of heat exchanger 616. In general,
heat stream 601 may comprise a flow of any medium or combination of
media suitable for use in transferring heat energy to the working
fluid in the organic Rankine cycle (ORC) system comprising heat
exchangers 616 and 622, expander 626, condenser 630 and associated
components including but not limited to radiator 633 and pump 635,
and ORC system pump 638. In one embodiment, heat stream 601
comprises a flow of compressed charge air communicated from a
turbocharger or any similar source of compressed air. In other
embodiments, heat stream 601 may comprise a flow of any other
pressurized or unpressurized gas or any liquid medium of sufficient
temperature and mass flow rate so as to increase the enthalpy of
the working fluid to at least a semi-vaporized state suitable for
expansion in expander 626. For the purpose of describing the basic
operation of the embodiment depicted in FIG. 6A, the heat stream
will be presumed to be a flow of compressed charge air of elevated
temperature and density produced by and communicated from a
turbocharger for use in an internal combustion engine. As such, the
temperature of heat stream 601 is at its greatest temperature at
heat exchanger input 616.
[0135] As the heated media of heat stream 601 passes through heat
exchanger 616, a first portion of heat energy is transferred from
said heat stream to the working fluid flowing through said heat
exchanger in the opposite (counterflowing) direction, ensuring that
the highest temperature media in the heat stream is proximate to
the working fluid at the working fluid output 624 of heat exchanger
616. Said heated working fluid is communicated to working fluid
inlet 625 of expander 626 where it is expanded, thereby generating
rotational mechanical energy communicated in this embodiment to
electric generator 627. Expanded working fluid exits expander 626
at outlet 628 and is communicated to condenser 630 at condenser
inlet 629. Condenser 630, radiator 633, and pump 635 comprise a
condensing and cooling system for the expanded working fluid
wherein the working fluid is cooled to a sufficiently liquid state
to permit repressurization by ORC system pump 638. In some
embodiments, the condensing and cooling system may comprise a
working fluid collector or other reservoir for condensed working
fluid. In other embodiments, said condensing and cooling system
does not required a working fluid collector or other reservoir. In
one embodiment, the condensing and cooling system may comprise one
or more forced air radiators, one or more liquid cooled radiators,
or any other apparatus known in the art wherein residual heat may
be extracted from expanded working fluid. FIG. 6 depicts a
circulating system whereby a separate heat transfer medium in heat
receiving communication with the expanded working fluid is
circulated under pressure from pump 635 through a heat transfer
medium input 636, heat transfer medium output 631, radiator 633
input 632, and radiator output 634. The individual elements shown
in FIG. 6A are not limiting on the scope of this disclosure, as any
known heat exchanger apparatus known in the art or later developed
is envisioned for the purpose at hand. Cooled working fluid exits
the condensing and cooling system at outlet 637 and is communicated
to ORC system pump 638 for repressurization.
[0136] Cooled and pressurized working fluid is communicated first
to heat exchanger 622, functioning as a preheater, and then to heat
exchanger 616. Heat stream 601, after having flowed through heat
exchanger 616 via heat source input 615 and heat source output 617,
is now communicated to heat exchanger 622 at heat source input 621.
As a first portion of the heat energy present at heat source input
615 of heat exchanger 616 has already been transferred to the
working fluid therein, the temperature of the working fluid in heat
exchanger 622 is lower than that of the working fluid in heat
exchanger 616. However, a second portion of heat energy is
communicated to the counterflowing working fluid entering heat
exchanger 622 at working fluid input 639 and exiting at working
fluid output 640, thereby increasing the temperature of said
working fluid which is then communicated to heat exchanger 616 via
working fluid input 641.
[0137] After having passed through heat exchangers 616 and 622,
heat stream 601 is next communicated from heat source output 623 of
heat exchanger 622 to a supplemental condensing and cooling system
comprising heat exchanger 609, radiator 644, pump 646, and a
separate heat transfer medium circulating therein under pressure
supplied by pump 646. The partially-depleted pressurized heat
stream 601 enters heat exchanger 609 at heat source input 608 and
exits at output 610, whereupon it is communicated to the engine air
inlet 611 for injection into the combustion chambers. During its
passage through heat exchanger 609, a third portion of heat energy
is transferred from heat stream 601 to the separate heat transfer
medium which is subsequently cooled via radiator 644, which may
also be any apparatus known in the art or later developed for the
purpose at hand.
[0138] During its passage through heat exchangers 616, 622, and
609, three portions of heat energy have been extracted from heat
stream 601, thereby reducing its temperature. The heat energy
transferred to the ORC working fluid via heat exchangers 616 and
622 have been utilized to heat said working fluid for expansion in
expander 626 and generation of mechanical and electrical power.
However, in some embodiments, the temperature of heat stream 601
exiting heat exchanger 622 at output 623 is not yet optimally
conditioned for application to the engine and further heat must be
removed; this function is provided by the supplemental condensing
and cooling system comprising heat exchanger 609. In view of
limitations on the temperature and mass flow rate necessary for
applications involving ORC systems, and the occasional need to
remove the ORC system from operation for maintenance, the
supplemental condensing and cooling system is configured to remove
any and all remaining excess heat energy from heat stream 601 so as
to present the compressed charge air to the engine air inlet 611 at
the desired temperature.
[0139] With reference now to FIG. 6B, a system further comprising
intermediate heat exchangers 603 and 606 is depicted. In this
embodiment, heat stream 601 passes through each of said
intermediate heat exchangers in sequence and communicates each of
the first and second portions of heat energy from said heat stream
to a heat transfer medium flowing therein. The heat transfer medium
of each intermediate heat exchanger is in further heat transfer
communication with each of the heat exchangers depicted in the
previously-described embodiment. For example, a first portion of
heat energy is extracted from heat stream 601 via passage through
intermediate heat exchanger 603 during its passage from heat source
input 602 to heat source output 604, and said first portion of heat
energy is communicated to the separate heat transfer medium
circulating between media input 613, media output 614, heat source
input 615 of heat exchanger 616, and heat source output 617 of heat
exchanger 616 under pressure from pump 612. Subsequently, a portion
of the heat energy present in the heat transfer media received at
input 615 of heat exchanger 616 is communicated to the ORC working
fluid as previously described before exiting at output 617 for
continuous recirculation.
[0140] In an identical manner, the second portion of heat energy is
extracted from heat stream 601 during passage between heat source
input 605 and heat source output 607 of heat exchanger 606,
whereupon said second portion of heat energy is communicated to the
heat transfer medium circulating between media input 619, media
output 620, heat source input 621 of heat exchanger 622, and heat
source output 623 of heat exchanger 622 under pressure from pump
618. Subsequently, heat energy present in the heat transfer media
received at input 621 of heat exchanger 622 is communicated to the
ORC working fluid as previously described before exiting at output
623 for continuous recirculation.
[0141] The use of intermediate heat exchangers generally introduces
a certain additional amount of heat loss between each stage since
any transfer of heat energy is subject to some degree of
unintentional dissipation. However. In some embodiments, it may be
impractical or unfeasible to locate the ORC system sufficiently
proximate to the flow of compressed charge air to achieve direct
heat transfer communication as depicted in FIG. 6A. For example,
the necessity for long working fluid conduits between the ORC and
the charge air stream may require an excessive quantity of working
fluid that would be problematic for various reasons. In such cases,
the ability to communicate said separate heat transfer media,
including but not limited to a liquid such as an oil or other
inexpensive and stable fluid with appropriate physical and thermal
characteristics, from an intermediate heat exchanger to each of the
ORC heat exchangers may be advantageous. Additionally, the reduced
size and configurability of said intermediate heat exchangers may
be preferred in some physical installations due to constraints on
available space, safety concerns, and the like.
[0142] The embodiments of FIGS. 6A and 6B comprise a direct and
sequential flow of heat stream 601 through more than one heat
exchangers. While this provides apportionment of the heat energy
into at least three portions, the previous embodiment does not
permit the distribution of heat energy from said heat stream to
each of the heat exchangers in a precise proportion as may be
desired. With reference to FIG. 7A, an embodiment is depicted
further comprising a series of valves added to the embodiment of
FIG. 6A. Specifically, heat stream 601 may be apportioned by valve
701 to direct a first applied portion of said heat stream received
at valve input 701A to the heat source input 615 of heat exchanger
616 via valve output 701B and to direct the first remaining portion
of said heat stream to subsequent heat exchangers via valve output
701C.
[0143] Subsequently, said first remaining portion of heat stream
601 received at input 702A of valve 702 may be further divided into
a second applied portion directed to valve output 702B for
injection into heat source input 621 of heat exchanger 622 via
valve 703, where it is combined with the flow between heat source
output 617 of heat exchanger 616 and heat source input 621 of heat
exchanger 622, with a second remaining portion of heat stream 601
provided by valve output 702C for use by subsequent heat
exchangers.
[0144] Next, the second remaining portion of heat stream 601
received at input 704A of valve 704 may be further divided into a
third applied portion directed to valve output 704B for injection
into heat source input 608 of heat exchanger 609 via valve 705,
where it is combined with the flow between heat source output 623
of heat exchanger 622 and heat source input 608 of heat exchanger
609, and a third remaining portion provided by valve output 704C
for subsequent communication.
[0145] Finally, valve 706 receives and combines said third
remaining portion of heat stream 601 at valve input 706A with the
flow received from heat source output 610 of heat exchanger 609 via
valve input 706B and provides said combined flow at valve output
706C for communication to engine air inlet 611.
[0146] This configuration provides the ability to direct a precise
portion of the heat energy present in heat stream 601 to each of
the several heat source inputs in the system (615, 621, and 608).
In this manner, each of the ORC heat exchangers may be configured
to operate with the desired quantity of heat energy. Depending on
the specific characteristics of the charge air heat stream 601,
which may vary according to ambient temperature, engine operating
characteristics, and the like, full control and optimization of the
ORC system may be successfully maintained. As described above and
depicted in the drawings, heat exchanger 609 provides only
supplemental cooling and operates independently of the ORC system.
Any excess heat energy present in the charge air heat stream 601
not beneficially utilized by either ORC heat exchanger (616 or 622)
may be safely provided to the supplemental condensing and cooling
system comprising heat exchanger 609, where it may be independently
consumed without adverse effect on ORC operation.
[0147] FIG. 7B depicts the embodiment of FIG. 6B further comprising
the same valves disposed at equivalent positions with respect to
intermediate heat exchangers 603 and 606. The same degree of
flexibility and control is provided, as each of the intermediate
heat exchangers is in exclusive heat transfer communication with
their respective ORC heat exchanger counterparts (603 with 616, and
606 with 622). This embodiment combines the advantages of the
embodiments of FIGS. 6B and 7A, namely the ability to precisely
control the distribution of heat energy from the charge air heat
stream 601 via a series of flow control valves while simultaneously
offering the option of using intermediate heat exchanges for the
reasons given above.
[0148] While the embodiments of FIGS. 7A and 7B permit injection of
precise portions of heat stream 601 at each of the various
component inputs, they require that portions of the heat stream
injected at upstream inputs also flow downstream through all
subsequent components. In some instances, this scheme may not
permit precise determination of the actual system temperatures. In
FIG. 8A, an embodiment comprising an alternate configuration of
valves is depicted wherein the heat source outputs of the ORC heat
exchangers are isolated from the heat source inputs of the
following heat exchangers. Specifically, valve 801 provides a first
applied portion of the heat stream 601 received at valve input 801A
to heat source input 615 of heat exchanger 616 via valve output
801B and a first remaining portion of said heat stream to valve
output 801C. Similarly, valve 802 provides a second applied portion
of the heat stream 601 received at valve input 802A to heat source
input 621 of heat exchanger 622 via valve output 802B and a second
remaining portion of said heat stream to valve output 802C.
[0149] In both cases, heat source inputs 615 and 621 receive their
respective portions of the charge air heat stream 601 directly from
the valves in communication with the source of said heat stream. As
such, the temperature of the portions of the heat stream provided
to inputs 615 and 621 will be at essentially identical
temperatures. In the previously-described embodiment of FIGS. 7A
and 7B, the temperature of the heat stream at heat source input 621
was determined by the temperature of the second portion of heat
stream 601 injected at valve 703, the temperature of the partially
heat-depleted first portion of the heat stream previously provided
to heat source input 615 and output at heat source output 604, and
the relative mass flow rates of each of those two portions as they
were combined at valve 703. Unlike those previously-described
embodiments, heat source output 617 of heat exchanger 616 is not
communicatively coupled to heat source input 621 of heat exchanger
622 in the embodiment of FIG. 8A. Instead, heat source output 617
is communicatively coupled to valve 805, where the output flow is
combined with the output flow from heat source output 623 of heat
exchanger 622. The combined flow from outputs 617 and 623 are then
communicated via valve 806, which is also in flow receiving
communication with valve 803 to receive the second remaining
portion of heat flow 601, and provide the combined flow to the
supplemental condensing and cooling system via heat source input
608 of heat exchanger 609 as previously described. The
configuration of this embodiment is particularly advantageous when
the heat stream portions applied to each of the heat exchangers
require the highest possible temperatures.
[0150] FIG. 8B depicts the embodiment of FIG. 8A further comprising
intermediate heat exchangers. Valve configuration and operation is
identical to the embodiment of FIG. 8A with the added advantages in
certain circumstances realized by the use of intermediate heat
exchangers previously presented.
[0151] In addition to providing precise control over the
portioning, distribution, and communication of heat energy from a
heat stream, the flow control valves are also operative to control
system pressures via restriction of mass flow rates. For example,
with respect to the embodiment of FIG. 8B, the pressure at heat
source input 602 of intermediate heat exchanger 603 is
significantly determined by the settings of both valve outputs 801B
and 801C. For a given setting of valve output 801B and a given heat
stream input mass flow rate at valve input 801A, the mass flow rate
delivered to heat source input 602 of heat exchanger 603 is also
determined by the setting of valve output 801C. Closing valve
output 801C would impose the entire mass flow and pressure at valve
input 801A upon heat source input 602. Likewise, opening valve
output 801C may significantly diminish the mass flow and pressure
at heat source input 602 to the point that insufficient heat energy
would be delivered to the ORC system for continued operation.
Further, restricting the output flow from heat source output 604 at
valve 805 would result in a significant increase in pressure within
heat exchanger 603 that may lead to system failure or catastrophic
failure of one or more individual components. It should be
appreciated that the flow control valves enable control of all
system operating parameters and that the methods associated with
their configuration is an essential novel element of the invention
described herein. As disclosed elsewhere herein with respect to
jacket water distribution subsystem 210, the numerous valves in the
other embodiments are also preferably controlled, and their status
monitored, by a suitable microprocessor-based control subsystem
(not shown) such as the DirectLogic series of programmable logic
controllers (PLCs) available from Automation Direct of Cumming,
Ga.
[0152] While the embodiments presented in FIGS. 6A through 8B have
disclosed a heat stream as that of a compressed high temperature
charge air heat stream, a single ORC system comprising multiple
heat exchangers, and one supplemental condensing and cooling
system, a person of ordinary skill in the art will immediately
recognize and appreciate any number of related embodiments enabled
by this disclosure. For example and not by way of limitation or
exclusion of any other embodiments not specifically mentioned
herein, heat stream 601 may comprise any other suitable source of
heat conveyed via any suitable medium. More than one ORC system may
be provided heat energy from a single heat stream, particularly a
charge air heat stream from the turbocharger associated with a
large industrial internal combustion engine. While the heat
exchangers depicted in said embodiments are disclosed to be
components of one or more ORC systems, portioned heat energy
obtained from a heat stream according to the various configurations
disclosed herein may be applied to heat exchangers operative for
non-ORC purposes in addition to, or in lieu of, use in one or more
ORC systems.
[0153] The embodiment depicted in FIG. 9 depicts a charge air
cooling system suitable for use with a single engine requiring
multiple ORC system to consume sufficient heat energy from the
charge air heat stream to provide a compressed air feed to the
engine at the desired temperature. System 900 comprises a single
engine 901, which in some embodiments may be a two stroke engine
producing in excess of 40 MW output power. Turbocharger 902
comprises a first turbine or other rotating machine 902A driven by
the exhaust gas of the engine mechanically coupled to a second
turbine, compressor, or other rotating machine 902B that receives
ambient intake air and provides a charge air feed with the
necessary degree of compression while increasing the temperature of
the compressed charge air stream to a temperature in excess of that
preferred by the engine. The charge air feed therefore comprises
the heat stream from which heat will be extracted and, to the
greatest degree possible, converted into useful mechanical or
electrical energy.
[0154] The charge air heat stream is fed to intermediate heat
exchanger 903, within which the bulk of the excess heat energy in
the charge air feed is communicated via counterflow to a heat
transfer medium of any suitable type know in the art. Said heat
transfer medium circulates within a closed loop system under
pressure supplied by pump 904. In this embodiment, three separate
but identical ORC systems are depicted operating with controllable
heat stream inputs essentially in a parallel heat energy receiving
configuration, but such a system may comprise any number of ORC
systems necessary or desired to consume up to all of the heat
energy required to be extracted from the charge air heat stream.
For the purpose of the following description, reference will often
be made to elements only by their base element descriptors since
each of the three ORC systems is identically configured. For
example, the three ORC systems each comprise a single expander
bearing a base element descriptor of 908 but individually
identified in FIG. 9 as expander 908A, expander 908B, and expander
908C. Reference to "expanders 908" is intended to apply to each
expander 908A, 908B, and 908C identically without distinction or
difference.
[0155] Pump 904 provides sufficient motive force to provide flow of
the heat transfer medium throughout the system beginning with heat
exchanger 903. After passing through said heat exchanger and
receiving heat energy communicated from the charge air heat stream,
heated heat transfer medium is delivered to flow control valves 905
in heat transfer medium communication with ORC heat exchangers 906
and 907 in sequence. Valves 905 are controllably configured to
accept a portion of the heat transfer medium as desired and
directed by the system operator. The passage of heated transfer
medium through said ORC heat exchangers communicates the heat
energy obtained from the charge air feed to the counterflowing ORC
working fluid in the heat exchangers, which fluid is then
communicated to expanders 908 for expansion. Mechanical energy is
produced by expansion of the ORC working fluid and said energy is
utilized to provide useful work, such as driving at least one of
any of an electric power generator, a prime mover, a pump, a
combustion engine, a fan, a turbine, or a compressor (not shown).
In the case of an electrical power generator, the electric power
may be coupled to a commercial power grid or used independently for
any beneficial purpose, including powering other systems or devices
associated with engine 901 or elsewhere within system 900.
[0156] Following expansion, the ORC working fluid is communicated
to condensers 909 for cooling sufficient to return said working
fluid to a sufficiently liquid state to permit repressurization by
pumps 910.
[0157] Having passed through ORC heat exchangers 906 and 907, the
portions of heat transfer medium allocated by valves 905 to each of
the separate ORC systems is now communicated to condensing and
cooling systems via heat exchangers 913 for removal of any
additional undesired excess heat energy not consumed by ORC heat
exchangers 906 and 907. Heat exchangers 913 function as condensers
and operate in separate closed loop circuits with ORC condensers
909, radiators 914, pumps 915, and a separate secondary heat
transfer medium. As described elsewhere, the supplemental
condensing and cooling systems are preferably configured to consume
any residual heat present in the portion of heated transfer medium
not previously consumed by the ORC system and to provide excess
heat consumption capacity when needed.
[0158] The outputs of heat exchangers are communicated to valves
917 in flow communication with the input of pumps 904. In one
embodiment, valves 917 are volume flow regulator valves configured
to operate based on the pressure drop across the valves.
[0159] Valves 905 and 912 are configured and operative to control
the flow of heat transfer medium through and across heat exchangers
906, 907, and 913. In addition to regulating the flow as desired by
the system operator, they are also configured to provide heat
transfer medium bypass across the heat exchangers. Valves 905 are
operative to bypass all or a portion of the heat transfer medium
across ORC heat exchangers 906 and 907 via bypass lines 911, while
valves 912 are operative to bypass all or a portion of the heat
transfer medium across supplemental heat exchangers 913 via bypass
lines 916. These bypasses provide additional flow control
capability and may also be utilized during system start-up and
shutdown. In addition, valve 918 provides the ability to bypass a
portion or all of the flow of heat transfer medium across all of
the ORC systems when necessary or advisable. In some embodiments,
valve 918 may be a differential pressure regulator such as the
Series 42 self-operated Differential Pressure and Flow Regulators
manufactured and sold by Samson Controls, Inc. Said device may be
configured to prevent transient pressure spikes and to provide a
degree of independence between the combined flow of heat transfer
medium through the ORC systems and the flow through intermediate
heat exchanger 903, particularly when the mass flow required
through intermediate heat exchanger 903 exceeds the flow that may
be accommodated by available ORC systems.
[0160] Following passage through intermediate heat exchanger 903,
the charge air heat stream is communicated to a secondary
condensing and cooling system comprising heat exchanger 920,
radiator 922, and pump 921. As previously disclosed, radiator 922
may comprise any known type of apparatus or device suitable for the
rejection of heat from the charge air heat stream, including but
not limited to forced air or liquid cooled radiators. In some
embodiments, the secondary condensing and cooling system may
comprise a closed loop circuit with a separate heat transfer medium
circulating therein.
[0161] Following passage through the secondary condensing and
cooling system, the charge air feed is now at a temperature
suitable for use by the engine and is communicated thereto.
[0162] FIG. 10 depicts an embodiment comprising more than one
intermediate heat exchanger operating at different temperatures and
configured to transfer heat energy from the charge air feed to
multiple ORC systems for conversion into mechanical or electrical
power. System 1000 comprises a single engine 1001 and associated
turbocharger 1002 comprises a first turbine or other rotating
machine 1002 A driven by the exhaust gas of the engine mechanically
coupled to a second turbine, compressor, or other rotating machine
1002B that receives ambient intake air and provides the charge air
feed as in the previous embodiment.
[0163] The charge air heat stream is fed to high temperature
intermediate heat exchanger 1003, within which the bulk of the
excess heat energy in the charge air feed is communicated via
counterflow to a heat transfer medium of any suitable type know in
the art. Said heat transfer medium circulates within a high
temperature closed loop system under pressure supplied by pump
1004. As before, the depiction of this embodiment comprises three
separate, identical ORC systems but is easily extensible to
comprise any number of ORC system necessary or desired to consume
up to all of the heat energy required to be extracted from the
charge air heat stream as will be exemplified below.
[0164] Pump 1004 provides sufficient motive force to provide flow
of the heat transfer medium throughout the high temperature portion
of the system beginning with heat exchanger 1003. After passing
through said heat exchanger and receiving heat energy communicated
from the charge air heat stream, heated heat transfer medium is
delivered to flow control valves 1005 in heat transfer medium
communication with ORC heat exchangers 1007 and 1008 sequentially.
Valves 1005 are controllably configured to accept a portion of the
heat transfer medium as desired and directed by the system
operator. The passage of heated transfer medium through said ORC
heat exchangers communicates the heat energy obtained from the
charge air feed to the counterflowing ORC working fluid in said
heat exchangers, which fluid is then communicated to expanders 1009
for expansion. Mechanical energy is produced by expansion of the
ORC working fluid and said energy is utilized to provide useful
work, such as driving at least one of any of an electric power
generator, a prime mover, a pump, a combustion engine, a fan, a
turbine, or a compressor (not shown). In the case of an electrical
power generator, the electric power may be coupled to a commercial
power grid or used independently for any beneficial purpose,
including powering other systems or devices associated with engine
1001 or elsewhere within system 1000.
[0165] Following expansion, the ORC working fluid is communicated
to condensers 1010 for cooling sufficient to return said working
fluid to a sufficiently liquid state to permit repressurization by
pumps 1025. The working fluid outputs of condensers 1010 are in
working fluid sending communication with radiators 1026 and pumps
1027 through which a separate heat transfer media may
circulate.
[0166] Having passed through ORC heat exchangers 1007 and 1008, the
portion of heat transfer medium apportioned by valves 1005 now
passes through valves 1013 for communication to the input of pump
1004 for repressurization. In one embodiment, valves 1013 are
volume flow regulator valves configured to operate based on the
pressure drop across the valves.
[0167] Valves 1005 and 1006 are configured and operative to control
the flow of heat transfer medium through and across heat exchangers
1007 and 1008. In addition to regulating the flow as desired by the
system operator, they are also configured to provide heat transfer
medium bypass across the heat exchangers. Valves 1005 are operative
to bypass all or a portion of the heat transfer medium across ORC
heat exchangers 1007 via bypass lines 1011, while valves 1006 are
operative to bypass all or a portion of the heat transfer medium
across ORC heat exchangers 1008 via bypass lines 1012. These
bypasses provide additional flow control capability and may also be
utilized during system start-up and shutdown. In addition, valve
1012 provides the ability to bypass all or a portion of the heat
transfer medium in the high temperature portions across of all of
the ORC systems when necessary or advisable. In some embodiments,
valve 1012 may be a differential pressure regulator such as the
Series 42 self-operated Differential Pressure and Flow Regulators
manufactured and sold by Samson Controls, Inc. Said regulator
valves may be configured to prevent transient pressure spikes and
to provide a degree of independence between the combined flow of
heat transfer medium through the ORC systems and the flow through
intermediate heat exchanger 1003, particularly when the mass flow
required through intermediate heat exchanger 1003 exceeds the flow
that may be accommodated by available ORC systems.
[0168] Following passage through high temperature intermediate heat
exchanger 1003, the charge air heat stream is communicated to low
temperature intermediate heat exchanger 1020. As a substantial
portion of the heat energy has been removed from the charge air
feed during passage through high temperature heat exchanger 1003,
the temperature of the feed is now significantly lower than
initially. Pump 1021 provides sufficient motive force to provide
flow of heat transfer medium throughout the low temperature portion
of the system beginning with heat exchanger 1020. After passing
through said heat exchanger and receiving heat energy communicated
from the charge air heat stream, heated heat transfer medium is
delivered to flow control valves 1022 in heat transfer medium
communication with ORC heat exchangers 1024. Valves 1022 are
controllably configured to accept a portion of the heat transfer
medium as desired and directed by the system operator. The passage
of heated transfer medium through said ORC heat exchangers
communicates the heat energy obtained from the charge air feed to
the counterflowing ORC working fluid in the heat exchangers, which
is then communicated to the inputs of ORC heat exchangers 1008 in
the high temperature portion of the system. In this manner, the low
temperature portion of the system provides pre-heating of the ORC
working fluid prior to communication of said fluid to the high
temperature portion of the system.
[0169] The outputs of heat exchangers 1024 are communicated to
valves 1023 in flow communication with the input of pumps 1021. In
one embodiment, valves 1023 are volume flow regulator valves
configured to operate based on the pressure drop across the
valves.
[0170] Valves 1022 are configured and operative to control the flow
of heat transfer medium through and across heat exchangers 1024. In
addition to regulating the flow as desired by the system operator,
they are also configured to provide heat transfer medium bypass
across said heat exchangers. Valves 1022 are operative to bypass
all or a portion of the heat transfer medium across ORC heat
exchangers 1024 via bypass lines 1028. These bypasses provide
additional flow control capability and may also be utilized during
system start-up and shutdown.
[0171] In addition, valve 1029 provides the ability to bypass all
or a portion of the heat transfer medium in the low temperature
portion of the system across of all of the ORC systems when
necessary or advisable. In some embodiments, valve 1029 may be a
differential pressure regulator such as the Series 42 self-operated
Differential Pressure and Flow Regulators manufactured and sold by
Samson Controls, Inc. Said regulator valve may be configured to
prevent transient pressure spikes and to provide a degree of
independence between the combined flow of heat transfer medium
through the ORC systems and the flow through low temperature
intermediate heat exchanger 1020, particularly when the mass flow
required through said heat exchanger exceeds the flow that may be
accommodated by available ORC systems.
[0172] Following passage through low temperature intermediate heat
exchanger 1020, the charge air heat stream is communicated to a
secondary condensing and cooling system comprising heat exchanger
1030, radiator 1033, and pump 1032. As previously disclosed,
radiator 1033 may comprise any know type of apparatus or device
suitable for the rejection of heat from the charge air heat stream,
including but not limited to forced air or liquid cooled radiators.
In some embodiments, the secondary condensing and cooling system
may comprise a closed loop circuit with a separate heat transfer
medium circulating therein.
[0173] Following passage through the secondary condensing and
cooling system, the charge air feed is now at a temperature
suitable for use by the engine and is communicated thereto.
[0174] In one non-limiting exemplary embodiment of the system
depicted in FIG. 10, the charge air feed associated with a
turbocharged MAN K80MC-S9 engine comprises the heat stream provided
by turbocharger 1002 B to ORC heat exchangers 1003 and 1020 and
secondary condensing and cooling system heat exchanger 1030
sequentially. Given an ambient inlet air temperature of 37.degree.
C., the turbocharger will produce a charge air feed with an
approximate mass flow rate of 100 kg/s at a approximate temperature
of 210.degree. C. The desired engine air inlet temperature is
specified by the manufacturer to be 44.degree. C., requiring a
reduction in temperature of 166.degree. C. by extracting heat
energy at the approximate rate of 17 MW (17 MJ/s) from the heat
stream. Extracting energy from the charge air heat stream via the
embodiment of FIG. 10 will provide significant mechanical power to
perform useful work or to generate electrical power.
[0175] As described above, a first portion of heat energy is
extracted via high temperature heat exchanger 1003 for
communication to the working fluid of multiple ORC systems via dual
heat exchangers 1007 and 1008. Although two heat exchangers are
depicted, it should be appreciated that one heat exchanger or more
than two heat exchangers may be utilized to accomplish the transfer
of heat energy from the heat transfer medium circulating in the
high temperature portion of the system. In this exemplary
embodiment, said heat transfer medium is water.
[0176] The temperature at the charge air heat stream input to heat
exchanger 1003 is approximately 210.degree. C., and after
extracting a portion of the heat energy at the approximate rate of
8.3 MW and communicating it to the heat transfer medium therein, it
exits said heat exchanger at an approximate temperature of
128.degree. C. The portion of extracted heat energy communicated to
the water heat transfer medium raises the temperature of the
circulating water to approximately 122.degree. C. A portion of the
heat energy in said heat transfer medium is then communicated to
the ORC working fluid in each of the multiple ORC systems via heat
exchangers 1007 and 1008.
[0177] The partially heat-depleted charge air heat stream is
communicated to low temperature intermediate heat exchanger 1020 at
an approximate input temperature of 128.degree. C., where
additional heat energy is extracted at the approximate rate of 3.7
MW prior to expulsion of the further heat depleted charge air heat
stream at an approximate temperature of 93.degree. C. The
temperature of the heat transfer medium flowing in the low
temperature portion of the system is raised to an approximate
temperature of 87.degree. C. and applied to ORC heat exchangers
1024 to preheat the ORC working fluid prior to its communication to
ORC heat exchangers 1008 and 1007.
[0178] Finally, the charge air heat stream is communicated to
secondary condensing and cooling system heat exchanger 1030 at an
approximate input temperature of 93.degree. C., wherein additional
heat energy is extracted at the rate of approximately 5.0 MW and
dissipated by radiator 1033. It should be appreciated that heat
exchanger 1030 and radiator 1033 may each comprise more than one
physical apparatus of identical or different types. When expelled
from heat exchanger 1030, the charge air stream is now
approximately at the required temperature of 44.degree. C. for
injection into engine 1001, having transferred heat energy at the
approximate total rate of 17 MW to the three charge air heat
exchangers described above.
[0179] The heat energy transferred to the multiple ORC systems via
the high temperature and low temperature portions of the system is
used to heat ORC working fluid for expansion in expanders 1009,
thereby generating mechanical power. Given the temperature of the
water heat transfer medium flowing in the each of the high and low
temperature portions of the system, a single ORC system of the type
manufactured and sold by ElectraTherm, Inc. of Reno, Nev. may be
expected to consume heat energy at the approximate combined rate of
1.25 MW. Given the total rate of heat energy consumption of the
charge air heat exchangers 1003 and 1020 of approximately 12.0 MW,
it can be seen that a total of ten identical ORC systems would be
required for full consumption of all charge air heat in this
embodiment. As previously described, the system depicted in FIG. 10
is easily extensible to any number of ORC systems, including the
ten required in this example, from the three shown for purposes of
enablement of the written description.
[0180] If the mechanical power developed via expansion of each of
the expanders 1009 is coupled to electrical generators, each system
would provide approximately 93 kW (0.093 MW) of electrical power
for a combined electrical power output of 930 kW (0.93 MW). As
such, the overall efficiency of the combined charge air cooling and
heat energy recovery system is approximately 7.8%. In practice,
additional heat losses not considered in the above example would
reduce the realized efficiency slightly, but the fact that close to
one megawatt of power may be generated from a charge air stream,
previously classified as a waste heat byproduct of the necessary
charge air cooling, represents a significant advantage over known
technologies.
[0181] The description of this invention is intended to be enabling
and not limiting. It will be evident to those skilled in the art
that numerous combinations of the embodiments described above may
be implemented together as well as separately, and all such
combinations constitute embodiments effectively described
herein.
* * * * *