Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
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REVERSIBLE TURBOMACHINES IN PUMPED HEAT ENERGY STORAGE
SYSTEMS
BACKGROUND
[0001] In a heat engine or heat pump, a heat exchanger may be employed to
transfer heat
between a thermal storage material and a working fluid for use with
turbomachinery. The heat
engine may be reversible, e.g., it may also be a heat pump, and the working
fluid and heat
exchanger may be used to transfer heat or cold to thermal storage media.
SUMMARY
[0002] A Pumped Heat Energy Storage ("PHES") system may include at least a
working fluid
circulated through a closed cycle fluid path including at least two heat
exchangers, at least one
turbine, and at least one compressor. In some systems, one or more
recuperative heat
exchangers may also be included. At least two temperature reservoirs may hold
thermal fluids
which may be sent through the heat exchangers, providing thermal energy to,
and/or extracting
thermal energy from, the working fluid. One or more motor/generators may be
used to obtain
work from the thermal energy in the system, preferably by generating
electricity from
mechanical energy received from the turbine.
[0003] In one aspect, a system includes PHES system that includes a first
reversible
turbomachine. a second reversible turbomachine, a hot-side heat exchanger, and
a cold-side
heat exchanger. The PHES system is configured to operate in a charge mode and
a discharge
mode. During charge mode, the first reversible turbomachine acts as a
compressor, the second
reversible turbomachine acts as a turbine, and a working fluid circulates
through, in sequence,
the first reversible turbomachine. the hot-side heat exchanger, the second
reversible
turbomachine, the cold-side heat exchanger, and back to the first reversible
turbomachine.
During discharge mode, the first reversible turbomachine acts as a turbine,
the second
reversible turbomachine acts as a compressor, and the working fluid circulates
through, in
sequence, the first reversible turbomachine, the cold-side heat exchanger, the
second reversible
turbomachine. the hot-side heat exchanger, and back to the first reversible
turbomachine.
[0004] In another aspect, a system includes PHES system that includes a first
reversible
turbomachine. a second reversible turbomachine, a hot-side heat exchanger, a
cold-side heat
exchanger, and a recuperator heat exchanger. The PHES system is configured to
operate in a
charge mode and a discharge mode. During charge mode, the first reversible
turbomachine
acts as a compressor, the second reversible turbomachine acts as a turbine,
and a working fluid
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circulates through, in sequence, the first reversible turbomachine, the hot-
side heat exchanger,
the recuperator heat exchanger, the second reversible turbomachine, the cold-
side heat
exchanger, the recuperator heat exchanger, and back to the first reversible
turbomachine.
During discharge mode, the first reversible turbomachine acts as a turbine,
the second
reversible turbomachine acts as a compressor, and the working fluid circulates
through, in
sequence, the first reversible turbomachine, the recuperator heat exchanger,
the cold-side heat
exchanger, the second reversible turbomachine, the recuperator heat exchanger,
the hot-side
heat exchanger, and back to the first reversible turbomachine.
[0005] In another aspect, a method includes operating a first reversible
turbomachine as a
compressor and circulating a working fluid from the first reversible
turbomachine to, in
sequence, a hot-side heat exchanger, a second turbomachine acting as a
turbine, a cold-side
heat exchanger, and back to the first reversible turbomachine. The method
further includes
operating the first reversible turbomachine as a turbine and circulating the
working fluid from
the first reversible turbomachine to, in sequence, the cold-side heat
exchanger, a third
turbornachine acting as a compressor, the hot-side heat exchanger, and back to
the first
reversible turbomachine.
BRIEF DESCRIPTION OF THE DRAWINGS
[0006] FIG. 1 schematically illustrates operation of a pumped thermal electric
storage system.
[0007] FIG. 2 is a schematic flow diagram of working fluid and heat storage
media of a pumped
thermal system in a charge/heat pump mode.
[0008] FIG. 3 is a schematic flow diagram of working fluid and heat storage
media of a pumped
thermal system in a discharge/heat engine mode.
[0009] FIG. 4 is a schematic pressure and temperature diagram of the working
fluid as it
undergoes the charge cycle in FIG. 2.
[0010] FIG. 5 is a schematic pressure and temperature diagram of the working
fluid as it
undergoes the discharge cycle in FIG. 3.
[0011] FIG. 6 is a schematic perspective view of a closed working fluid system
in the pumped
thermal system in FIGs. 2-3.
[0012] FIG. 7 is a schematic perspective view of the pumped thermal system in
FIGs. 2-3 with
hot side and cold side storage tanks and a closed cycle working fluid system.
[0013] FIG. 8 shows a heat storage charge cycle for a water/molten salt system
with ric = 0.9
and it = 0.95. The dashed lines correspond to Tic = it = 1.
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[0014] FIG. 9 shows a heat storage discharge (extraction) cycle for the
water/molten salt
system in FIG. 8 with Tic = 0.9 and lit = 0.95. The dashed lines correspond to
tic = it = 1.
[0015] FIG. 10 shows a heat storage cycle in a pumped thermal system with
variable
compression ratios between the charge and discharge cycles.
[0016] FIG. 11 shows roundtrip efficiency contours for a water/salt system.
The symbols e
and 0 represent an approximate range of present large turbomachinery adiabatic
efficiency
values. The dashed arrows represent the direction of increasing efficiency.
[0017] FIG. 12 shows roundtrip efficiency contours for a colder storage/salt
system. The
symbols G and 0 represent an approximate range of present large turbomachinery
adiabatic
efficiency values.
[0018] FIG. 13 is a schematic flow diagram of working fluid and heat storage
media of a
pumped thermal system with a gas-gas heat exchanger for the working fluid in a
charge/heat
pump mode.
[0019] FIG. 14 is a schematic flow diagram of working fluid and heat storage
media of a
pumped thermal system with a gas-gas heat exchanger for the working fluid in a
discharge/heat
engine mode.
[0020] FIG. 15 is a schematic flow diagram of working fluid and heat storage
media of a
pumped thermal system with a gas-gas heat exchanger for the working fluid in a
charge/heat
pump mode with indirect heat rejection to the environment.
[0021] FIG. 16 is a schematic flow diagram of working fluid and heat storage
media of a
pumped thermal system with a gas-gas heat exchanger for the working fluid in a
discharge/heat
engine mode with indirect heat rejection to the environment.
[0022] FIG. 17 shows a heat storage charge cycle for a storage system with a
gas-gas heat
exchanger, a cold side storage medium capable of going down to temperatures
significantly
below ambient temperature and tic = 0.9 and -qt. = 0.95.
[0023] FIG. 18 shows a heat storage discharge cycle for a storage system with
a gas-gas heat
exchanger, a cold side storage medium capable of going down to temperatures
significantly
below ambient temperature and tic = 0.9 and it = 0.95.
[0024] FIG. 19 is a schematic flow diagram of hot side recharging in a pumped
heat cycle in
solar mode with heating of a solar salt solely by solar power.
[0025] FIG. 20 is a schematic flow diagram of a pumped thermal system
discharge cycle with
heat rejection to ambient.
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[0026] FIG. 21 is a schematic flow diagram of a pumped thermal system
discharge cycle with
heat rejection to an intermediate fluid circulated in a thermal bath at
ambient temperature.
[0027] FIGs. 22 and 23 are pumped thermal systems with separate
compressor/turbine pairs
for charge and discharge modes.
[0028] FIGs. 24 and 25 show pumped thermal systems configured in a combustion
heat input
generation mode.
[0029] FIG. 26 is a schematic flow diagram of hot side recharging in a pumped
heat cycle
through heating by a combustion heat source or a waste heat source.
[0030] FIG. 27 shows an example of a pumped thermal system with pressure
regulated power
control.
[0031] FIG. 28 shows an example of a pumped thermal system with a pressure
encased
generator.
[0032] FIG. 29 is an example of variable stators in a compressor/turbine pair.
[0033] FIG. 30 shows a computer system that is programmed to implement various
methods
and/or regulate various systems of the present disclosure.
[0034] FIGs. 31A and 31B are schematic flow diagrams of working fluid and heat
storage
media of a pumped thermal system with reversible turbomachinery in charge and
discharge
modes.
[0035] FIGs. 32A and 32B are schematic flow diagrams of working fluid and heat
storage
media of a pumped thermal system with reversible turbomachinery in charge and
discharge
modes.
[0036] FIG. 33 is a schematic diagram of a combined turbomachine drivetrain.
[0037] FIG. 34 is a schematic diagram of individual turbomachine diivetrains.
[0038] FIG. 35 is a representation of a pair of reversible turbomachines,
according to an
example embodiment.
[0039] FIGs. 36-39 are illustrations of reversible turbomachine blades
compared to
conventional turbine and compressor blades, according to example embodiments.
[0040] FIG. 40 is a representation of a compressor or turbine stage, viewed
axially, according
to an example embodiment.
[0041] FIG. 42 is an illustration of an example blade arc, according to an
example embodiment.
[0042] FIGs. 43-45 are illustrations of blade shapes based on the blade arc of
FIG. 42,
according to example embodiments.
[0043] FIGs. 46-49 are illustrations of reversible turbomachine blades
compared to
conventional turbine and compressor blades, according to example embodiments.
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DETAILED DESCRIPTION
[0044] Example systems and methods are described herein. It should be
understood that the
words "example- and/or "exemplary" are used herein to mean "serving as an
example,
instance, or illustration." Any embodiment or feature described herein as
being an "example"
or "exemplary" is not necessarily to be construed as preferred or advantageous
over other
embodiments or features. The example embodiments described herein are not
meant to be
limiting. It will be readily understood that certain aspects of the disclosed
systems and methods
can be arranged and combined in a wide variety of different configurations,
all of which are
contemplated herein.
[0045] While various embodiments of the invention have been shown and
described herein, it
will be obvious to those skilled in the art that such embodiments are provided
by way of
example only. Numerous variations, changes, and substitutions may occur to
those skilled in
the art without departing from the invention. It should be understood that
various alternatives
to the embodiments of the invention described herein may be employed. It shall
be understood
that different aspects of the invention can be appreciated individually,
collectively, or in
combination with each other.
[0046] It is to be understood that the terminology used herein is used for the
purpose of
describing specific embodiments, and is not intended to limit the scope of the
present invention.
It should be noted that as used herein, the singular forms of "a", "an" and
"the" include plural
references unless the context clearly dictates otherwise. In addition, unless
defined otherwise,
all technical and scientific terms used herein have the same meaning as
commonly understood
by one of ordinary skill in the art to which this invention belongs.
[0047] While preferable embodiments of the present invention are shown and
described herein,
it will be obvious to those skilled in the art that such embodiments are
provided by way of
example only. Numerous variations, changes, and substitutions will now occur
to those skilled
in the art without departing from the invention. It should be understood that
various alternatives
to the embodiments of the invention described herein may be employed in
practicing the
invention. It is intended that the following claims define the scope of the
invention and that
methods and structures within the scope of these claims and their equivalents
be covered
thereby.
I. Overview
[0048] A closed thermodynamic cycle power generation or energy storage system,
such as a
reversible Brayton cycle system, may use a generator/motor connected to a
turbine and a
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compressor which act on a working fluid circulating in the system. Examples of
working fluids
include air, argon, carbon dioxide, or gaseous mixtures. A closed
thermodynamic cycle power
generation or energy storage system, such as a reversible Brayton cycle
system, may have a
hot side and/or a cold side. Each side may include a heat exchanger coupled to
one or more
cold storage containers and/or one or more hot storage containers. Preferably,
the heat
exchangers may be arranged as counterflow heat exchangers for higher thermal
efficiency.
Liquid thermal storage medium be utilized and may include, for example,
liquids that are stable
at high temperatures, such as molten nitrate salt or solar salt, or liquids
that are stable at low
temperatures, such as glycols or alkanes such as hexane. For an example in a
molten salt and
hexane system, the hot side molten salt may include a hot storage at
approximately 565 C and
a cold storage at approximately 290 C and the cold side hexane may include a
hot storage at
approximately 35 C and a cold storage at approximately -60 C.
[0049] Large power generation systems may be slow to ramp up to full power
generation. In
addition, out-of-phase transient spikes in generated power in such large
systems may be
disruptive. Instead, it may be desirable to utilize a plurality of power
subunits, each power
subunit generating a portion of the total maximum power. Each power subunit
may include a
valve arrangement configurable to be in a connected state or an isolated state
with respect to a
shared hot side thermal store and a shared cold side thermal store (i.e., to
be in use or not in
use).
Illustrative Reversible Heat Engine
[0050] While various embodiments of the invention have been shown and
described herein, it
will be obvious to those skilled in the art that such embodiments are provided
by way of
example only. Numerous variations, changes, and substitutions may occur to
those skilled in
the art without departing from the invention. It should be understood that
various alternatives
to the embodiments of the invention described herein may be employed. It shall
be understood
that different aspects of the invention can be appreciated individually,
collectively, or in
combination with each other.
[0051] It is to be understood that the terminology used herein is used for the
purpose of
describing specific embodiments, and is not intended to limit the scope of the
present invention.
It should be noted that as used herein, the singular forms of "a", "an" and
"the" include plural
references unless the context clearly dictates otherwise. In addition, unless
defined otherwise,
all technical and scientific terms used herein have the same meaning as
commonly understood
by one of ordinary skill in the art to which this invention belongs.
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[0052] While preferable embodiments of the present invention are shown and
described herein,
it will be obvious to those skilled in the art that such embodiments are
provided by way of
example only. Numerous variations, changes, and substitutions will now occur
to those skilled
in the art without departing from the invention. It should be understood that
various alternatives
to the embodiments of the invention described herein may be employed in
practicing the
invention. It is intended that the following claims define the scope of the
invention and that
methods and structures within the scope of these claims and their equivalents
be covered
thereby.
[0053] The term "reversible," as used herein, generally refers to a process or
operation that can
be reversed via infinitesimal changes in some property of the process or
operation without
substantial entropy production (e.g., dissipation of energy). A reversible
process may be
approximated by a process that is at thermodynamic equilibrium. In some
examples, in a
reversible process, the direction of flow of energy is reversible. As an
alternative, or in addition
to, the general direction of operation of a reversible process (e.g., the
direction of fluid flow)
can be reversed, such as, e.g., from clockwise to counterclockwise, and vice
versa.
[0054] The term "sequence," as used herein, generally refers to elements
(e.g., unit operations)
in order. Such order can refer to process order, such as, for example, the
order in which a fluid
flows from one element to another. In an example, a compressor, heat storage
unit and turbine
in sequence includes the compressor upstream of the heat exchange unit, and
the heat exchange
unit upstream of the turbine. In such a case, a fluid can flow from the
compressor to the heat
exchange unit and from the heat exchange unit to the turbine. A fluid flowing
through unit
operations in sequence can flow through the unit operations sequentially. A
sequence of
elements can include one or more intervening elements. For example, a system
comprising a
compressor, heat storage unit and turbine in sequence can include an auxiliary
tank between
the compressor and the heat storage unit. A sequence of elements can be
cyclical.
Pumped thermal systems
[0055] The disclosure provides pumped thermal systems capable of storing
electrical energy
and/or heat, and releasing energy (e.g., producing electricity) at a later
time. The pumped
thermal systems of the disclosure may include a heat engine, and a heat pump
(or refrigerator).
In some cases, the heat engine can be operated in reverse as a heat pump. In
some cases, the
heat engine can be operated in reverse as a refrigerator. Any description of
heat pump/heat
engine systems or refrigerator/heat engine systems capable of reverse
operation herein may
also be applied to systems comprising separate and/or a combination of
separate and reverse-
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operable heat engine system(s), heat pump system(s) and/or refrigerator
system(s). Further, as
heat pumps and refrigerators share the same operating principles (albeit with
differing
objectives), any description of configurations or operation of heat pumps
herein may also be
applied to configurations or operation of refrigerators, and vice versa.
[0056] Systems of the present disclosure can operate as heat engines or heat
pumps (or
refrigerators). In some situations, systems of thc disclosure can alternately
operate as heat
engines and heat pumps. In some examples, a system can operate as a heat
engine to generate
power, and subsequently operate as a heat pump to store energy, or vice versa.
Such systems
can alternately and sequentially operate as heat engines as heat pumps. In
some cases, such
systems reversibly or substantially reversibly operate as heat engines as heat
pumps.
[0057] Reference will now be made to the figures, wherein like numerals refer
to like parts
throughout. It will be appreciated that the figures and features therein are
not necessarily drawn
to scale.
[0058] FIG. I schematically illustrates operating principles of pumped thermal
electric storage
using a heat pump/heat engine electricity storage system. Electricity may be
stored in the form
of thermal energy of two materials or media at different temperatures (e.g.,
thermal energy
reservoirs comprising heat storage fluids or thermal storage media) by using a
combined heat
pump/heat engine system. In a charging or heat pump mode, work may be consumed
by the
system for transferring heat from a cold material or medium to a hot material
or medium, thus
lowering the temperature (e.g., sensible energy) of the cold material and
increasing the
temperature (i.e., sensible energy) of the hot material. In a discharging or
heat engine mode,
work may be produced by the system by transferring heat from the hot material
to the cold
material, thus lowering the temperature (i.e., sensible energy) of the hot
material and increasing
the temperature (i.e., sensible energy) of the cold material. The system may
be configured to
ensure that the work produced by the system on discharge is a favorable
fraction of the energy
consumed on charge. The system may be configured to achieve high roundtrip
efficiency,
defined herein as the work produced by the system on discharge divided by the
work consumed
by the system on charge. Further, the system may be configured to achieve the
high roundtrip
efficiency using components of a desired (e.g., acceptably low) cost. Arrows H
and W in FIG.
I represent directions of heat flow and work, respectively.
[0059] Heat engines, heat pumps and refrigerators of the disclosure may
involve a working
fluid to and from which heat is transferred while undergoing a thermodynatnic
cycle. The heat
engines, heat pumps and refrigerators of the disclosure may operate in a
closed cycle. Closed
cycles allow, for example, a broader selection of working fluids, operation at
elevated cold side
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pressures, operation at lower cold side temperatures, improved efficiency, and
reduced risk of
turbine damage. One or more aspects of the disclosure described in relation to
systems having
working fluids undergoing closed cycles may also be applied to systems having
working fluids
undergoing open cycles.
[0060] In one example, the heat engines may operate on a Brayton cycle and the
heat
pumps/refrigerators may operate on a reverse Brayton cycle (also known as a
gas refrigeration
cycle). Other examples of thermodynamic cycles that the working fluid may
undergo or
approximate include the Rankine cycle, the ideal vapor-compression
refrigeration cycle, the
Stirling cycle, the Ericsson cycle or any other cycle advantageously employed
in concert with
heat exchange with heat storage fluids of the disclosure.
[0061] The working fluid can undergo a thermodynamic cycle operating at one,
two or more
pressure levels. For example, the working fluid may operate in a closed cycle
between a low
pressure limit on a cold side of the system and a high pressure limit on a hot
side of the system.
In some implementations, a low pressure limit of about 10 atmospheres (atm) or
greater can be
used. In some instances, the low pressure limit may be at least about 1 atm,
at least about 2
atm, at least about 5 atm, at least about 10 atm. at least about 15 atm, at
least about 20 atm, at
least about 30 atm, at least about 40 atm, at least about 60 atm, at least
about 80 atm, at least
about 100 atm, at least about 120 atm, at least about 160 atm, or at least
about 200 atm, 500
atm, 1000 atm, or more. In some instances, a sub-atmospheric low pressure
limit may be used.
For example, the low pressure limit may be less than about 0.1 atm, less than
about 0.2 atm,
less than about 0.5 atm, or less than about 1 atm. In some instances, the low
pressure limit may
be about 1 atmosphere (atm). In the case of a working fluid operating in an
open cycle, the low
pressure limit may be about 1 atm or equal to ambient pressure.
[0062] In some cases, the value of the low pressure limit may be selected
based on desired
power output and/or power input requirements of the thermodynamic cycle. For
example, a
pumped thermal system with a low pressure limit of about 10 atm may be able to
provide a
power output comparable to an industrial gas turbine with ambient (1 atm) air
intake. The value
of the low pressure limit may also be subject to cost/safety tradeoffs.
Further, the value of the
low pressure limit may be limited by the value of the high pressure limit, the
operating ranges
of the hot side and cold side heat storage media (e.g., pressure and
temperature ranges over
which the heat storage media are stable), pressure ratios and operating
conditions (e.g.,
operating limits, optimal operating conditions, pressure drop) achievable by
turbomachinery
and/or other system components, or any combination thereof. The high pressure
limit may be
determined in accordance with these system constraints. In some instances,
higher values of
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the high pressure limit may lead to improved heat transfer between the working
fluid and the
hot side storage medium.
[0063] Working fluids used in pumped thermal systems may include air, argon,
other noble
gases, carbon dioxide, hydrogen, oxygen, or any combination thereof, and/or
other fluids in
gaseous, liquid, critical, or supercritical state (e.g., supercritical CO,,).
The working fluid can
be a gas or a low viscosity liquid (e.g., viscosity below about 500x10-6 Poise
at 1 atm),
satisfying the requirement that the flow be continual. In some
implementations, a gas with a
high specific heat ratio may be used to achieve higher cycle efficiency than a
gas with a low
specific heat ratio. For example, argon (e.g., specific heat ratio of about
1.66) may be used to
substitute air (e.g., specific heat ratio of about 1.4). In some cases, the
working fluid may be a
blend of one, two, three or more fluids. In one example, helium (having a high
thermal
conductivity and a high specific heat) may be added to the working fluid
(e.g., argon) to
improve heat transfer rates in heat exchangers.
[0064] Pumped thermal systems herein may utilize heat storage media or
materials, such as
one or more heat storage fluids. The heat storage media can be gases or low
viscosity liquids,
satisfying the requirement that the flow be continual. The systems may utilize
a first heat
storage medium on a hot side of the system ("hot side thermal storage (HTS)
medium" or
"HTS" herein) and a second heat storage medium on a cold side of the system
("cold side
thermal storage (CTS) medium" or "CTS" herein). The thermal storage media
(e.g., low
viscosity liquids) can have high heat capacities per unit volume (e.g., heat
capacities above
about 1400 Joule (kilogram Kelvin)-') and high thermal conductivities (e.g.,
thermal
conductivities above about 0.7 Watt (meter Kelvin)-1). In some
implementations, several
different thermal storage media (also "heat storage media" herein) on either
the hot side, cold
side or both the hot side and the cold side may be used.
[0065] The operating temperatures of the hot side thermal storage medium can
be in the liquid
range of the hot side thermal storage medium, and the operating temperatures
of the cold side
thermal storage medium can be in the liquid range of the cold side thermal
storage medium. In
some examples, liquids may enable a more rapid exchange of large amounts of
heat by
convective counter-flow than solids or gases. Thus, in some cases, liquid HTS
and CTS media
may advantageously be used. Pumped thermal systems utilizing thermal storage
media herein
may advantageously provide a safe, non-toxic and geography-independent energy
(e.g.,
electricity) storage alternative.
[0066] In some implementations, the hot side thermal storage medium can be a
molten salt or
a mixture of molten salts. Any salt or salt mixture that is liquid over the
operating temperature
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range of the hot side thermal storage medium may be employed. Molten salts can
provide
numerous advantages as thermal energy storage media, such as low vapor
pressure, lack of
toxicity, chemical stability, low chemical reactivity with typical steels
(e.g., melting point
below the creep temperature of steels, low corrosiveness, low capacity to
dissolve iron and
nickel), and low cost. In one example, the HTS is a mixture of sodium nitrate
and potassium
nitrate. In some examples, the HTS is a eutectic mixture of sodium nitrate and
potassium
nitrate. In some examples, the HTS is a mixture of sodium nitrate and
potassium nitrate having
a lowered melting point than the individual constituents, an increased boiling
point than the
individual constituents, or a combination thereof. Other examples include
potassium nitrate,
calcium nitrate, sodium nitrate, sodium nitrite, lithium nitrate, mineral oil,
or any combination
thereof. Further examples include any gaseous (including compressed gases),
liquid or solid
media (e.g., powdered solids) having suitable (e.g., high) thermal storage
capacities and/or
capable of achieving suitable (e.g., high) heat transfer rates with the
working fluid. For
example, a mix of 60% sodium nitrate and 40% potassium nitrate (also referred
to as a solar
salt in some situations) can have a heat capacity of approximately 1500 Joule
(Kelvin mole)-'
and a thermal conductivity of approximately 0.75 Watt (meter Kelvin)-1 within
a temperature
range of interest. The hot side thermal storage medium may be operated in a
temperature range
that structural steels can handle.
[0067] In some cases, liquid water at temperatures of about 0 C to 100 C
(about 273 K-373 K)
and a pressure of about 1 atm may be used as the cold side thermal storage
medium. Due to a
possible explosion hazard associated with presence of steam at or near the
boiling point of
water, the operating temperature can be kept below about 100 C or less while
maintaining an
operating pressure of 1 atm (i.e., no pressurization). In some cases, the
temperature operating
range of the cold side thermal storage medium may be extended (e.g., to -30 C
to 100 C at 1
atm) by using a mixture of water and one or more antifreeze compounds (e.g.,
ethylene glycol,
propylene glycol, or glycerol).
[0068] As described in greater detail elsewhere herein, improved storage
efficiency may be
achieved by increasing the temperature difference at which the system
operates, for example,
by using a cold side heat storage fluid capable of operating at lower
temperatures. In some
examples, the cold side thermal storage media may comprise hydrocarbons, such
as, for
example, alkanes (e.g., hexane or heptane), alkenes, alkynes, aldehydes,
ketones, carboxylic
acids (e.g., HCOOH), ethers, cycloalkanes, aromatic hydrocarbons, alcohols
(e.g., butanol),
other type(s) of hydrocarbon molecules, or any combinations thereof. In some
cases, the cold
side thermal storage medium can be hexane (e.g., n-hexane). Hexane has a wide
liquid range
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and can remain fluid (i.e., runny) over its entire liquid range (-94 C to 68 C
at 1 atm). Hexane's
low temperature properties are aided by its immiscibility with water. Other
liquids, such as, for
example, ethanol or methanol can become viscous at the low temperature ends of
their liquid
ranges due to pre-crystallization of water absorbed from air. In some cases,
the cold side
thermal storage medium can be heptane (e.g., n-heptane). Heptane has a wide
liquid range and
can remain fluid (i.e., runny) over its entire liquid range (-91 C to 98 C at
1 atm). fleptane's
low temperature properties are aided by its immiscibility with water. At even
lower
temperatures, other heat storage media can be used, such as, for example,
isohexane (2-
methylpentane). In some examples, cryogenic liquids having boiling points
below about -
150 C (123 K) or about -180 C (93.15 K) may be used as cold side thermal
storage media
(e.g., propane, butane, pentane, nitrogen, helium, neon, argon and krypton,
air, hydrogen,
methane, or liquefied natural gas). In some implementations, choice of cold
side thermal
storage medium may be limited by the choice of working fluid. For example,
when a gaseous
working fluid is used, a liquid cold side thermal storage medium having a
liquid temperature
range at least partially or substantially above the boiling point of the
working fluid may be
required.
[0069] In some cases, the operating temperature range of CTS and/or HTS media
can be
changed by pressurizing (i.e., raising the pressure) or evacuating (i.e.,
lowering the pressure)
the tanks and thus changing the temperature at which the storage media undergo
phase
transitions (e.g., going from liquid to solid, or from liquid to gas).
[0070] In some cases, the hot side and the cold side heat storage fluids of
the pumped thermal
systems are in a liquid state over at least a portion of the operating
temperature range of the
energy storage device. The hot side heat storage fluid may be liquid within a
given range of
temperatures. Similarly, the cold side heat storage fluid may be liquid within
a given range of
temperatures. The heat storage fluids may be heated, cooled or maintained to
achieve a suitable
operating temperature prior to, during or after operation.
[0071] Pumped thermal systems of the disclosure may cycle between charged and
discharged
modes. In some examples, the pumped thermal systems can he fully charged,
partially charged
or partially discharged, or fully discharged. In some cases, cold side heat
storage may be
charged (also "recharged" herein) independently from hot side heat storage.
Further, in some
implementations, charging (or some portion thereof) and discharging (or some
portion thereof)
can occur simultaneously. For example, a first portion of a hot side heat
storage may be
recharged while a second portion of the hot side heat storage together with a
cold side heat
storage are being discharged.
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[0072] The pumped thermal systems may be capable of storing energy for a given
amount of
time. In some cases, a given amount of energy may be stored for at least about
1 second, at
least about 30 seconds, at least about 1 minute, at least about 5 minutes, at
least about 30
minutes, at least about 1 hour, at least about 2 hours, at least about 3
hours, at least about 4
hours, at least about 5 hours, at least about 6 hours, at least about 7 hours,
at least about S hours,
at least about 9 hours, at least about 10 hours, at least about 12 hours at
least about 14 hours,
at least about 16 hours, at least about 18 hours, at least about 20 hours, at
least about 22 hours,
at least about 24 hours (1 day), at least about 2 days, at least about 4 days,
at least about 6 days,
at least about 8 days, at least about 10 days, 20 days. 30 days, 60 days, 100
days. 1 year or
more.
[0073] Pumped thermal systems of the disclosure may be capable of
storing/receiving input of,
and/or extracting/providing output of a substantially large amount of energy
and/or power for
use with power generation systems (e.g., intermittent power generation systems
such as wind
power or solar power), power distribution systems (e.g. electrical grid),
and/or other loads or
uses in grid-scale or stand-alone settings. During a charge mode of a pumped
thermal system,
electric power received from an external power source (e.g., a wind power
system, a solar
photovoltaic power system, an electrical grid etc.) can be used operate the
pumped thermal
system in a heat pump mode (i.e., transferring heat from a low temperature
reservoir to a high
temperature reservoir, thus storing energy). During a discharge mode of the
pumped thermal
system, the system can supply electric power to an external power system or
load (e.g., one or
more electrical grids connected to one or more loads, a load, such as a
factory or a power-
intensive process, etc.) by operating in a heat engine mode (i.e.,
transferring heat from a high
temperature reservoir to a low temperature reservoir, thus extracting energy).
As described
elsewhere herein, during charge and/or discharge, the system may receive or
reject thermal
power, including, but not limited to electromagnetic power (e.g., solar
radiation) and thermal
power (e.g., sensible energy from a medium heated by solar radiation, heat of
combustion etc.).
[0074] In some implementations, the pumped thermal systems are grid-
synchronous.
Synchronization can be achieved by matching speed and frequency of
motors/generators and/or
turbomachinery of a system with the frequency of one or more grid networks
with which the
system exchanges power. For example, a compressor and a turbine can rotate at
a given, fixed
speed (e.g., 3600 revolutions per minute (rpm)) that is a multiple of grid
frequency (e.g., 60
hertz (Hz)). In some cases, such a configuration may eliminate the need for
additional power
electronics. In some implementations, the turbomachinery and/or the
motors/generators are not
grid synchronous. In such cases, frequency matching can be accomplished
through the use of
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power electronics. In some implementations, the turbomachinery and/or the
motors/generators
are not directly grid synchronous but can be matched through the use of gears
and/or a
mechanical gearbox. As described in greater detail elsewhere herein, the
pumped thermal
systems may also be ranapable. Such capabilities may enable these grid-scale
energy storage
systems to operate as peaking power plants and/or as a load following power
plants. In some
cases, the systems of the disclosure may be capable of operating as base load
power plants.
[0075] Pumped thermal systems can have a given power capacity. In some cases,
power
capacity during charge may differ from power capacity during discharge. For
example, each
system can have a charge and/or discharge power capacity of less than about 1
megawatt (MW),
at least about 1 megawatt, at least about 2 MW, at least about 3 MW, at least
about 4 MW, at
least about 5 MW, at least about 6 MW, at least about 7 MW, at least about 8
MW, at least
about 9 MW, at least about 10 MW, at least about 20 MW, at least about 30 MW,
at least about
40 MW, at least about 50 MW, at least about 75 MW, at least about 100 MW, at
least about
200 MW, at least about 500 MW, at least about 1 gigawatt (GW), at least about
2 GW, at least
about 5 GW, at least about 10 GW, at least about 20 GW, at least about 30 GW,
at least about
40 GW, at least about 50 GW, at least about 75 GW, at least about 100 GW, or
more.
[0076] Pumped thermal systems can have a given energy storage capacity. In one
example, a
pumped thermal system is configured as a 100 MW unit operating for 10 hours.
In another
example, a pumped thermal system is configured as a 1 GW plant operating for
12 hours. In
some instances, the energy storage capacity can be less than about 1 megawatt
hour (MWh), at
least about 1 megawatt hour, at least about 10 MWh, at least about 100 MWh, at
least about 1
gigawatt hour (GWh), at least about 5 GWh, at least about 10 GWh, at least
about 20 GWh, at
least 50 GWh, at least about 100 GWh, at least about 200 GWh, at least about
500 GWh, at
least about 700 GWh, at least about 1000 GWh, or more.
[0077] In some cases, a given power capacity may be achieved with a given
size, configuration
and/or operating conditions of the heat engine/heat pump cycle. For example,
size of
turbomachinery, ducts, heat exchangers, or other system components may
correspond to a
given power capacity.
[0078] In some implementations, a given energy storage capacity may be
achieved with a given
size and/or number of hot side thermal storage tanks and/or cold side thermal
storage tanks.
For example, the heat engine/heat pump cycle can operate at a given power
capacity for a given
amount of time set by the heat storage capacity of the system or plant. The
number and/or heat
storage capacity of the hot side thermal storage tanks may be different from
the number and/or
heat storage capacity of the cold side thermal storage tanks. The number of
tanks may depend
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on the size of individual tanks. The size of hot side storage tanks may differ
from the size of
cold side thermal storage tanks. In some cases, the hot side thermal storage
tanks, the hot side
heat exchanger and the hot side thermal storage medium may be referred to as a
hot side heat
(thermal) storage unit. In some cases, the cold side thermal storage tanks,
the cold side heat
exchanger and the cold side thermal storage medium may be referred to as a
cold side heat
(thermal) storage unit.
[0079] A pumped thermal storage facility can include any suitable number of
hot side storage
tanks, such as at least about 2, at least about 4, at least about 10, at least
about 50, at least about
100, at least about 500, at least about 1,000, at least about 5,000, at least
about 10,000, and the
like. In some examples, a pumped thermal storage facility includes 2, 3, 4, 5.
6, 7, 8, 9, 10, 15,
20, 30, 40, 50, 60, 70, 80, 90, 100, 200, 300, 400, 500, 600, 700, 800, 900,
1,000 or more hot
side tanks.
[0080] A pumped thermal storage facility can also include any suitable number
of cold side
storage tanks, such as at least about 2, at least about 4, at least about 10,
at least about 50, at
least about 100, at least about 500, at least about 1,000, at least about
5,000, at least about
10,000, and the like. In some examples, a pumped thermal storage facility
includes 2, 3, 4, 5,
6, 7, 8, 9, 10, 15, 20, 30, 40, 50, 60, 70, 80, 90, 100, 200, 300, 400, 500,
600, 700, 800, 900,
1,000 or more cold side tanks.
Pumped thermal storage cycles
[0081] An aspect of the disclosure relates to pumped thermal systems operating
on pumped
thermal storage cycles. In some examples, the cycles allow electricity to be
stored as heat (e.g.,
in the form of a temperature differential) and then converted back to
electricity through the use
of at least two pieces of turbomachinery, a compressor and a turbine. The
compressor consumes
work and raises the temperature and pressure of a working fluid (WF). The
turbine produces
work and lowers the temperature and pressure of the working fluid. In some
examples, more
than one compressor and more than one turbine is used. In some cases, the
system can include
at least 1, at least 2, at least 3, at least 4, or at least 5 compressors. In
some cases, the system
can include at least 1, at least 2, at least 3, at least 4, or at least 5
turbines. The compressors
may be arranged in series or in parallel. The turbines may be arranged in
series or in parallel.
[0082] FIGs. 2 and 3 are schematic flow diagrams of working fluid and heat
storage media of
an exemplary pumped thermal system in a charge/heat pump mode and in a
discharge/heat
engine mode, respectively. The system may be idealized for simplicity of
explanation so that
there are no losses (i.e., entropy generation) in either the turbomachinery or
heat exchangers.
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The system_ can include a working fluid 20 (e.g., argon gas) flowing in a
closed cycle between
a compressor 1, a hot side heat exchanger 2, a turbine 3 and a cold side heat
exchanger 4. Fluid
flow paths/directions for the working fluid 20 (e.g., a gas), a hot side
thermal storage (HTS)
medium 21 (e.g., a low viscosity liquid) and a cold side thermal storage (CTS)
medium 22
(e.g., a low viscosity liquid) are indicated by arrows.
[0083] FIGs. 4 and 5 are schematic pressure and temperature diagrams of the
working fluid 20
as it undergoes the charge cycles in FIGs. 2 and 3, respectively, once again
simplified in the
approximation of no entropy generation. Normalized pressure is shown on the y-
axes and
temperature is shown on the x-axes. The direction of processes taking place
during the cycles
is indicated with arrows, and the individual processes taking place in the
compressor 1, the hot
side CFX 2, the turbine 3 and the cold side CFX 4 are indicated on the diagram
with their
respective numerals.
[0084] The heat exchangers 2 and 4 can be configured as counter-flow heat
exchangers
(CFXs), where the working fluid flows in one direction and the substance it is
exchanging heat
with is flowing in the opposite direction. In an ideal counter-flow heat
exchanger with correctly
matched flows (i.e., balanced capacities or capacity flow rates), the
temperatures of the working
fluid and thermal storage medium flip (i.e., the counter-flow heat exchanger
can have unity
effectiveness).
[0085] The counter-flow heat exchangers 2 and 4 can be designed and/or
operated to reduce
entropy generation in the heat exchangers to negligible levels compared to
entropy generation
associated with other system components and/or processes (e.g., compressor
and/or turbine
entropy generation). In some cases, the system may be operated such that
entropy generation
in the system is minimized. For example, the system may be operated such that
entropy
generation associated with heat storage units is minimized. In some cases, a
temperature
difference between fluid elements exchanging heat can be controlled during
operation such that
entropy generation in hot side and cold side heat storage units is minimized.
In some instances,
the entropy generated in the hot side and cold side heat storage units is
negligible when
compared to the entropy generated by the compressor, the turbine, or both the
compressor and
the turbine. In some instances, entropy generation associated with heat
transfer in the heat
exchangers 2 and 4 and/or entropy generation associated with operation of the
hot side storage
unit, the cold side storage unit or both the hot side and cold side storage
units can be less than
about 50%, less than about 25%, less than about 20%, less than about 15%, less
than about
10%, less than about 5%, less than about 4%, less than about 3%, less than
about 2%, or less
than about 1% of the total entropy generated within the system (e.g., entropy
generated by the
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compressor 1, the hot side heat exchanger 2, the turbine 3, the cold side heat
exchanger 4 and/or
other components described herein, such as, for example, a recuperator). For
example, entropy
generation can be reduced or minimized if the two substances exchanging heat
do so at a local
temperature differential AT ¨> 0 (i.e., when the temperature difference
between any two fluid
elements that are in close thermal contact in the heat exchanger is small). In
some examples,
the temperature differential AT between any two fluid elements that are in
close thermal contact
may be less than about 300 Kelvin (K), less than about 200 K, less than about
100 K, less than
about 75 K, less than about 50 K, less than about 40 K, less than about 30 K,
less than about
20 K, less than about 10 K, less than about 5 K, less than about 3 K, less
than about 2 K, or
less than about 1 K. In another example, entropy generation associated with
pressure drop can
be reduced or minimized by suitable design. In some examples, the heat
exchange process can
take place at a constant or near-constant pressure. Alternatively, a non-
negligible pressure drop
may be experienced by the working fluid and/or one or more thermal storage
media during
passage through a heat exchanger. Pressure drop in heat exchangers may be
controlled (e.g.,
reduced or minimized) through suitable heat exchanger design. In some
examples, the pressure
drop across each heat exchanger may be less than about 20% of inlet pressure,
less than about
10% of inlet pressure, less than about 5% of inlet pressure, less than about
3% of inlet pressure,
less than about 2% of inlet pressure, less than about 1% of inlet pressure,
less than about 0.5%
of inlet pressure, less than about 0.25% of inlet pressure, or less than about
0.1% of inlet
pressure.
[0086] Upon entering the heat exchanger 2, the temperature of the working
fluid can either
increase (taking heat from the HTS medium 21, corresponding to the discharge
mode in FIGs.
3 and 5) or decrease (giving heat to the HTS medium 21, corresponding to the
charge mode in
FIGs. 2 and 4), depending on the temperature of the HTS medium in the heat
exchanger relative
the temperature of the working fluid. Similarly, upon entering the heat
exchanger 4, the
temperature of the working fluid can either increase (taking heat from the CTS
medium 22,
corresponding to the charge mode in FIGs. 2 and 4) or decrease (giving heat to
the CTS medium
22, corresponding to the discharge mode in FIGs. 3 and 5), depending on the
temperature of
the CTS medium in the heat exchanger relative the temperature of the working
fluid.
[0087] As described in more detail with reference to the charge mode in FIGs.
2 and 4, the heat
addition process in the cold side CFX 4 can take place over a different range
of temperatures
than the heat removal process in the hot side CFX 2. Similarly, in the
discharge mode in FIGs.
3 and 5, the heat rejection process in the cold side CFX 4 can take place over
a different range
of temperatures than the heat addition process in the hot side CFX 2. At least
a portion of the
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temperature ranges of the hot side and cold side heat exchange processes may
overlap during
charge, during discharge, or during both charge and discharge.
[0088] As used herein, the temperatures To, Ti, To and Ti + are so named
because To, Ti + are
the temperatures achieved at the exit of a compressor with a given compression
ratio r,
adiabatic efficiency i and inlet temperatures of To, T1 respectively_ The
examples in FIGs_ 2,
3, 4 and 5 can be idealized examples where Tic = 1 and where adiabatic
efficiency of the turbine
Tit also has the value Tit = 1.
[0089] With reference to the charge mode shown in FIGs. 2 and 4, the working
fluid 20 can
enter the compressor 1 at position 30 at a pressure P and a temperature T
(e.g., at Ti, P2). As
the working fluid passes through the compressor, work Wi is consumed by the
compressor to
increase the pressure and temperature of the working fluid (e.g.. to Ti, Pi),
as indicated by PT
and TT at position 31. In the charge mode, the temperature Ti + of the working
fluid exiting the
compressor and entering the hot side CFX 2 at position 31 is higher than the
temperature of the
HTS medium 21 entering the hot side CFX 2 at position 32 from a second hot
side thermal
storage tank 7 at a temperature To+ (i.e., To+ < T1+). As these two fluids
pass in thermal contact
with each other in the heat exchanger, the working fluid's temperature
decreases as it moves
from position 31 position 34, giving off heat Qi to the HTS medium, while the
temperature of
the HTS medium in turn increases as it moves from position 32 to position 33,
absorbing heat
Qi from the working fluid. In an example, the working fluid exits the hot side
CFX 2 at position
34 at the temperature To and the HTS medium exits the hot side CFX 2 at
position 33 into a
first hot side thermal storage tank 6 at the temperature Ti. The heat exchange
process can take
place at a constant or near-constant pressure such that the working fluid
exits the hot side CFX
2 at position 34 at a lower temperature but same pressure Pi, as indicated by
P and TI at position
34. Similarly, the temperature of the HTS medium 21 increases in the hot side
CFX 2, while
its pressure can remain constant or near-constant.
[0090] Upon exiting the hot side CFX 2 at position 34 (e.g., at To, Pi), the
working fluid 20
undergoes expansion in the turbine 3 before exiting the turbine at position
35. During the
expansion, the pressure and temperature of the working fluid decrease (e.g.,
to To, P2), as
indicated by PT and TI at position 35. The magnitude of work W7 generated by
the turbine
depends on the enthalpy of the working fluid entering the turbine and the
degree of expansion.
In the charge mode, heat is removed from the working fluid between positions
31 and 34 (in
the hot side CFX 2) and the working fluid is expanded back to the pressure at
which it initially
entered the compressor at position 30 (e.g., 1212). The compression ratio
(e.g., Pi/P)) in the
compressor 1 being equal to the expansion ratio in the turbine 3, and the
enthalpy of the gas
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entering the turbine being lower than the enthalpy of the gas exiting the
compressor, the work
W2 generated by the turbine 3 is smaller than the work WI consumed by the
compressor 1 (i.e.,
W2 < Wi).
[0091] Because heat was taken out of the working fluid in the hot side CFX 2,
the temperature
To at which the working fluid exits the turbine at position 35 is lower than
the temperature Ti
at which the working fluid initially entered the compressor at position 30. To
close the cycle
(i.e., to return the pressure and temperature of the working fluid to their
initial values T1, P2 at
position 30), heat Q2 is added to the working fluid from the CTS medium 22 in
the cold side
CFX 4 between positions 35 and 30 (i.e., between the turbine 3 and the
compressor 1). In an
example, the CTS medium 22 enters the cold side CFX 4 at position 36 from a
first cold side
thermal storage tank 8 at the temperature Ti and exits the cold side CFX 4 at
position 37 into a
second cold side thermal storage tank 9 at the temperature To, while the
working fluid 20 enters
the cold side CFX 4 at position 35 at the temperature To and exits the cold
side CFX 4 at
position 30 at the temperature Ti. Again, the heat exchange process can take
place at a constant
or near-constant pressure such that the working fluid exits the cold side CFX
2 at position 30
at a higher temperature but same pressure P2, as indicated by P and TT at
position 30. Similarly,
the temperature of the CTS medium 22 decreases in the cold side CFX 2, while
its pressure can
remain constant or near-constant.
[0092] During charge, the heat Q2 is removed from the CTS medium and the heat
Qi is added
to the HTS medium, wherein Qi > ci/. A net amount of work Wi - W/ is consumed,
since the
work Wi used by the compressor is greater than the work W2 generated by the
turbine. A device
that consumes work while moving heat from a cold body or thermal storage
medium to a hot
body or thermal storage medium is a heat pump; thus, the pumped thermal system
in the charge
mode operates as a heat pump.
[0093] In an example, the discharge mode shown in FIGs. 3 and 5 can differ
from the charge
mode shown in FIGs. 2 and 4 in the temperatures of the thermal storage media
being introduced
into the heat exchangers. The temperature at which the HTS medium enters the
hot side CFX
2 at position 32 is Ti + instead of To+, and the temperature of the CTS medium
entering the cold
side CFX 4 at position 36 is To instead of Ti. During discharge, the working
fluid enters the
compressor at position 30 at To and P2, exits the compressor at position 31 at
To < Ti and Pi,
absorbs heat from the HTS medium in the hot side CFX 2, enters the turbine 3
at position 34
at Ti' and Pi, exits the turbine at position 35 at Ti > To and P2, and finally
rejects heat to the
CTS medium in the cold side CFX 4, returning to its initial state at position
30 at To and
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[0094] The HTS medium at temperature Ti + can be stored in a first hot side
thermal storage
tank 6, the HTS medium at temperature To can be stored in a second hot side
thermal storage
tank 7, the CTS medium at temperature Ti can be stored in a first cold side
thermal storage
tank 8, and the CTS medium at temperature To can be stored in a second cold
side thermal
storage tank 9 during both charge and discharge modes. In one implementation,
the inlet
temperature of the HTS medium at position 32 can be switched between Ii + and
To by
switching between tanks 6 and 7, respectively. Similarly, the inlet
temperature of the CTS
medium at position 36 can be switched between Ti and To by switching between
tanks 8 and
9, respectively. Switching between tanks can be achieved by including a valve
or a system of
valves (e.g., valve systems 12 and 13 in FIG. 7) for switching connections
between the hot side
heat exchanger 2 and the hot side tanks 6 and 7, and/or between the cold side
heat exchanger 4
and the cold side tanks 8 and 9 as needed for the charge and discharge modes.
In some
implementations, connections may be switched on the working fluid side
instead, while the
connections of storage tanks 6, 7, 8 and 9 to the heat exchangers 2 and 4
remain static. In some
examples, flow paths and connections to the heat exchangers may depend on the
design (e.g.,
shell-and-tube) of each heat exchanger. In some implementations, one or more
valves can be
used to switch the direction of both the working fluid and the heat storage
medium through the
counter-flow heat exchanger on charge and discharge. Such configurations may
be used, for
example, due to high thermal storage capacities of the heat exchanger
component, to decrease
or eliminate temperature transients, or a combination thereof. In some
implementations, one or
more valves can be used to switch the direction of only the working fluid,
while the direction
of the HTS or CTS can be changed by changing the direction of pumping, thereby
maintaining
the counter-flow configuration. In some implementations, different valve
configurations may
be used for the HTS and the CTS. Further, any combination of the valve
configurations herein
may be used. For example, the system may be configured to operate using
different valve
configurations in different situations (e.g., depending on system operating
conditions).
[0095] In the discharge mode shown in FIGs. 3 and 5, the working fluid 20 can
enter the
compressor 1 at position 30 at a pressure P and a temperature T (e.g., at To,
Pi). As the working
fluid passes through the compressor, work Wi is consumed by the compressor to
increase the
pressure and temperature of the working fluid (e.g., to Tot, Pi), as indicated
by P-1 and TT at
position 31. In the discharge mode, the temperature To of the working fluid
exiting the
compressor and entering the hot side CFX 2 at position 31 is lower than the
temperature of the
HTS medium 21 entering the hot side CFX 2 at position 32 from a first hot side
thermal storage
tank 6 at a temperature Ti+ (i.e., To+ < Ti). As these two fluids pass in
thermal contact with
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each other in the heat exchanger, the working fluid's temperature increases as
it moves from
position 31 position 34, absorbing heat Qi from the HTS medium, while the
temperature of the
HTS medium in turn decreases as it moves from position 32 to position 33,
giving off heat Qi
to the working fluid. In an example, the working fluid exits the hot side CFX
2 at position 34
at the temperature T1+ and the HTS medium exits the hot side CFX 2 at position
33 into the
second hot side thermal storage tank 7 at the temperature Tot The heat
exchange process can
take place at a constant or near-constant pressure such that the working fluid
exits the hot side
CFX 2 at position 34 at a higher temperature but same pressure Pi, as
indicated by P and Ti at
position 34. Similarly, the temperature of the HTS medium 21 decreases in the
hot side CFX
2, while its pressure can remain constant or near-constant.
[0096] Upon exiting the hot side CFX 2 at position 34 (e.g., at T1+, Pi), the
working fluid 20
undergoes expansion in the turbine 3 before exiting the turbine at position
35. During the
expansion, the pressure and temperature of the working fluid decrease (e.g.,
to Ti, P2), as
indicated by Pi and 11, at position 35. The magnitude of work W2 generated by
the turbine
depends on the enthalpy of the working fluid entering the turbine and the
degree of expansion.
In the discharge mode, heat is added to the working fluid between positions 31
and 34 (in the
hot side CFX 2) and the working fluid is expanded back to the pressure at
which it initially
entered the compressor at position 30 (e.g., P2). The compression ratio (e.g.,
Pi/P2) in the
compressor 1 being equal to the expansion ratio in the turbine 3, and the
enthalpy of the gas
entering the turbine being higher than the enthalpy of the gas exiting the
compressor, the work
W2 generated by the turbine 3 is greater than the work Wi consumed by the
compressor 1 (i.e.,
W2 > W 1)-
[0097] Because heat was added to the working fluid in the hot side CFX 2, the
temperature Ti
at which the working fluid exits the turbine at position 35 is higher than the
temperature To at
which the working fluid initially entered the compressor at position 30. To
close the cycle (i.e.,
to return the pressure and temperature of the working fluid to their initial
values To, P2 at
position 30), heat Q2 is rejected by the working fluid to the CTS medium 22 in
the cold side
CFX 4 between positions 35 and 30 (i.e., between the turbine 3 and the
compressor 1). The
CTS medium 22 enters the cold side CFX 4 at position 36 from a second cold
side thermal
storage tank 9 at the temperature To and exits the cold side CFX 4 at position
37 into a first
cold side thermal storage tank 8 at the temperature T1, while the working
fluid 20 enters the
cold side CFX 4 at position 35 at the temperature Ti and exits the cold side
CFX 4 at position
30 at the temperature To. Again, the heat exchange process can take place at a
constant or near-
constant pressure such that the working fluid exits the cold side CFX 2 at
position 30 at a higher
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temperature but same pressure P), as indicated by P and Tt at position 30.
Similarly, the
temperature of the CTS medium 22 increases in the cold side CFX 2, while its
pressure can
remain constant or near-constant.
[0098] During discharge, the heat Q, is added to the CTS medium and the heat
Q1 is removed
from the HTS medium, wherein Qi > Q,. A net amount of work W, - Wi is
generated, since
the work W I used by the compressor is smaller than the work W2 generated by
the turbine. A
device that generates work while moving heat from a hot body or thermal
storage medium to a
cold body or thermal storage medium is a heat engine; thus, the pumped thermal
system in the
discharge mode operates as a heat engine.
[0099] FIG. 6 is a simplified schematic perspective view of a closed working
fluid system in
the pumped thermal system in FIGs. 2-3. As indicated, the working fluid 20
(contained inside
tubing) circulates clockwise between the compressor 1, the hot side heat
exchanger 2, the
turbine 3, and the cold side heat exchanger 4. The compressor 1 and the
turbine 3 can be ganged
on a common mechanical shaft 10 such that they rotate together. In some
implementations, the
compressor 1 and the turbine 3 can have separate mechanical shafts. A
motor/generator 11
(e.g., including a synchronous motor - synchronous generator converter on a
single common
shaft) provides power to and from the turbomachinery. In this example, the
compressor, the
turbine and the motor/generator are all located on a common shaft. Pipes at
positions 32 and
33 transfer hot side thermal storage fluid to and from the hot side heat
exchanger 2,
respectively. Pipes at positions 36 and 37 transfer cold side thermal storage
fluid to and from
the cold side heat exchanger 4. respectively.
[0100] Although the system of FIG. 6 is illustrated as comprising a compressor
1 and turbine
3, the system can include one or more compressors and one or more turbines,
which may
operate, for example, in a parallel configuration, Or alternatively in a
series configuration or in
a combination of parallel and series configurations. In some examples, a
system of compressors
or turbines may be assembled such that a given compression ratio is achieved.
In some cases,
different compression ratios (e.g., on charge and discharge) can be used
(e.g., by connecting or
disconnecting, in a parallel and/or series configuration, one or more
compressors or turbines
from the system of compressors or turbines). In some examples, the working
fluid is directed
to a plurality of compressors and/or a plurality of turbines. In some
examples, the compressor
and/or turbine may have temperature dependent compression ratios. Arrangement
and/or
operation of the turbomachinery and/or other elements of the system may be
adjusted in
accordance with the temperature dependence (e.g., to optimize performance).
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[0101] FIG. 7 is a simplified schematic perspective view of the pumped thermal
system in
FIGs. 2-3 with hot side and cold side storage tanks and a closed cycle working
fluid system. In
this example, the HTS medium is a molten salt and the CTS medium is a low
temperature
liquid. One, two or more first hot side tanks 6 (at the temperature Ti) and
one, two or more
second hot side tanks 7 (at the temperature To), both for holding the HTS
medium, are in fluid
communication with a valve 13 configured to transfer the HTS medium to and
from the hot
side heat exchanger 2. One, two or more first cold side tanks 8 (at the
temperature Ti) and one,
two or more second cold side tanks 9 (at the temperature To), both for holding
the CTS medium,
are in fluid communication with a valve 12 configured to transfer the CTS
medium to and from
the cold side heat exchanger 4.
[0102] The thermal energy reservoirs or storage tanks may be thermally
insulated tanks that
can hold a suitable quantity of the relevant thermal storage medium (e.g.,
heat storage fluid).
The storage tanks may allow for relatively compact storage of large amounts of
thermal energy.
In an example, the hot side tanks 6 and/or 7 can have a diameter of about 80
meters, while the
cold side tanks 8 and/or 9 can have a diameter of about 60 meters. In another
example, the size
of each (i.e., hot side or cold side) thermal storage for a 1 GW plant
operating for 12 hours can
be about 20 medium-sized oil refinery tanks.
[0103] In some implementations, a third set of tanks containing storage media
at intermediate
temperatures between the other tanks may be included on the hot side and/or
the cold side. In
an example, a third storage or transfer tank (or set of tanks) at a
temperature intermediate to
the temperatures of a first tank (or set of tanks) and a second tank (or set
of tanks) may be
provided. A set of valves may be provided for switching the storage media
between the
different tanks and heat exchangers. For example, thermal media may be
directed to different
sets of tanks after exiting the heat exchangers depending on operating
conditions and/or cycle
being used. In some implementations, one or more additional sets of storage
tanks at different
temperatures may be added on the hot side and/or the cold side.
[0104] The storage tanks (e.g., hot side tanks comprising hot side thermal
storage medium
and/or cold side tanks comprising cold side thermal storage medium) may
operate at ambient
pressure. In some implementations, thermal energy storage at ambient pressure
can provide
safety benefits. Alternatively, the storage tanks may operate at elevated
pressures, such as, for
example, at a pressure of at least about 2 atm, at least about 5 atm, at least
about 10 atm, at
least about 20 atm, or more. Alternatively, the storage tanks may operate at
reduced pressures,
such as, for example, at a pressure of at most about 0.9 atm, at most about
0.7 atm, at most
about 0.5 atm, at most about 0.3 atm, at most about 0.1 atm, at most about
0.01 atm, at most
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about 0.001 atm, or less. In some cases (e.g., when operating at
higher/elevated or lower
pressures or to avoid contamination of the thermal storage media), the storage
tanks can be
sealed from the surrounding atmosphere. Alternatively, in some cases, the
storage tanks may
not be sealed. In some implementations, the tanks may include one or more
pressure regulation
or relief systems (e.g., a valve for safety or system optimization).
[0105] As used herein, the first hot side tank(s) 6 (at the temperature Ti)
can contain HTS
medium at a higher temperature than the second hot side tank(s) 7 (at the
temperature Tot), and
the first cold side tank(s) 8 (at the temperature Ti) can contain CTS medium
at a higher
temperature than the second cold side tank(s) 9 (at the temperature To).
During charge, HTS
medium in the first (higher temperature) hot side tank(s) 6 and/or CTS medium
in the second
(lower temperature) cold side tank(s) 9 can be replenished. During discharge,
HTS medium in
the first (higher temperature) hot side tank(s) 6 and/or CTS medium in the
second (lower
temperature) cold side tank(s) 9 can be consumed.
[0106] In the foregoing examples, in either mode of operation, two of the four
storage tanks 6,
7, 8 and 9 are feeding thermal storage medium to the heat exchangers 2 and 4
at the inlets 32
and 36, respectively, and the other two tanks are receiving thermal storage
medium from the
heat exchangers 2 and 4 from the exits 33 and 37, respectively. In this
configuration, the feed
tanks can contain a storage medium at a given temperature due to prior
operating conditions,
while the receiving tanks' temperatures can depend on current system operation
(e.g., operating
parameters, loads and/or power input). The receiving tank temperatures may be
set by the
Brayton cycle conditions. In some cases, the receiving tank temperatures may
deviate from
desired values due to deviations from predetermined cycle conditions (e.g.,
variation of
absolute pressure in response to system demand) and/or due to entropy
generation within the
system. In some cases (e.g., due to entropy generation), at least one of the
four tank
temperatures can be higher than desired. In some implementations, a radiator
can be used to
reject or dissipate this waste heat to the environment. In some cases, heat
rejection to the
environment may be enhanced (e.g., using evaporative cooling etc.). The waste
heat generated
during operation of the pumped thermal systems herein can also be utilized for
other purposes.
For example, waste heat from one part of the system may be used elsewhere in
the system. In
another example, waste heat may be provided to an external process or system,
such as, for
example, a manufacturing process requiring low grade heat, commercial or
residential heating,
thermal desalination, commercial drying operations etc.
[0107] Components of pumped thermal systems of the disclosure may exhibit non-
ideal
performance, leading to losses and/or inefficiencies. The major losses in the
system may occur
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due to inefficiencies of the turbomachinery (e.g., compressor and turbine) and
the heat
exchangers. The losses due to the heat exchangers may be small compared to the
losses due to
the turbomachinery. In some implementations, the losses due to the heat
exchangers can be
reduced to near zero with suitable design and expense. Therefore, in some
analytical examples,
losses due to the heat exchangers and other possible small losses due to
pumps, the
motor/generator and/or other factors may be neglected.
[0108] Losses due to turbomachinery can be quantified in terms of adiabatic
efficiencies tic and
tit (also known as isentropic efficiencies) for compressors and turbines,
respectively. For large
turbomachinery, typical values may range between /7, = 0.85 - 0.9 for
compressors and qt =
0.9 - 0.95 for turbines. The actual amount of work produced or consumed by a
cycle can then
w (out) w (in) w (out) 1 w (in)
be expressed as AIN
where, in an example
= v'actual ''actual = rlt"ideal vvideal,
assuming constant specific heats of the working fluid, W
id( tena) = cpTinlet OP ¨ 1), wid(oeuati) =
v-,
cpTinlet(1 ), where = r r is the compression ratio
(i.e., ratio of the higher pressure
to the lower pressure), and 7 = cp/c, is the ratio of specific heats of the
working fluid. Due to
compressor and turbine inefficiencies, more work is required to achieve a
given compression
ratio during compression, and less work is generated during expansion for a
given compression
ratio. Losses can also be quantified in terms of the polytropic, or single
stage, efficiencies, rfrp
and IN, for compressors and turbines, respectively. The polytropic
efficiencies are related to
ip -1 -
tp-ntP
the adiabatic efficiencies ric and gi by the equations Tic = 1, and rh =
Ucp_i 1-
1p-i =
[0109] In examples where qc = = 1, pumped thermal cycles of the disclosure can
follow
identical paths in both charge and discharge cycles (e.g., as shown in FIGs. 4
and 5). In
examples where tic < 1 and/or tit < 1, compression in the compressor can lead
to a greater
temperature increase than in the ideal case for the same compression ratio,
and expansion in
the turbine can lead to a smaller temperature decrease than in the ideal case.
[0110] In some implementations, the polytropic efficiency of the compressor
qcp may be at
least about 0.3, at least about 0.5, at least about 0.6, at least about 0.7,
at least about 0.75, at
least about 0.8, at least about 0.85, at least about 0.9, at least about 0.91,
at least about 0.92, at
least about 0.93, at least about 0.96, or more. In some implementations, the
polytropic
efficiency of the compressor rhp may be at least about 0.3, at least about
0.5, at least about 0.6,
at least about 0.7, at least about 0.75, at least about 0.8, at least about
0.85, at least about 0.9,
at least about 0.91, at least about 0.92, at least about 0.93, at least about
0.96, at least about
0.97 or more.
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[0111] To, Ti were previously defined as the temperatures achieved at the exit
of a
compressor with a given compression ratio r, adiabatic efficiency rif and
inlet temperatures of
To, Ti respectively. In some examples, these four temperatures are related by
the equation T# =
Ti = oi/ncp
[0112] FIG. 8 shows an exemplary heat storage charge cycle for a water
(CTS)/molten salt
(HTS) system with ric = 0.9 and rit = 0.95. The dashed lines correspond to tit
= tit = 1 and the
solid lines show the charge cycle with lit= 0.95 and Tic = 0.9. In this
example, the CTS medium
on the cold side is water, and the HTS medium on the hot side is molten salt.
In some cases,
the system can include 4 heat storage tanks. In the charge cycle, the working
fluid at To and P2
can exchange heat with a CTS medium in the cold side heat exchanger 4, whereby
its
temperature can increase to T1 (assuming negligible pressure drop, its
pressure can remain P2).
In the compressor 1 with i = 0.9, the temperature and pressure of the working
fluid can
increase from Ti , P2 to Ti+, Pi. The working fluid can then exchange heat
with an HIS medium
in the hot side heat exchanger 2, such that its temperature can decrease (at
constant pressure
P1, assuming negligible pressure drop). If the working fluid enters the
turbine 3 with rit= 0.95
at the temperature To+ and expands back to its original pressure 132, its
temperature when exiting
the turbine may not be To. Instead, the working fluid may enter the turbine at
a temperature Do+
and exit the turbine at the temperature To and pressure P2. In some examples,
the temperatures
are related by the relation ¨T = letp . In some examples, -T6' is the
temperature at which the
To
working fluid enters the inlet of a turbine with adiabatic efficiency nt and
compression ratio r
in order to exit at the temperature To.
[0113] In some implementations, the temperature Dit may be incorporated into
charge cycles
of the disclosure by first heat exchanging the working fluid with the HTS
medium from Ti+ to
To+, followed by further cooling the working fluid from To+ to Do+, as
illustrated by section 38
of the cycle in FIG. 8.
[0114] FIG. 9 shows an exemplary heat storage discharge (extraction) cycle for
the
water/molten salt system in FIG. 8 with = 0.9 and /yr = 0.95. The dashed lines
correspond to
17, = Tit= 1 and the solid lines show the charge cycle with lit= 0.95 and nc=
0.9. In the discharge
cycle, the working fluid at Ti and P2 can exchange heat with a CTS medium in
the cold side
heat exchanger 4, whereby its temperature can decrease to To (assuming
negligible pressure
drop, its pressure can remain P2). In the compressor 1 with
= 0.9, the temperature and
pressure of the working fluid can increase from To, P2 to To+, Pi. The working
fluid can then
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exchange heat with an HTS medium in the hot side heat exchanger 2, such that
its temperature
can increase (at constant pressure Pi, assuming negligible pressure drop).
Working fluid
entering the turbine 3 at Ti+ may not exit the turbine at the temperature Ti
as in the charge
cycle, but may instead exit at a temperature D1, where, in some examples, -D1
= Ti+1/-1hP . In
some examples, 11 is the temperature at which the working fluid exits the
outlet of a turbine
with adiabatic efficiency nt and compression ratio r after entering the inlet
of the turbine at the
temperature T.
[0115] In some implementations, the temperature Di may be incorporated into
the discharge
cycles of the disclosure by first cooling the working fluid exiting the
turbine at T1 to Ti, as
illustrated by section 39 of the cycle in FIG. 9, followed by heat exchanging
the working fluid
with the CTS medium from Ti to To.
[0116] The charge and discharge cycles may be closed by additional heat
rejection operations
in sections 38 (between To+ and Do+) and 39 (between T1 and Ti), respectively.
In some cases,
closing the cycles through heat rejection in sections of the cycles where the
working fluid can
reject heat to ambient at low cost may eliminate the need for additional heat
input into the
system. The sections of the cycles where the working fluid can reject heat to
ambient may be
limited to sections where the temperature of the working fluid is high enough
above ambient
temperature for ambient cooling to be feasible. In some examples, heat may be
rejected to the
environment in sections 38 and/or 39. For example, heat may be rejected using
one or more
working fluid to air radiators, intermediate water cooling, or a number of
other methods. In
some cases, heat rejected in sections 38 and/or 39 may be used for another
useful purpose, such
as, for example, cogeneration, thermal desalination and/or other examples
described herein.
[0117] In some implementations, the cycles may be closed by varying the
compression ratios
between the charge and discharge cycles, as shown, for example, in FIG. 10.
The ability to vary
compression ratio on charge and discharge may be implemented, for example, by
varying the
rotation speed of the compressor and/or turbine, by variable stator pressure
control, by
bypassing a subset of the compression or expansion stages on charge or
discharge by the use
of valves, or by using dedicated compressor/turbine pairs for charge and
discharge mode. In
one example, the compression ratio in the discharge cycle in FIG_ 9 can be
changed such that
heat rejection in section is 39 is not used, and only heat rejection in
section 38 in the charge
cycle is used. Varying the compression ratio may allow heat (i.e., entropy) to
be rejected at a
lower temperature, thereby increasing overall roundtrip efficiency. In some
examples of this
lincp
configuration, the compression ratio on charge, rci, can be set such that .-=
Ipc , and on
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discharge, the compression ratio ID can be set such that ¨ = 11.1ntp In some
cases, the upper
temperatures Ti+ and Ti can be identical on charge and discharge and no heat
removal may be
needed in this portion (also "leg" herein) of the cycle. In such cases, the
temperature To on
Vtp
linp
c
charge (e.g.. To +(c) = Topc ) and the temperature To + on discharge (e.g.,
To+(d) = TAD )
can be dissimilar and heat may be rejected (also "dissipated" or "dumped"
herein) to the
+(d)
environment between the temperatures To+(c) and To . In an implementation
where only the
storage media exchange heat with the environment, a heat rejection device
(e.g., devices 55
and 56 shown in FIG. 16) can be used to lower the temperature of the CTS from
T0+ (d) to
To+(c)between discharge and charge.
[0118] FIG. 10 shows an example of a cycle with variable compression ratios.
The
compression ratio can be higher on discharge (when work is produced by the
system) than on
charge (when work is consumed by the system), which may increase an overall
round trip
efficiency of the system. For example, during a charge cycle 80 with To+(c), a
lower
compression ratio of < 3 can be used; during a discharge cycle 81 with To
+(d), a compression
ratio of > 3 can he used. The upper temperatures reached in both cycles 80 and
81 can be Ti
and Ti+, and no excess heat may be rejected.
[0119] The compression ratio may be varied between charge and discharge such
that the heat
dissipation to the environment needed for closing the cycle on both charge and
discharge occurs
(c)
between the temperatures TOT (the temperature of the working fluid before it
enters the turbine
during the charge cycle) and To (D)(the temperature of the working fluid as it
exits the
compressor on discharge) and not above the temperature Ti (the temperature of
the working
fluid before it enters the compressor on charge and/or exits the turbine on
discharge). In some
examples, none of the heat is rejected at a temperature above the lowest
temperature of the
HTS medium.
[0120] In the absence of system losses and/or inefficiencies, such as, for
example, in the case
of pumped thermal systems comprising heat pump(s) and heat engine(s) operating
at the zero
entropy creation/isentropic limit, a given amount of heat QH can be
transferred using a given
quantity of work Win heat pump (charge) mode, and the same QH can be used in
heat engine
(discharge) mode to produce the same work W, leading to a unity (i.e., 100%)
roundtrip
efficiency. In the presence of system losses and/or inefficiencies, roundtrip
efficiencies of
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pumped thermal systems may be limited by how much the components deviate from
ideal
performance.
[0121] The roundtrip efficiency of a pumped thermal system may be defined as
71
'store =
Iwcevxtract Iwccvharge
In some examples, with an approximation of ideal heat exchange, the
roundtrip efficiency can be derived by considering the net work output during
the discharge
cycle, iwcevxtracti = Thwioduetat wIal, and the net work input during the
charge cycle,
,charge W (LI
I vvev I ¨ ¨ ¨ ThWic:luetai using the equations for work and
temperature given above.
nc
[0122] Roundtrip efficiencies may be calculated for different configurations
of pumped
thermal systems (e.g., for different classes of thermal storage media) based
on turbomachinery
component efficiencies, 17, and Tit.
[0123] In one example, FIG. 11 shows roundtrip efficiency contours for a
water/salt system,
such as, for example, the water/salt system in FIGs. 8 and 9 with To = 273 K
(0 C), Ti = 373
K (100 C) and a compression ratio of r = 5.65 chosen to achieve compatibility
with the salt(s)
on the hot side. Exemplary roundtrip efficiency contours at values of 17
-,storeOf 10%, 20%, 30%,
40%, 50%, 60%, 70%, 80% and 90% are shown as a function of component
efficiencies lie and
qi on the x- and y-axes, respectively. The symbols e and 0 represent the
approximate range
of present large turbomachinery adiabatic efficiency values. The dashed arrows
represent the
direction of increasing efficiency.
[0124] FIG. 12 shows roundtrip efficiency contours for a colder storage/salt
system, such as,
for example a hexane/salt system with a gas-gas heat exchanger in FIGs. 13,
14, 17 and 18 with
To = 194 K (-79 C), Ti = 494 K (221 C) and a compression ratio of r = 3.28.
Exemplary
roundtrip efficiency contours at values of nstoreof 10%, 20%, 30%, 40%, 50%,
60%, 70%,
-,
80% and 90% are shown as a function of component efficiencies i and i on the x-
and y-axes,
respectively. The symbols e and 0 represent the approximate range of present
large
turbomachinery adiabatic efficiency values. As discussed in detail elsewhere
herein, using
hexane, heptane and/or another CTS medium capable of low temperature operation
may result
in significant improvements of system efficiency.
Pumped thermal storage cycles with recuperation
[0125] Another aspect of the disclosure is directed to pumped thermal systems
with
recuperation. In some situations, the terms regeneration and recuperation can
be used
interchangeably, although they may have different meanings. As used herein,
the terms
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"recuperation" and "recuperator" generally refer to the presence of one or
more additional heat
exchangers where the working fluid exchanges heat with itself during different
segments of a
thermodynamic cycle through continuous heat exchange without intermediate
thermal storage.
The roundtrip efficiency of pumped thermal systems may be substantially
improved if the
allowable temperature ranges of the storage materials can be extended. In some
implementations, this may be accomplished by choosing a material or medium on
the cold side
that can go to temperatures below 273 K (0 C). For example, a CTS medium
(e.g., hexane)
with a low temperature limit of approximately To = 179 K (-94 C) may be used
in a system
with a molten salt HTS medium. However, To (i.e., the lowest temperature of
the working
fluid in the hot side heat exchanger) at some (e.g., modest) compression
ratios may be below
the freezing point of the molten salt, making the molten salt unviable as the
HTS medium. In
some implementations, this can be resolved by including a working fluid to
working fluid (e.g.,
gas-gas) heat exchanger (also "recuperator" herein) in the cycle.
[0126] FIG. 13 is a schematic flow diagram of working fluid and heat storage
media of a
pumped thermal system in a charge/heat pump mode with a gas-gas heat exchanger
5 for the
working fluid. The use of the gas-gas heat exchanger can enable use of colder
heat storage
medium on the cold side of the system. The working fluid can be air. The
working fluid can be
dry air. The working fluid can be nitrogen. The working fluid can be argon.
The working fluid
can be a mixture of primarily argon mixed with another gas such as helium. For
example, the
working fluid may comprise at least about 50% argon, at least about 60% argon,
at least about
70% argon, at least about SO% argon, at least about 90% argon, or about 100%
argon, with
balance helium.
[0127] FIG. 17 shows a heat storage charge cycle for the storage system in
FIG. 13 with a cold
side storage medium (e.g., liquid hexane) capable of going down to
approximately to 179 K (-
94 C) and a molten salt as the hot side storage, and 17, = 0.9 and qt = 0.95.
The CTS medium
can be hexane or heptane and the HTS medium can be molten salt In some cases,
the system
can include four heat storage tanks.
[0128] In one implementation, during charge in FIGs. 13 and 17, the working
fluid enters the
compressor at Ti and P2, exits the compressor at Ti + and P1, rejects heat Qi
to the HTS medium
21 in the hot side CFX 2, exiting the hot side CFX 2 at Ti and P1, rejects
heat Qrecup (also
"Qregen" herein, as shown, for example, in the accompanying drawings) to the
cold (low
pressure) side working fluid in the heat exchanger or recuperator 5, exits the
recuperator 5 at
To and Pi, rejects heat to the environment (or other heat sink) in section 38
(e.g., a radiator),
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enters the turbine 3 at T6' and P1, exits the turbine at TO and P?, absorbs
heat Q2 from the CTS
medium 22 in the cold side CFX 4, exiting the cold side CFX 4 at To and P2,
absorbs heat
Qrecup from the hot (high pressure) side working fluid in the heat exchanger
or recuperator 5,
and finally exits the recuperator 5 at T1 and P2, returning to its initial
state before entering the
compressor.
[0129] FIG. 14 is a schematic flow diagram of working fluid and heat storage
media of the
pumped thermal system in FIG. 13 in a discharge/heat engine mode. Again, the
use of the gas-
gas heat exchanger can enable use of colder heat storage fluid (CTS) and/or
colder working
fluid on the cold side of the system.
[0130] FIG. 18 shows a heat storage discharge cycle for the storage system for
the storage
system in FIG. 14 with a cold side storage medium (e.g., liquid hexane)
capable of going down
to 179 K (-94 C) and a molten salt as the hot side storage, and tic = 0.9 and
tit = 0.95. Again,
the CTS medium can be hexane or heptane and the HTS medium can be molten salt,
and the
system may include 4 heat storage tanks.
[0131] During discharge in FIGs. 14 and 18, the working fluid enters the
compressor at To and
P2, exits the compressor at To and Pi, absorbs heat 0
,r.cup from the cold (low pressure) side
working fluid in the heat exchanger or recuperator 5, exits the recuperator 5
at T1 and P1,
absorbs heat Qt from the HTS medium 21 in the hot side CFX 2, exiting the hot
side CFX 2 at
Tih and Pi, enters the turbine 3 at Ti + and Pi, exits the turbine at T1 and
P), rejects heat to the
environment (or other heat sink) in section 39 (e.g., a radiator), rejects
heat o
_re. to the hot
(high pressure) side working fluid in the heat exchanger or recuperator 5,
enters the cold side
CFX 4 at To+ and P2, rejects heat Q7 to the CTS medium 22 in the cold side CFX
4, and finally
exits the cold side CFX 4 at To and P2, returning to its initial state before
entering the
compressor.
[0132] In another implementation, shown in FIG. 15, the charge cycle remains
the same as in
FIGs. 13 and 17, except that the working fluid exits the recuperator 5 at To+
and P1 (instead of
at To and Pi as in FIGs. 13 and 17), enters the turbine 3 at 'To+and Pi, exits
the turbine at To
and P2, absorbs heat Q/ from the CTS medium 22 having a temperature Tjt
(instead of at To+
as in FIGs. 13 and 17) in the cold side CFX 4, and exits the cold side CFX 4
at To+ and P?
(instead of at To and P7 as in FIG. 13) before reentering the recuperator 5.
The heat between
temperatures To+ and T6E is no longer rejected from the working fluid to the
environment
directly (as in section 38 in FIGs. 13 and 17).
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[0133] During discharge in FIG. 16, the discharge cycle remains the same as in
FIGs. 14 and
8B, except that the temperature of the HTS medium being deposited in tank 7 is
changed. The
working fluid exits the recuperator 5 at D1 and Pi (instead of at Ti and Pi as
in FIGs. 14 and
8B) and absorbs heat Qi from the HTS medium 21 in the hot side CFX 2. The HTS
medium
exits the hot side CFX 2 having a temperature 71 (instead of at T1 as in FIGs.
14 and 18).The
working fluid then exits the hot side CFX 2 at Ti and Pi, enters the turbine 3
at Ti' and Pi,
and exits the turbine at T1 and 13/ before reentering the recuperator 5. The
heat between
temperatures Di_ and T1 is no longer rejected from the working fluid to the
environment directly
(as in section 39 in FIGs. 14 and 18). As in FIG. 14, the CTS medium enters
the tank 8 at
temperature Tot
[0134] After the discharge in FIG. 16, in preparation for the charge in FIG.
15, heat exchange
with ambient may be used to cool the HTS medium 21 from the temperature pi
used in the
discharge cycle to the temperature Ti used in the charge cycle. Similarly,
heat exchange with
ambient may be used to cool the CTS medium 22 from the temperature To used in
the
discharge cycle to the temperature T used in the charge cycle. Unlike in the
configuration in
FIGs. 13 and 14, where the working fluid may need to reject a substantial
amount of heat (in
sections 38 and 39, respectively) at a fast rate, in this configuration, the
hot side and cold side
storage media may he cooled at an arbitrarily slow rate (e.g., by radiating
away or by other
means of giving off the heat to the environment).
[0135] As shown in FIG. 16, in some implementations, heat can be rejected from
the CTS
medium to the environment by circulating the CTS medium in the tank 8 in a
heat rejection
device 55 that can absorb heat from the CTS medium and reject heat to the
environment until
the CTS medium cools from the temperature To+ to the temperature T6E. In some
examples, the
heat rejection device 55 can be, for example, a radiator, a thermal bath
containing a substance
such as water or salt water, or a device immersed in a natural body of water
such as a lake,
river or ocean. In some examples, the heat rejection device 55 can also be an
air cooling
apparatus, or a series of pipes which are thermally connected to a solid
reservoir (e.g., pipes
embedded in the ground).
[0136] Similarly, in some implementations, heat can be rejected from the HTS
medium to the
environment by circulating the HTS in the tank 7 in a heat rejection device 56
that can absorb
heat from the HTS medium and reject heat to the environment until the HTS
medium cools
from the temperature Di to the temperature Ti. In some examples, the heat
rejection device 56
can be, for example, a radiator, a thermal bath containing a substance such as
water or salt
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water, or a device immersed in a natural body of water such as a lake, river
or ocean. In some
examples, the heat rejection device 56 can also be an air cooling apparatus or
a series of pipes
which are thermally connected to a solid reservoir (e.g., pipes embedded in
the ground).
[0137] In some implementations, rejecting heat to ambient through the use of
the thermal
storage media may be used in conjunction with the variable compression ratio
charge and/or
discharge cycles described, for example, in FIG. 10. In this system, only the
CTS medium may
exchange heat with ambient. Such a system can also be implemented with a
recuperator to
extend the temperature ranges of the HTS and CTS media in the cycles.
[0138] In some implementations, three separate cold side storage tanks at
respective
temperatures To, -1-'6E and T6E may be used (e.g., an extra tank may be used
in addition to the
tanks 8 and 9). During heat exchange in the cold side CFX 4 in the discharge
cycle, heat from
the working fluid exiting the recuperator 5 may be transferred to the CTS
medium in the Td' -
tank. The CTS medium may be cooled in/by, for example, the heat rejection
device 55 prior to
entering the V-tank. In some implementations, three separate hot side storage
tanks at
respective temperatures T1, t and TijE may be used (e.g., an extra tank may be
used in addition
to the tanks 6 and 7). During heat exchange in the hot side CFX 2 in the
discharge cycle, heat
from the working fluid exiting the recuperator 5 may be transferred to the HTS
medium in the
Ti -tank. The HTS medium may be cooled in/by, for example, the heat rejection
device 56 prior
to entering the T1-tank. Heat rejection to the environment in such a manner
may present several
advantages. In a first example, it may eliminate the need for a potentially
expensive working
fluid to ambient beat exchanger that is capable of absorbing heat from the
working fluid at a
rate proportional to the power input/output of the system. The HTS and CTS
media may instead
reject heat over extended time periods, thus reducing the cost of the cooling
infrastructure. In
a second example, it may allow the decision regarding when heat is rejected to
the environment
to be delayed such that heat exchange to ambient may be performed when
temperature (e.g.,
the ambient temperature) is most favorable.
[0139] In the charge and discharge cycles of FIGs. 13 and 17, and FIGs. 14 and
18,
respectively, the same compression ratios and temperature values are used for
both charge and
discharge. In this configuration, the roundtrip efficiency can be about n
store ¨ 74%, as given by
To = 194 K (-79 C), Tt = 494 K (221 C). = 0.95, 17, = 0.9 and r = 3.3.
[0140] Thus, in some examples involving working fluid to working fluid
recuperation, heat
rejection on the hot side (high pressure) side of the closed charge cycle can
take place in three
operations (heat exchange with the HTS medium, followed by recuperation,
followed by heat
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rejection to the environment), and heat rejection on the cold side (low
pressure) side of the
closed discharge cycle can take place in three operations (heat rejection to
the environment,
followed by recuperation, followed by heat exchange with the CTS medium). As a
result of
recuperation, the higher temperature HTS tank(s) 6 can remain at T1+ while the
lower
temperature HTS tank(s) 7 can now be at the temperature T1 >To+, and the lower
temperature
CTS tank(s) 9 can remain at To while the higher temperature CTS tank(s) 8 can
now be at the
temperature To < Ti.
[0141] In some cases, recuperation may be implemented using the heat exchanger
5 for direct
transfer of heat between the working fluid on the high pressure side and the
working fluid on
the low pressure side. In an alternative configuration, an additional pair (or
plurality) of heat
exchangers together with an additional heat transfer medium or fluid (e.g., a
dedicated thermal
heat transfer fluid that is liquid in an appropriate temperature range, such
as, for example,
Thermino10) may be used to achieve recuperation. For example, an additional
heat exchanger
may be added in series with the cold side heat exchanger and an additional
heat exchanger may
be added in series with the hot side heat exchanger. The additional heat
transfer medium may
circulate between the two additional heat exchangers in a closed loop. In
other examples, one
or more additional heat exchangers may be placed elsewhere in the system to
facilitate
recuperation. Further, one or more additional heat transfer media or mixtures
thereof may be
used. The one or more additional heat transfer media fluids may be in fluid or
thermal
communication with one or more other components, such as, for example, a
cooling tower or
a radiator.
[0142] In one example, hexane or heptane can be used as a CTS medium, and
nitrate salt can
be used as an HTS medium. On the low pressure side of the cycle, the operating
temperatures
of the pumped thermal storage cycles may be limited by the melting point of
hexane (178 K or
-95 C) at To and by the melting point of the nitrate (494 K or 221 C) at Ti.
On the high pressure
side of the cycle, the operating temperatures may be limited by the boiling
point of hexane (341
K or 68 C) at To+and by the decomposition of nitrate (873 K or 600 C) at Ti+.
At these
conditions, the high pressure and low pressure temperature ranges can overlap
such that
recuperation can be implemented. The actual temperatures To, Ti, To and Ti+and
pressure
ratios implemented in hexane/nitrate systems may differ from the limits above.
[0143] In some examples, recuperation may enable the compression ratio to be
reduced. In
some cases, reducing the compression ratio may result in reduced compressor
and turbine
losses. In some cases, the compression ratio may be at least about 1.2, at
least about 1.5, at
least about 2, at least about 2.5, at least about 3, at least about 3.5, at
least about 4, at least
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about 4.5, at least about 5, at least about 6, at least about 8, at least
about 10, at least about 15,
at least about 20, at least about 30, or more.
[0144] In some cases, To may be at least about 30 K, at least about 50 K, at
least about 80 K,
at least about 100 K, at least about 120 K, at least about 140 K, at least
about 160 K, at least
about 180 K, at least about 200 K, at least about 220 K, at least about 240 K,
at least about
260 K, or at least about 280 K. In some cases, To+ may be at least about 220
K, at least about
240 K, at least about 260 K, at least about 280 K, at least about 300 K, at
least about 320 K, at
least about 340 K, at least about 360 K, at least about 380 K, at least about
400 K. or more. In
some cases, the temperatures To and To+ can be constrained by the ability to
reject excess heat
to the environment at ambient temperature. In some cases. the temperatures To
and 77 can be
constrained by the operating temperatures of the CTS (e.g., a phase transition
temperature). In
some cases, the temperatures To and T can be constrained by the compression
ratio being
used. Any description of the temperatures To and/or To+ herein may apply to
any system or
method of the disclosure.
[0145] In sonic cases, 7'1 may be at least about 350K, at least about 400 K,
at least about 440
K, at least about 480 K, at least about 520 K, at least about 560 K. at least
about 600 K, at least
about 640 K, at least about 680 K, at least about 720 K, at least about 760 K,
at least about
800 K, at least about 840 K, at least about 880 K, at least about 920 K, at
least about 960 K, at
least about 1000 K, at least about 1100 K, at least about 1200 K, at least
about 1300 K, at least
about 1400 K, or more. In some cases, Ti+ may be at least about 480 K, at
least about 520 K,
at least about 560 K, at least about 600 K, at least about 640 K, at least
about 680 K, at least
about 720 K, at least about 760 K, at least about 800 K, at least about 840 K,
at least about
880 K, at least about 920 K, at least about 960 K, at least about 1000 K, at
least about 1100 K,
at least about 1200 K, at least about 1300 K, at least about 1400 K, at least
about 1500 K, at
least about 1600 K, at least about 1700 K, or more. In some cases. the
temperatures T1 and Ti+
can be constrained by the operating temperatures of the HTS. In some cases.
the
temperatures T1 and TiE can be constrained by the thermal limits of the metals
and materials
being used in the system. For example, a conventional solar salt can have a
recommended
temperature range of approximately 560-840 K. Various system improvements,
such as, for
example, increased roundtrip efficiency, increased power and increased storage
capacity may
be realized as available materials, metallurgy and storage materials improve
over time and
enable different temperature ranges to be achieved. Any description of the
temperatures Ti
and/or Ti+ herein may apply to any system or method of the disclosure.
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[0146] In some cases, the roundtrip efficiency ristor, (e.g., electricity
storage efficiency) with
and/or without recuperation can be at least about 5%, at least about 10%, at
least about 15%,
at least about 20%, at least about 25%, at least about 30%, at least about
35%, at least about
40%, at least about 45%, at least about 50%, at least about 55%, at least
about 60%, at least
about 65%, at least about 70%, at least about 75%, at least about 80%, at
least about 85%, at
least about 90%, or at least about 95%.
[0147] In some implementations, at least a portion of heat transfer in the
system (e.g., heat
transfer to and from the working fluid) during a charge and/or discharge cycle
includes heat
transfer with the environment (e.g., heat transfer in sections 38 and 39). The
remainder of the
heat transfer in the system can occur through thermal communication with
thermal storage
media (e.g., thermal storage media 21 and 22), through heat transfer in the
recuperator 5 and/or
through various heat transfer processes within system boundaries (i.e., not
with the surrounding
environment). In some examples, the environment may refer to gaseous or liquid
reservoirs
surrounding the system (e.g., air, water), any system or medium capable of
exchanging thermal
energy with the system (e.g., another thermodynamic cycle or system,
heating/cooling systems,
etc.), or any combination thereof. In some examples, heat transferred through
thermal
communication with the heat storage media can be at least about 25%, at least
about 50%, at
least about 60%, at least about 70%. at least about 80%, or at least about 90%
of all heat
transferred in the system. In some examples, heat transferred through heat
transfer in the
recuperator can be at least about 5%, at least about 10%, at least about 15%,
at least about 20%,
at least about 25%, at least about 50%, or at least about 75% of all heat
transferred in the
system. In some examples, heat transferred through thermal communication with
the heat
storage media and through heat transfer in the recuperator can be at least
about 25%, at least
about 50%, at least about 60%, at least about 70%, at least about 80%, at
least about 90%, or
even about 100% of all heat transferred in the system. In some examples, heat
transferred
through heat transfer with the environment can be less than about 5%, less
than about 10%,
less than about 15%, less than about 20%, less than about 30%, less than about
40%, less than
about 50%, less than about 60%, less than about 70%, less than about 80%, less
than about
90%, less than about 100%, or even 100% of all heat transferred in the systcm.
In some
implementations, all heat transfer in the system may be with the thermal
storage media (e.g.,
the CTS and HTS media), and only the thermal storage media may conduct heat
transfer with
the environment.
[0148] Pumped thermal cycles of the disclosure (e.g., the cycles in FIGs. 13
and 14) may be
implemented through various configurations of pipes and valves for
transporting the working
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fluid between the turbomachinery and the heat exchangers. In some
implementations, a valving
system may be used such that the different cycles of the system can be
interchanged while
maintaining the same or nearly the same temperature profile across at least
one, across a subset
or across all of counter-flow heat exchangers in the system. For example, the
valving may be
configured such that the working fluid can pass through the heat exchangers in
opposite flow
directions on charge and discharge and flow directions of the HTS and CTS
media arc reversed
by reversing the direction of the pumps.
[0149] In some implementations, the system with a recuperator may have a
different
compression and/or expansion ratio on charge and discharge. This may then
involve heat
rejection at only one or both of the heat rejection locations 38 and 39 as
shown in Figure 5C
along the lines described above.
[0150] FIG. 19 is a schematic flow diagram of hot side recharging in a pumped
heat cycle in
solar mode with heating of a solar salt solely by solar power. The system can
comprise a solar
heater for heating the hot side heat storage. The HTS medium 21 in the second
hot thermal
storage tank 7 of a discharge cycle, such as, for example, the HTS medium of
the discharge
cycle in FIG. 14, can be recharged within element 17 using heating provided by
solar radiation.
The HTS medium (e.g., molten salt) can be heated by solar heating from the
temperature T1 in
the second hot thermal storage tank 7 to the temperature T1+ in the first hot
thermal storage
tank 6.
[0151] In some implementations, such as, for example, for the systems in FIGs.
19 solar heat
for heating the HTS medium (e.g., from Ti = 493 K (220 C) to T1+ = 873 K (600
C)) may be
provided by a concentrating solar facility. In some examples, a small scale
concentrating
facility may be utilized for providing heat. In some cases, the concentrating
solar facility may
include one or more components for achieving high solar concentrating
efficiency, including,
for example, high-performance actuators (e.g., adaptive fluidic actuators
manufactured from
polymers), mutiplexing control system, dense heliostat layout etc. In some
examples, the heat
provided for heating the HTS medium (e.g., in the element 17) may be a waste
heat stream
from the concentrating solar facility.
[0152] FIG. 20 is a schematic flow diagram of a pumped thermal system
discharge cycle that
can be coupled with external heat input (e.g., solar, combustion) with heat
rejection to ambient.
Such a discharge cycle may be used, for example, in situations where the
capacity for hot side
recharging (e.g., using solar heating, waste heat or combustion) is greater
than the capacity for
cold side recharging. Solar heat may be used to charge the HTS medium 21 in
the hot side
storage tanks from T1 to T,+, as described elsewhere herein. The discharge
cycle can operate
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similarly to the discharge cycle in FIG. 3, but after exiting the turbine 3,
the working fluid 20
can proceed to the cold side CFX 4 heat exchanger 4 where it exchanges heat
with an
intermediate thermal storage (ITS) medium 61 having a lower temperature To at
or near
ambient temperature. The ITS medium 61 enters the cold side CFX 4 from a
second
intermediate thermal storage tank 59 at the temperature To (e.g., ambient
temperature) and exits
the cold side CFX 4 into a first intermediate thermal storage tank 60 at the
temperature 71 ,
while the working fluid 20 enters the cold side CFX 4 at the temperature Di
and exits the cold
side CFX 4 at the temperature To. The working fluid enters the compressor 1 at
To and P2, exits
the compressor at To and Pi, absorbs heat Qi from the HTS medium 21 in the hot
side CFX 2,
exits the hot side CFX 2 at Ti+ and P1, enters the turbine 3 at Ti+ and P1,
exits the turbine at
and P2, rejects heat Q2 from the ITS medium 61 in the cold side CFX 4, and
exits the cold
side CFX 4 at To and P2, returning to its initial state before entering the
compressor.
[0153] In some implementations, the ITS medium 61 may be a liquid over the
entire range
from To to T1. In other implementations, the ITS medium 61 may not be a liquid
over the entire
range from To to D1, but may be provided to the counter-flow heat exchanger 4
at a higher flow
rate in order to achieve a lower temperature rise across the counter-flow heat
exchanger (e.g.,
such that the temperature of the ITS medium at the exit of the counter-flow
heat exchanger 4
is lower than Dip while still cooling the working fluid from T1 to To. In this
instance, the
temperature of the ITS medium in the tank 60 can be lower than T1. The ITS
medium in the
tank 60 can exchange heat with ambient (e.g., through a radiator or other
implementations
described herein) in order to cool back to the temperature To. In some cases,
the ITS medium
can then be returned to the tank 59. In some cases, the heat deposited in the
ITS medium may
be used for various useful purposes such as, for example, residential or
commercial heating,
thermal desalination or other uses described elsewhere herein.
[0154] FIG. 21 is a schematic flow diagram of a pumped thermal system
discharge cycle in
solar mode or combustion heated mode with heat rejection to an intermediate
fluid circulated
in a thermal bath at ambient temperature. The discharge cycle can operate
similarly to the
discharge cycle in FIG. 20, but after exiting the turbine 3, the working fluid
20 can proceed to
the cold side CFX 4 where it exchanges heat with an intermediate medium or
fluid 62
circulating through a thermal bath 63 at the temperature To at or near ambient
temperature. The
intermediate medium or fluid 62 (e.g., Therminol , Or a heat transfer oil) may
be used for
exchanging heat between the working fluid 20 and a thermal bath 63 in the cold
side CFX 4.
The use of the intermediate fluid 62 may provide an advantage over contacting
an inexpensive
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thermal sink or medium (e.g., water) directly with the working fluid. For
example, directly
contacting such a thermal medium with the working fluid in the cold side CFX 4
may cause
problems, such as, for example, evaporation or over-pressurization (e.g.,
explosion) of the
thermal medium. The intermediate fluid 62 can remain in liquid phase
throughout all, at least
a portion of, or a significant portion of the operation in the cold side CFX
4. As the intermediate
fluid 62 passes through the thermal bath 58, it can be sufficiently cooled to
circulate back into
the cold side CFX 4 for cooling the working fluid from 71 to To. The thermal
bath 63 may
contain a large amount of inexpensive heat sink material or medium, such as,
for example,
water. In some cases, the heat deposited in the heat sink material may be used
for various useful
purposes such as, for example, residential or commercial heating, thermal
desalination or other
uses described elsewhere herein. In some cases, the heat sink material may be
re-equilibrated
with ambient temperature (e.g., through a radiator or other implementations
described herein).
[0155] In some implementations, the discharge cycles in FIGs. 20 and/or 21 may
include a
recuperator, as described in greater detail in examples throughout the
disclosure. Such systems
may be implemented using the temperatures T1+, T1, To+ and To described in
greater detail
elsewhere herein.
Solar assisted pumped thermal storage cycles with intercooling
[0156] In some instances, the pumped thermal system may provide heat sources
and/or cold
sources to other facilities or systems such as, for example, through co-
location with a gas to
liquids (GTL) facility or a desalination facility. In one example, the GTL
facilities may make
use of one or more of the cold reservoirs in the system (e.g., the CTS medium
in the tank 9 for
use in oxygen separation in the GTL facility) and/or one or more hot
reservoirs in the system
(e.g., the HTS medium in the tank 6 for use in a Fischer-Tropsch process in
the GTL facility).
In another example, one or more hot reservoirs or one or more cold reservoirs
in the pumped
thermal system may be used for the operation of thermal desalination methods.
Further
examples of possible heat and cold uses include co-location or heat exchange
with
building/area heating and cooling systems.
[0157] Conversely, in some cases, the pumped thermal system may make use of
waste heat
sources and/or waste cold sources from other facilities or systems such as,
for example, through
co-location with a liquefied natural gas import or export terminal. For
example, a waste cold
source may be used for cooling the cold side thermal storage media 22. In some
implementations, recharging of the cold side using waste cold may be combined
with
recharging of the hot side thermal storage media 21 by external heat input
(e.g., solar,
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combustion, waste heat, etc.). In some cases, the recharged storage media can
then be used in
a discharge cycle such as, for example, the discharge cycles in FIGs. 14 or
16. In some cases,
the pumped thermal system may be used as a heat engine with a waste heat
source serving as
the hot side heat input and a waste cold source serving as the cold side heat
sink. In another
implementation, the hot side storage media may he recharged using a modified
version of the
cycle shown in FIG. 15, where the temperature To is about the ambient
temperature and Do+
corresponds to a temperature above the ambient temperature. In some examples,
a waste heat
source can be used to provide the heat needed at a temperature of at least t
for heating the
working fluid and/or the CTS medium to Dot In another implementation, an
intermediate fluid
(e.g., Thernlinol ) which can remain liquid between the temperatures T6' and
To may be used
to transfer the heat from the waste heat source to the working fluid.
Pumped thermal systems with dedicated compressor/turbine pairs
[0158] In a further aspect of the disclosure, pumped thermal systems
comprising multiple
working fluid systems, or working fluid flow paths are provided. In some
cases, pumped
thermal system components in the charge and discharge modes may be the same.
For example,
the same compressor/turbine pair may be used in charge and discharge cycles.
Alternatively,
one or more system components may differ between charge and discharge modes.
For example,
separate compressor/turbine pairs may be used in charge and discharge cycles.
In one
implementation, the system has one set of heat exchangers, and a common set of
HTS and CTS
tanks which are charged or discharged by two pairs or sets of compressors and
turbines. In
another implementation, the system has a common set of HTS and CTS tanks, but
separate sets
of heat exchangers and separate sets of compressors and turbines.
[0159] Pumped thermal systems with recuperation, utilization of external
sources of heat, cold
and/or waste heat/cold may benefit from having separate compressor/turbine
pairs as a result
of operation of turbomachinery over large and/or different temperature ranges
in charge and
discharge modes. For example, temperature changes between charge and discharge
cycles may
lead to a thermal adjustment period or other difficulties during transition
between the cycles
(e.g., issues or factor related to metallurgy, thermal expansion, Reynolds
number, temperature
dependent compression ratios, tip clearance and/or bearing friction etc.). In
another example,
turbomachinery (e.g., turbomachinery used in systems with recuperation) may
operate over a
relatively low pressure ratio (e.g., with relatively few compression stages)
but over relatively
large temperature during both compression and expansion. The temperature
ranges may change
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(e.g., switch as in FIGs. 17 and 18) between charge and discharge modes. In
some cases, the
operation over large temperature ranges during compression and/or expansion
may complicate
design of a combined compressor/turbine for both charge and discharge.
Furthermore,
recuperation, waste heat/cold incorporation and/or other pumped thermal system
features may
reduce the compression ratio of the compressor/turbine in the charge cycle
and/or the discharge
cycle, thereby reducing the cost associated with duplicating
compressor/turbine sets.
[0160] FIGs. 22 and 23 show pumped thermal systems with separate compressor
1/turbine 3
pairs for charge mode C and discharge mode D. The separate compressor/turbine
pairs may or
may not be ganged on a common mechanical shaft. In this example, the
compressor/turbine
pairs C and D can have separate shafts 10. The shafts 10 may rotate at the
same speed or at
different speeds. The separate compressor/turbine pairs or working fluid
systems may or may
not share heat exchangers (e.g., the heat exchangers 2 and 4).
[0161] In the example in FIG. 22, the system has a common set of HTS tanks 6
and 7 and CTS
tanks 8 and 9. The system has separate pairs of heat exchangers 2 and 4 and
separate
compressor 1/turbine 3 pairs for the charge mode C and the discharge mode D.
The HTS and
CTS storage media flow paths for the charging cycle are shown as solid black
lines. The HTS
and CTS storage media flow paths for the discharge cycle are shown as the
dashed grey lines.
[0162] In the example in FIG. 23, the system, shown in a charge configuration,
has one set of
heat exchangers 2 and 4, and a common set of HTS tanks 6 and 7 and CTS tanks 8
and 9. The
HTS and CTS tanks can be charged by a compressor/turbine set C, or discharged
by a
compressor/turbine set D, each set comprising a compressor 1 and a turbine 3.
The system may
switch between the sets C and D using valves 83. In the example in FIG. 22,
the system, again
shown in a charge configuration, has a common set of HTS tanks 6 and 7 and CTS
tanks 8 and
9. The HTS and CTS tanks can be charged by the charge set C that includes a
first set of the
heat exchangers 2 and 4, the compressor 1 and the turbine 3. The HTS and CTS
tanks can be
discharged by switching to a separate discharge set C that includes a second
set of the heat
exchangers 2 and 4, the compressor 1 and the turbine 3.
[0163] In one example, if the charge and discharge sets of compressors and
turbines in FIGs.
22 and 23 are not operated at the same time, the charge and discharge sets may
share a common
set of heat exchangers that are switched between the turbomachinery pairs
using the valves 83.
In another example, if the charge and discharge turbomachinery sets or pairs
in FIGs. 22 and
23 are operated at the same time (e.g., in order for one set to charge,
following intermittent
generation, and the other set to discharge at the same time, following load),
then each set of
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turbomachinery may have a dedicated set of heat exchangers. In this instance,
the charge and
discharge sets may or may not share a set of HTS and CTS tanks.
[0164] In some implementations, separate compressor/turbine sets or pairs may
advantageously be used in pumped thermal systems used with intermittent and/or
variable
electric power inputs. For example, a first compressor/turbine set can be used
in a charge cycle
that follows wind and/or solar powcr (e.g., electric power input from wind
and/or solar power
systems) while a second compressor/turbine set can be used in a discharge
cycle that follows
load (e.g., electric power output to a power grid). In this configuration,
pumped thermal
systems placed between a power generation system and a load may aid in
smoothing
variations/fluctuations in input and/or output power requirements.
Hybrid pumped thermal systems
[0165] In accordance with another aspect of the disclosure, pumped thermal
systems can be
augmented by additional energy conversion processes and/or be directly
utilized as energy
conversion systems without energy storage (i.e., as power generation systems).
In some
examples, pumped thermal systems herein can be modified to allow for direct
power generation
using natural gas, Diesel fuel, petroleum gas (e.g., propane/butane), dimethyl
ether, fuel oil,
wood chips, landfill gas, hexane, hydrocarbons or any other combustible
substance (e.g., fossil
fuel or biomass) for adding heat to the working fluid on a hot side of a
working fluid cycle, and
a cold side heat sink (e.g., water) for removing heat from the working fluid
on a cold side of
the working fluid cycle.
[0166] FIGs. 24 and 25 show pumped thermal systems configured in generation
mode. In some
examples, pumped thermal systems herein can be modified by adding two
additional heat
exchangers 40 and 41, four additional valves 19a, 19b, 19c and 19d, a heat
sink (e.g., a water
cooling system; water from a fresh water reservoir such as a river, a lake or
a reservoir; salt
water from a salt water reservoir such as a sea or an ocean; air cooling using
radiators,
fans/blowers, convection; or an environmental heat sink such as ground/soil,
cold air etc.) 42,
and a heat source (e.g., a combustion chamber with a fuel-oxidant mixture) 43.
The heat source
43 can exchange heat with a first of the two additional heat exchangers 40,
and the heat sink
42 can exchange heat with a second of the two additional heat exchangers 41.
The heat source
43 may be used to for exchanging heat with the working fluid 20.
[0167] The heat source 43 may be a combustion heat source. In some examples,
the
combustion heat source can comprise a combustion chamber for combusting a
combustible
substance (e.g., a fossil fuel, a synthetic fuel, municipal solid waste (MSW)
or biomass). In
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some cases, the combustion chamber may be separate from the heat exchanger 40.
In some
cases, the heat exchanger 40 may comprise the combustion chamber. The heat
source 43 may
be a waste heat source, such as, for example waste heat from a power plant, an
industrial
process (e.g., furnace exhaust).
[0168] In some examples, a solar heater, a combustion heat source, a waste
heat source, or any
combination thereof may be used for heating the hot side heat storage fluid
and/or the working
fluid. In an example, the working fluid can be heated directly using any of
these heat sources.
In another example, the hot side heat storage fluid (or HTS medium) can be
heated using any
of these heat sources. In another example, the hot side heat storage fluid (or
HTS medium) can
be heated in parallel with the working fluid using any of these heat sources.
[0169] The pumped thermal systems in FIGs. 24 and 25 may be operated as hybrid
systems.
For example, the valves 19a, 19b, 19c and 19d can be used to switch between
two modes. When
the valves are in a first position, the system can operate as a pumped thermal
storage system
(e.g., closed system in charge/discharge mode). In this configuration, the
working fluid 20 (e.g.,
argon or air) can exchange heat with an HTS medium (e.g., molten salt) in the
hot side heat
exchanger 2 and with a CTS medium (e.g., hexane) in the cold side heat
exchanger 4. When
the valves are in a second position, the system can operate as a power
generation system (e.g.,
open system in generation mode). In this configuration, the heat exchangers 2
and 4 may be
bypassed, and the working fluid 20 can exchange heat with the combustion
chamber 43 in the
hot side heat exchanger 40 and with the heat sink 42 in the cold side heat
exchanger 41. Any
description of configuration and/or design of heat transfer processes (e.g.,
heat transfer in heat
exchangers) described herein in relation to pumped thermal systems may also be
applied to
hybrid pumped thermal systems, and vice versa. For example, the heat sink 42,
the heat source
43, the heat exchangers 40 and 41, and/or the quantity of heat transferred on
the cold side and/or
the hot side may be configured to decrease or minimize entropy generation
associated with heat
transfer processes and/or to maximize system efficiency.
[0170] In some implementations, the hybrid systems may operate in storage and
generation
modes simultaneously. For example, the valves 19a, 19b, 19c and 19d can be
configured to
allow a given split between a working fluid flow rate to the heat exchangers
40 and 41 and a
working fluid flow rate to the heat exchangers 2 and 4. Alternatively, the
hybrid systems may
operate exclusively in storage mode, or exclusively in generation mode (e.g.,
as a natural gas
peaking plant). In some cases, the split between modes may be selected based
on system
efficiency, available electric power input (e.g., based on availability),
desired electric power
output (e.g., based on load demand) etc. For example, thermal efficiency of an
ideal system
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(i.e., assuming isentropic compression and expansion processes, ideal heat
transfer processes)
operating exclusively in generation mode can be the thermal efficiency of a
working fluid
undergoing an ideal Brayton cycle. In some examples, thermal efficiencies in
hybrid systems
of the disclosure (e.g., exclusive and/or split mode operation) can heat least
about 10%, at least
about 20%, at least about 30%, at least about 40%, at least about 50%, at
least about 60%, or
more.
[0171] The heat source 43 may be used for exchanging heat with an HTS medium
(e.g., a
molten salt). For example, the combustion heat source 43 may be used for
heating the HTS
medium 21. In some instances, rather than using the combustion heat source 43
for exchanging
heat in the heat exchanger 40 or for directly exchanging heat between flue
gases from the
combustion heat source and the working fluid, the combustion heat source 43
may be used to
heat up the HTS medium 21 between the two HTS tanks 7 and 6.
[0172] FIG. 26 is a schematic flow diagram of hot side recharging in a pumped
heat cycle
through heating by the heat source 43 (e.g., combustion heat source, waste
heat source). In an
example, the heat source 43 is a waste heat source, such as a waste heat
source from a refinery
or other processing plant. In an example, the heat source 43 is obtained from
combusting
natural gas in order to ensure the delivery of electricity even if the pumped
thermal system runs
out of charged storage media. For example, recharging of the hot side storage
media using the
heat source 43 may provide an advantage over recharging using electricity or
other means (e.g.,
the price of electricity at the time may be too high). The heat source 43 can
be used to heat up
the HTS medium 21 from the temperature Ti in the tank 7 to the temperature Ti+
in the tank 6.
The HTS medium can then be used in the CFX 2 for exchanging heat with the
working fluid
to in a discharge cycle, such as, for example, the discharge cycles in FIGs.
20 and 21.
[0173] In some examples, such as, for example, when the CTS medium is a
combustible
substance such as a fossil fuel (e.g., hexane or heptanes), burning of the CTS
medium stored
in the CTS tanks (e.g., the tanks 8 and 9) may be used for providing thermal
energy for heating
the HTS medium as shown, for example, in FIG. 26 or for operation of the
cycles in the
configurations shown, for example, in FIGs. 24 and 25.
[0174] The systems of the disclosure may be capable of operating both in an
electricity only
storage cycle (comprising heat transfer with an HTS medium and a CTS medium
below
ambient temperature) and in a heat engine to ambient cycle, where, in a
discharge mode, heat
is input from the HTS medium to the working fluid and rejected to the ambient
environment
rather than to the CTS medium. This capability may enable the use of heating
of the HTS with
combustible substances (e.g., as shown in FIG. 26) or the use of solar heating
of the HTS (e.g.,
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as shown in FIG. 19). Heat rejection to ambient may be implemented using, for
example, the
discharge cycles in FIGs. 20 and 21. In some cases, heat may be rejected to
the environment
with the aid of the ITS medium 61 or the intermediate medium 62.
[0175] Aspects of the disclosure may be synergistically combined. For example,
the systems
capable of operating both in an electricity only storage cycle and in a heat
engine to ambient
cycle may comprise a recuperator. Any description in relation to such hybrid
systems without
a recuperator may readily be applied to hybrid systems with a recuperator at
least in some
configurations. In some instances, the hybrid systems may be implemented
using, for example,
the parallel, valved configuration in FIG. 24. For example, the counter-flow
heat exchangers 4
in FIGs. 20 and 21 may be implemented as separate counter-flow heat exchangers
67 for
exchanging heat with the ambient environment, and may be used in combination
with cold side
counter-flow heat exchangers 4 of the disclosure.
[0176] In some implementations, the systems herein may be configured to enable
switching
between different cycles of the disclosure using a shared set of valves and
pipes. For example,
the system may be configured to switch between the electricity only charge
cycle (such as
shown in, for example, FIG. 15), the electricity only discharge cycle (such as
shown in, for
example, FIG. 16), and the heat engine to ambient cycle (such as shown in FIG.
21).
Pumped thermal systems with pressure regulation power control
[0177] In an aspect of the disclosure, the pressure of working fluids in
pumped thermal systems
can be controlled to achieve power control. In an example, the power provided
to a closed
system in charge mode and/or the power extracted from the closed system in
discharge and/or
generation mode (e.g., work input/output using the shaft 10) is proportional
to the molar or
mass flow rate of the circulating working fluid. The mass flow rate is
proportional to density,
area, and flow velocity. The flow velocity can be kept fixed in order to
achieve a fixed shaft
speed (e.g., 3600 rpm or 3000 rpm in accordance with power grid requirements
of 60 and 50
Hz respectively). Thus, as the pressure of the working fluid changes, the mass
flow rate and
the power can change. In an example, as the mass flow rate increases in a
discharge and/or
generation mode, more load should be added to the system to maintain a
constant speed of the
rotating shaft, and vice versa. In another example, if load is reduced during
operation in a
discharge and/or generation mode, the reduced load can cause the shaft speed
to increase, thus
increasing the mass flow rate. For some period of time, before the heat stored
in the thermal
capacity of the heat exchangers themselves is dissipated, this increased mass
flow rate can lead
to an increase in the power delivered, in turn increasing the shaft speed. The
shaft speed and
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the power can continue to increase uncontrollably, resulting in a runaway of
the rotating shaft.
In some examples, pressure regulation may enable control, and thus
stabilization of runaway,
through adjustment of the amount (e.g., density) of circulating working fluid
in accordance
with system requirements. In an example where shaft speed (and turbomachinery,
such as a
turbine, attached to the shaft) begins to run away, a controller can reduce
the mass of circulating
working fluid (e.g., mass flow rate) in order to decrease the power delivered,
in turn decreasing
the shaft speed. Pressure regulation may also allow for an increase in mass
flow rate in
response to an increase in load. In each of these instances, the flow rates of
the HTS and CTS
media through the heat exchangers can be matched to the heat capacity of the
working fluid
passing through the heat exchangers.
[0178] In some examples, the working fluid pressure in the closed system can
be varied by
using an auxiliary working fluid tank in fluid communication with the closed
system. In this
configuration, power input/output can be decreased by transferring the working
fluid from the
closed cycle loop to the tank, and power input/output can be increased by
transferring the
working fluid from the tank to the closed cycle loop. In an example, when the
pressure of the
working fluid is decreased, less heat can be transferred between the thermal
storage tanks on
the hot and cold sides of the system as a result of the decreased mass flow
rate and less power
can be input to/output by the system.
[0179] As the pressure of the working fluid is varied, the compression ratios
of turbomachinery
components may remain substantially unchanged. In some cases, one or more
operating
parameters and/or configuration (e.g., variable stators, shaft speed) of
turbomachinery
components can be adjusted in response to a change in working fluid pressure
(e.g., to achieve
a desired performance of the system). Alternatively, one or more pressure
ratios may change
in response to a change in working fluid pressure.
[0180] In some cases, reduced cost and/or reduced parasitic energy consumption
may be
achieved using the power control configuration relative to other
configurations (e.g., using a
choke valve for controlling the flow of the working fluid). In some examples,
variation of
working fluid pressure while keeping the temperature and flow velocity
constant (or near-
constant) may lead to negligible entropy generation. In some examples, an
increase or decrease
in system pressure may lead to changes in, for example, turbomachinery
efficiencies.
[0181] FIG. 27 shows an example of a pumped thermal system with power control.
The
temperature of the working fluid on the hot and cold sides of the system may
remain constant
or near-constant for a given period of time regardless of working fluid mass
flow rate due to
large heat capacities of the heat exchangers 2 and 4 and/or the hot and cold
side thermal storage
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media in the tanks 6, 7, 8 and 9. In some examples, the flow rates of the HTS
and CTS media
through the heat exchangers 2 and 4 are varied in concert with a change in the
pressure of the
working fluid in order to keep the temperatures in the heat exchangers and
working fluid
optimized over longer time periods. Thus, pressure can be used to vary the
mass flow rate in
the system. One or more auxiliary tanks 44 filled with the working fluid 20
(e.g., air, argon or
argon-helium mix) can be in fluid communication with a hot (e.g., high
pressure) side of the
pumped thermal system and/or a cold (e.g., low pressure) side of the pumped
thermal system.
In some examples, the auxiliary tank can be in fluid communication with the
working fluid
adjacent to an inlet of the compressor 1 and/or adjacent to an outlet of the
compressor 1. In
some examples, the auxiliary tank can be in fluid communication with the
working fluid
adjacent to an inlet of the turbine 3 and/or adjacent to an outlet of the
turbine 3. In further
examples, the auxiliary tank can be in fluid communication with the working
fluid in one or
more locations system (e.g., one or more locations on the high pressure side
of the system, on
the low pressure side of the system, or any combination thereof). For example,
the auxiliary
tank can be in fluid communication with the working fluid on a high pressure
side and a low
pressure side of the closed cycle. In some cases, the fluid communication on
the high pressure
side can be provided after the compressor and before the turbine. In some
cases, the fluid
communication on the low pressure side can be provided after the turbine and
before the
compressor. In some instances, the auxiliary tank can contain working fluid at
a pressure
intermediate to the high and low pressures of the system. The working fluid in
the auxiliary
tank can be used to increase or decrease the amount of working fluid 20
circulating in the closed
cycle of the pumped thermal system. The amount of working fluid circulating in
the closed
cycle loop can be decreased by bleeding the working fluid from the high
pressure side of the
closed cycle loop into the tank through a fluid path containing a valve or
mass flow controller
46, thereby charging the tank 44. The amount of working fluid circulating in
the closed cycle
loop can be increased by bleeding the working fluid from the tank into the low
pressure side of
the closed cycle loop through a fluid path containing a valve or mass flow
controller 45, thereby
discharging the tank 44.
[0182] Power control over longer timescalcs may be implemented by changing the
pressure of
the working fluid and by adjusting the flow rates of the hot side 21 and cold
side 22 thermal
storage fluids through the heat exchangers 2 and 4, respectively.
[0183] In some examples, flow rates of the thermal storage media 21 and/or 22
may be
controlled (e.g., by a controller) to maintain given heat exchanger inlet and
outlet temperatures.
In some examples, a first controller(s) may be provided for controlling the
flow rates (e.g.,
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mass flow rates) of thermal storage media, and a second controller may be
provided for
controlling the mass flow rate (e.g., by controlling mass, mass flow rate,
pressure etc.) of the
working fluid.
Pumped thermal systems with pressure-encased motor/generator
[0184] In another aspect of the disclosure, pumped thermal systems with a
pressure-encased
motor/generator are provided. The pressure-encased motor/generator may be
provided as an
alternative to configurations where a shaft (also "crankshaft- herein)
penetrates through a
working fluid containment wall (where it can be exposed to one or more
relatively high
pressure differentials) in order to connect to a motor/generator outside the
working fluid
containment wall. In some cases, the shaft may be exposed to pressures and
temperatures of
the working fluid in the low pressure portion of the working fluid cycle, in
the high pressure
portion of the working fluid cycle, or both. In some cases, crankshaft seal(s)
capable of holding
back the pressures which the crankshaft is exposed to inside the working fluid
containment
wall can be difficult to manufacture and/or difficult to maintain. In some
cases, a rotating seal
between high and low pressure environments may be difficult to achieve. Thus,
coupling of the
compressor and turbine to the motor/generator can be challenging. In some
implementations,
the motor/generator can therefore be placed entirely within the low pressure
portion of the
working fluid cycle, such that the exterior pressure vessel or working fluid
containment wall
may not need to be penetrated.
[0185] FIG. 28 shows an example of a pumped thermal system with a pressure
encased
generator 11. The motor/generator is encased within the pressure vessel or
working fluid
containment wall (shown as dashed lines) and only feed-through electric leads
49 penetrate
through the pressure vessel. A thermal insulation wall 48 is added between the
motor/generator
11 and the working fluid in the low pressure portion of the cycle. The
technical requirements
for achieving an adequate seal through the thermal insulation wall can be less
stringent due to
the pressure being the same on both sides of the thermal insulation wall
(e.g., both sides of the
thermal insulation wall can be located in the low pressure portion of the
cycle). In an example,
the low pressure value can be about 10 atm. In some cases, the motor/generator
may be adapted
for operation at elevated surrounding pressures. An additional thermal
insulation wall 50 can
be used to create a seal between the outlet of the compressor 1 and the inlet
of the turbine 3 in
the high pressure portion of the cycle. In some examples, placing the
motor/generator on the
cold side of the pumped thermal systems may be beneficial to the operation of
the
motor/generator (e.g., cooling of a superconducting generator).
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Pumped thermal systems with variable stator pressure ratio control
[0186] A further aspect of the disclosure relates to control of pressure in
working fluid cycles
of pumped thermal systems by using variable stators. In some examples, use of
variable stators
in turbomachinery components can allow pressure ratios in working fluid cycles
to be varied.
The variable compression ratio can be accomplished by having movable stators
in the
turbomachinery.
[0187] In some cases, pumped thermal systems (e.g., the systems in FIGs. 17
and 18) can
operate at the same compression ratio in both the charge and the discharge
cycles. In this
configuration, heat can be rejected (e.g., to the environment) in section 38
in the charge cycle
and in section 39 in the discharge cycle, wherein the heat in section 38 can
be transferred at a
lower temperature than the heat in section 39. In alternative configurations,
the compression
ratio can be varied when switching between the charge cycle and the discharge
cycle. In an
example, variable stators can be added to both the compressor and the turbine,
thus allowing
the compression ratio to be tuned. The ability to vary compression ratio
between charge and
discharge modes may enable heat to be rejected at the lower temperature only
(e.g., heat may
be rejected in section 38 in the charge cycle but not in section 39 in the
discharge cycle). In
some examples, a greater portion (or all) of the heat rejected to the
environment is transferred
at a lower temperature, which may increase the roundtrip efficiency of the
system.
[0188] FIG. 29 is an example of variable stators in a compressor/turbine pair.
The compressor
1 and the turbine 3 can both have variable stators, so that the compression
ratio for each can be
tuned. Such tuning may increase roundtrip efficiency.
[0189] The compressor and/or the turbine can (each) include one or more
compression stages.
For example, the compressor and/or the turbine can have multiple rows of
repeating features
distributed along its circumference. Each compression stage can comprise one
or more rows of
features. The rows may be arranged in a given order. In one example, the
compressor 1 and the
turbine 3 each comprise a sequence of a plurality of inlet guide vanes 51, a
first plurality of
rotors 52, a plurality of stators 53, a second plurality of rotors 52 and a
plurality of outlet guide
vanes 54. Each plurality of features can be arranged in a row along the
circumference of the
compressor/turbine. The configuration (e.g., direction or angle) of the
stators 53 can be varied,
as indicated in FIG. 29.
[0190] The compressor/turbine pair can be matched. In some cases, an outlet
pressure of the
compressor can be about the same as an inlet pressure of the turbine, and an
inlet pressure of
the compressor can be about the same as the outlet pressure of the turbine;
thus, the pressure
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ratio across the turbine can be the same as the pressure ratio across the
compressor. In some
cases, the inlet/outlet pressures and/or the pressure ratios may differ by a
given amount (e.g.,
to account for pressure drop in the system). The use of variable stators on
both the compressor
and the turbine can allow the compressor and the turbine to remain matched as
the compression
ratio is varied. For example, using the variable stators, operation of the
compressor and the
turbine can remain within suitable operating conditions (e.g. within a given
range or at a given
point on their respective operating maps) as the compression ratio is varied.
Operation within
given ranges or at given points on turbomachinery operating maps may allow
turbomachinery
efficiencies (e.g., isentropic efficiencies) and resulting roundtrip storage
efficiency to be
maintained within a desired range. In some implementations, the use of
variable stators can be
combined with other methods for varying the compression ratios (e.g. variable
shaft rotation
speed, bypassing of turbomachinery stages, gears, power electronics, etc.).
Pumped thermal system units comprisin2 pumped thermal system subunits
[01911 A further aspect of the disclosure relates to control of charging and
discharging rate
over a full range from maximum charging/power input to maximum
discharging/power output
by building composite pumped thermal system units comprised of pumped thermal
system
subunits. In some examples, pumped thermal systems may have a minim UM power
input and/or
output (e.g., minimum power input and/or minimum power output) above 0% of
maximum
power input and/or output (e.g., maximum power input and/or maximum power
output),
respectively. In such cases, a single unit by itself may be able to
continuously ramp from the
minimum power input to the maximum power input and from. the minimum power
output to
the maximum power output, but may not be able to continuously ramp from the
minimum
power input to the minimum power output (i.e., from the minimum power input to
zero power
input/output, and from zero power input/output to the minimum power output).
An ability to
continuously ramp from the minimum power input to the minimum power output may
enable
the system to continuously ramp from the maximum power input to the maximum
power
output. For example, if both the output power and the input power may he
turned down all the
way to zero during operation, the system may be able to continuously vary the
power consumed
or supplied across a range from the maximum input (e.g., acting as a load on
the grid) to the
maximum output (e.g., acting as a generator on the grid). Such functionality
may increase (e.g.,
more than double) the continuously rampable range of the pumped thermal
system. Increasing
the continuously rampahle range of the pumped thermal system may be
advantageous, for
example, when continuously rampa.ble power range is used as a metric for
deteiraining the
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value of grid assets. Further, such functionality may enable the systems of
the disclosure to
follow variable load, variable generation, intermittent generation, or any
combination thereof.
[0192] In some implementations, composite pumped thermal system units
comprised of
multiple pumped thermal system subunits may be used. In some cases, each
subunit may have
a minimum power input and/or output above 0%. The continuous ramping of the
power from
the maximum power input to the maximum power output may include combining a
given
quantity of the subunits. For example, a suitable (e.g., sufficiently large)
number of subunits
may be needed to achieve continuous ramping. In some examples, the number of
subunits can
be at least about 2, 5, 10, 20, 30, 40, 50, 100, 200, 500, 750, 1000, and the
like. In some
examples, the number of subunits is 2, 5, 10, 20, 30, 40, 50, 100, 150, 200,
250, 300, 350, 400,
450, 500, 550, 600, 650, 700, 750, 800, 850, 900, 950, 1000 or more. Each
subunit may have
a given power capacity. For example, each subunit can have a power capacity
that is less than
about 0.1%, less than about 0.5%, less than about 1%, less than about 5%, less
than about 10%,
less than about 25%, less than about 50%, or less than about 90% of the total
power capacity
of the composite pumped thermal system. In some cases, different subunits may
have different
power capacities. In some examples, a subunit has a power capacity of about 10
kW, 100 kW,
500 kW, 1 MW, 2 MW, 5 MW, 10 MW, 20 MW, 50 MW, 100 MW, or more. The continuous
ramping of the power from the maximum power input to the maximum power output
may
include controlling each subunit's power input and/or output (e.g., power
input and/or power
output) separately. In some cases, the subunits may be operated in opposing
directions (e.g.,
one or more subunits may operate in power input mode while one or more
subunits may operate
in power output mode). In one example, if each pumped thermal system subunit
can be
continuously ramped between a maximum power input and/or output down to about
50% of
the maximum power input and/or output, respectively, three or more such pumped
thermal
system subunits may be combined into a composite pumped thermal system unit
that can be
continuously ramped from the maximum input power to the maximum output power.
In some
implementations, the composite pumped thermal system may not have a fully
continuous range
between the maximum input power and the maximum output power, but may have an
increased
number of operating points in this range compared to a non-composite system.
Energy storage system units comprising energy storage system subunits
[0193] A further aspect of the disclosure relates to control of charging and
discharging rate
over a full range from maximum charging/power input to maximum
discharging/power output
by building composite energy storage system units comprised of energy storage
system
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subunits. In some examples, energy storage systems may have a minimum power
input and/or
output (e.g., minimum power input and/or minimum power output) above 0% of
maximum
power input andior output (e.g., maximum power input and/or maximum power
output),
respectively. In such cases, a single unit by itself may be able to
continuously ramp from the
minimum power input to the maximum power input and from the minimum power
output to
the maximum power output, but may not be able to continuously ramp from the
minimum
power input to the minimum power output (i.e., from the minimum power input to
zero power
input/output, and from zero power input/output to the minimum power output).
An. ability to
continuously ramp from the minimum power input to the minimum power output may
enable
the system to continuously ramp from the maximum power input to the maximum
power
output. For example, if both the output power and the input power may be
turned down all the
way to zero during operation, the system may be able to continuously vary the
power consumed
or supplied across a range from the maximum input (e.g., acting as a load on
the grid) to the
maximum output (e.g., acting as a generator on the grid). Such functionality
may increase (e.g.,
more than double) the continuously rampable range of the energy storage
system. Increasing
the continuously rarnpable range of the energy storage system may be
advantageous, for
example, when continuously rampable power range is used as a metric for
determining the
value of grid assets. Further, such functionality may enable the systems of
the disclosure to
follow variable load, variable generation, intermittent generation, or any
combination thereof.
[0194] In some implementations, composite energy storage system units
comprised of multiple
energy storage system subunits may be used. In some examples, any energy
storage system
having power input/output characteristics that may benefit from a composite
configuration may
be used. In some examples, systems having power input and/or power output
characteristics
that may benefit from a composite configuration may include various power
storage and/or
generation systems such as, for example, natural gas or combined cycle power
plants, fuel cell
systems, battery systems, compressed air energy storage systems, pumped
hydroelectric
systems, etc. In some cases, each subunit may have a minimum power input
and/or output
above 0%. The continuous ramping of the power from the maximum power input to
the
maximum power output may include combining a given quantity of the subunits.
For example,
a suitable (e.g., sufficiently large) number of subunits may be needed to
achieve continuous
ramping. In some examples, the number of subunits can be at least about 2, 5,
10, 20, 30, 40,
50, 100, 200, 500, 750, 1000, and the like. In some examples, the number of
subunits is 2, 5,
10, 20, 30, 40, 50, 100, 150, 200, 250, 300, 350, 400, 450, 500, 550, 600,
650, 700, 750, 800,
850, 900, 950, 1000 or more. Each subunit may have a given power capacity. For
example,
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each subunit can have a power capacity that is less than about 0.1%, less than
about 0.5%, less
than about 1%, less than about 5%, less than about 10%, less than about 25%,
less than about
50%, or less than about 90% of the total power capacity of the composite
energy storage
system. In some cases, different subunits may have different power capacities.
In some
examples, a subunit has a power capacity of about 10 kW, 100 kW. 500 kW, 1 MW,
2 MW, 5
MW, 10 MW, 20 MW, 50 MW, 100 MW, or more. The continuous ramping of the power
from
the maximum power input to the maximum power output may include controlling
each
subunit' s power input and/or output (e.g., power input and/or power output)
separately. In some
cases, the subunits may be operated in opposing directions (e.g., one or more
subunits may
operate in power input mode while one or more subunits may operate in power
output mode).
In one example, if each energy storage system subunit can be continuously
ramped between a
maximum power input and/or output down to about 50% of the maximum power input
and/or
output, respectively, three or more such energy storage system subunits may be
combined into
a composite energy storage system unit that can be continuously ramped from
the maximum
input power to the maximum output power. In some implementations, the
composite energy
storage system may not have a fully continuous range between the maximum input
power and
the maximum output power, but may have an increased number of operating points
in this
range compared to a non-composite system.
Control systems
[0195] The present disclosure provides computer control systems (or
controllers) that are
programmed to implement methods of the disclosure. FIG. 30 shows a computer
system 1901
(or controller) that is programmed or otherwise configured to regulate various
process
parameters of energy storage and/or retrieval systems disclosed herein. Such
process
parameters can include temperatures, flow rates, pressures and entropy
changes.
[0196] The computer system 1901 includes a central processing unit (CPU, also
"processor"
and "computer processor" herein) 1905, which can be a single core or multi
core processor, or
a plurality of processors for parallel processing. The computer system 1901
also includes
memory or memory location 1910 (e.g., random-access memory, read-only memory,
flash
memory), electronic storage unit 1915 (e.g., hard disk). communication
interface 1920 (e.g.,
network adapter) for communicating with one or more other systems, and
peripheral devices
1925, such as cache, other memory, data storage and/or electronic display
adapters. The
memory 1910, storage unit 1915, interface 1920 and peripheral devices 1925 are
in
communication with the CPU 1905 through a communication bus (solid lines),
such as a
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motherboard. The storage unit 1915 can be a data storage unit (or data
repository) for storing
data. The computer system 1901 can be operatively coupled to a computer
network
("network") 1930 with the aid of the communication interface 1920. The network
1930 can be
the Internet, an internet and/or extranet, or an intranet and/or extranet that
is in communication
with the Internet. The network 1930 in some cases is a telecommunication
and/or data network.
The network 1930 can include one or more computer servers, which can enable
distributed
computing, such as cloud computing. The network 1930, in some cases with the
aid of the
computer system 1901, can implement a peer-to-peer network, which may enable
devices
coupled to the computer system 1901 to behave as a client or a server.
[0197] The computer system 1901 is coupled to an energy storage and/or
retrieval system
1935, which can be as described above or elsewhere herein. The computer system
1901 can
be coupled to various unit operations of the system 1935, such as flow
regulators (e.g., valves),
temperature sensors, pressure sensors, compressor(s), turbine(s), electrical
switches, and
photovoltaic modules. The system 1901 can be directly coupled to, or be a part
of, the system
1935, or be in communication with the system 1935 through the network 1930.
[0198] The CPU 1905 can execute a sequence of machine-readable instructions,
which can be
embodied in a program or software. The instructions may be stored in a memory
location, such
as the memory 1910. Examples of operations performed by the CPU 1905 can
include fetch,
decode, execute, and writeback.
[0199] With continued reference to FIG. 30, the storage unit 1915 can store
files, such as
drivers, libraries and saved programs. The storage unit 1915 can store
programs generated by
users and recorded sessions, as well as output(s) associated with the
programs. The storage
unit 1915 can store user data, e.g., user preferences and user programs. The
computer system
1901 in some cases can include one or more additional data storage units that
are external to
the computer system 1901, such as located on a remote server that is in
communication with
the computer system 1901 through an intranet or the Internet.
[0200] The computer system 1901 can communicate with one or more remote
computer
systems through the network 1930. For instance, the computer system 1901 can
communicate
with a remote computer system of a user (e.g., operator). Examples of remote
computer
systems include personal computers, slate or tablet PC's, telephones, Smart
phones, or personal
digital assistants. The user can access the computer system 1901 via the
network 1930.
[0201] Methods as described herein can be implemented by way of machine (e.g.,
computer
processor) executable code stored on an electronic storage location of the
computer system
1901, such as, for example, on the memory 1910 or electronic storage unit
1915. The machine
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executable or machine readable code can be provided in the form of software.
During use, the
code can be executed by the processor 1905. In some cases, the code can be
retrieved from the
storage unit 1915 and stored on the memory 1910 for ready access by the
processor 1905. In
some situations, the electronic storage unit 1915 can be precluded, and
machine-executable
instructions are stored on memory 1910.
[0202] The code can be pre-compiled and configured for use with a machine have
a processor
adapted to execute the code, or can be compiled during runtime. The code can
be supplied in
a programming language that can be selected to enable the code to execute in a
pre-compiled
or as-compiled fashion.
[0203] Aspects of the systems and methods provided herein, such as the
computer system
1901, can be embodied in programrning. Various aspects of the technology may
be thought of
as "products" or "articles of manufacture" typically in the form of machine
(or processor)
executable code and/or associated data that is carried on or embodied in a
type of machine
readable medium. Machine-executable code can be stored on an electronic
storage unit, such
memory (e.g., read-only memory, random-access memory, flash memory) or a hard
disk.
"Storage" type media can include any or all of the tangible memory of the
computers,
processors or the like, or associated modules thereof, such as various
semiconductor memories,
tape drives, disk drives and the like, which may provide non-transitory
storage at any time for
the software programming. All or portions of the software may at times be
communicated
through the Internet or various other telecommunication networks. Such
communications, for
example, may enable loading of the software from one computer or processor
into another, for
example, from a management server or host computer into the computer platform
of an
application server. Thus, another type of media that may bear the software
elements includes
optical, electrical and electromagnetic waves, such as used across physical
interfaces between
local devices, through wired and optical landline networks and over various
air-links. The
physical elements that carry such waves, such as wired or wireless links,
optical links or the
like, also may be considered as media bearing the software. As used herein,
unless restricted
to non-transitory, tangible "storage" media, terms such as computer or machine
"readable
medium" refer to any medium that participates in providing instructions to a
processor for
execution.
[0204] Hence, a machine readable medium, such as computer-executable code, may
take many
forms, including but not limited to, a tangible storage medium, a carrier wave
medium or
physical transmission medium. Non-volatile storage media include, for example,
optical or
magnetic disks, such as any of the storage devices in any computer(s) or the
like, such as may
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be used to implement the databases, etc. shown in the drawings. Volatile
storage media include
dynamic memory, such as main memory of such a computer platform. Tangible
transmission
media include coaxial cables; copper wire and fiber optics, including the
wires that comprise a
bus within a computer system. Carrier-wave transmission media may take the
form of electric
or electromagnetic signals, or acoustic or light waves such as those generated
during radio
frequency (RF) and infrared (IR) data communications. Common forms of computer-
readable
media therefore include for example: a floppy disk, a flexible disk, hard
disk, magnetic tape,
any other magnetic medium, a CD-ROM, DVD or DVD-ROM, any other optical medium,
punch cards paper tape, any other physical storage medium with patterns of
holes, a RAM, a
ROM, a PROM and EPROM, a FLASH-EPROM, any other memory chip or cartridge, a
carrier
wave transporting data or instructions, cables or links transporting such a
carrier wave, or any
other medium from which a computer may read programming code and/or data. Many
of these
forms of computer readable media may be involved in carrying one or more
sequences of one
or more instructions to a processor for execution.
III. Illustrative Reversible Turbomachinery in a PHES System
[0205] In some embodiments of a pumped heat energy storage system, reversible
turbines/compressors (turbomachines) can be used in place of one or more
conventional
compressors 1 and turbines 3. In a PHES system with reversible turbomachines,
the flow of
working fluid 20 may be reversed between charge and discharge cycles, such
that the working
fluid may flow in one direction through a reversible turbomachine during a
charge cycle and
flow in the opposite direction through the reversible turbomachine during a
discharge cycle,
such that the reversible turbomachine is acting as a compressor during one
cycle and acting as
a turbine during another cycle. Additionally, the turbomachine may spin in one
direction
during the charge cycle and spin in the opposite direction during the
discharge cycle.
Reversible turbomachines can be implemented in the arrangements previously
described
herein, provided that the working fluid flow direction is reversed between
charge and
discharge_ For example, reversible turbomachinery can be used in place of some
or all of the
compressors 1 and/or turbines 3 in the PHES systems shown in FIGS. 2-3, 6-7,
13-16, 20-21,
22, 23. 24-25, and 27-28. As an example, reversible turbomachinery can be used
to replace
both compressor 1 and turbine 3 in FIG. 2. As another example, reversible
turbomachinery
can be used to replace one compressor 1 and/or one turbine 3 in FIG. 23.
[0206] As non-limiting illustrative examples, FlGs. 31A and 31B illustrate a
non-recuperated
PHES system, such as in FIGs. 2-3, with a pair of reversible turbomachines 101
and 103,
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operating in charge mode and discharge mode respectively. Similarly, FIGs. 32A
and 32B
illustrate a recuperated PHES system, such as in FIGs. 13-16, with a pair of
reversible
turbomachines 101 and 103, operating in charge mode and discharge mode
respectively. Fluid
flow paths/directions for the working fluid 20 are indicated by arrows.
[0207] FIG. 31A illustrates a charge-mode fluid flow path of working fluid 20
from reversible
turbomachine 101 (acting as a compressor), through hot side heat exchange 2,
through
reversible turbomachine 103 (acting as a turbine), through cold side heat
exchanger 4, and back
to reversible turbomachine 101.
[0208] FIG. 31B illustrates a discharge-mode fluid flow path of working fluid
20 from
reversible turbomachine 103 (acting as a compressor), through hot side heat
exchange 2,
through reversible turbomachine 101 (acting as a turbine), through cold side
heat exchanger 4,
and back to reversible turbomachine 103.
[0209] FIG. 32A illustrates a charge-mode fluid flow path of working fluid 20
from reversible
turbomachine 101 (acting as a compressor), through hot side heat exchange 2,
through
recuperator 5, through reversible turbomachine 103 (acting as a turbine),
through cold side heat
exchanger 4, through recuperator 5, and back to reversible turbomachine 101.
[0210] FIG. 32B illustrates a discharge-mode fluid flow path of working fluid
20 from
reversible turbomachine 103 (acting as a compressor), through recuperator 5,
through hot side
heat exchange 2, through reversible turbomachine 101 (acting as a turbine),
through
recuperator 5, through cold side heat exchanger 4, and back to reversible
turbomachine 103.
[0211] Excess heat in the PHES systems of FIGs 31A, 31B, 32A, and 32B may be
removed
from the systems through heat rejection devices, such as heat rejection device
56 acting on, and
in thermal contact with, the HTS medium 21, heat rejection device 55 acting
on, and in thermal
contact with, the CTS medium 22, and/or heat rejection device 57 acting on,
and in therinal
contact with, the working fluid 20. The heat rejection devices 55, 56, 57 may
be operated by
opening or closing valves that allow the thermal storage mediums 21, 22 or the
working fluid
20, as appropriate for the location, to flow through them with thermal contact
or bypass them
as needed for heat removal. The heat rejection devices 55, 56, 57 may be, for
example and not
limited to, passive or active ambient air coolers, radiators, or heat
exchangers in thermal contact
with heat sinks. The heat rejection devices 55, 56, 57 are illustrated in
specific locations for
example purposes only and may be in different locations or not present at all.
For example,
heat rejection devices may additionally or alternatively be located on cold-
side tank(s) 6, hot
side tank(s) 8, or at another location or locations along the flow path of
working fluid 20.
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[0212] In each of the described examples, additional components may be
included in or along
the fluid paths, and/or illustrated components may be sets of components. For
example, one
or more valved bypass or recirculation paths may be included in or along the
fluid paths. As
another example, relief, drain, or fluid injection valves may be included in
or along the fluid
paths. As another example, an illustrated heat exchanger, heat rejection
device, tank, or
reversible turbomachinc may include a set of heat exchangers, heat rejection
devices, tanks, or
turbomachines, respectively, operating or capable of operating in series, in
parallel, or in some
other arrangement.
[0213] The PHES systems of FIGs 31A, 31B, 32A, and 32B may include an
inventory control
system in which one or more tanks 64 are connected to, or isolated from, the
working fluid 20
flow.
[0214] Under some operating conditions, the one or more tanks 64 may be
connected by one
or more valves (e.g., a three-way valve) to working fluid 20 flow at a high-
pressure side of the
PHES system (e.g., between the outlet of a reversible turbomachine acting as a
compressor and
the inlet of a reversible turbomachine acting as a turbine). At the same time
the one or more
tanks 64 may be isolated by one or more valves (e.g., a three-way valve) from
working fluid
20 flow at a low-pressure side of the PHES system (e.g., between the outlet of
a reversible
turbomachine acting as a turbine and the inlet of a reversible turbomachine
acting as a
compressor), thereby accepting and storing high-pressure working fluid 20 into
the one or more
tanks 64.
[0215] Under other operating conditions, the one or more tanks 64 may be
isolated from the
high-pressure side of the PHES system by the one or more valves (e.g., the
three-way valve),
while at the same time being connected to the low-pressure side of the PHES
system by one or
more valves (e.g., the three-way valve), thereby releasing high-pressure
working fluid 20 from
the one or more tanks 64 into the low-pressure side of the PHES system.
[0216] Under other operating conditions, the one or more tanks 64 may be
isolated from both
the high-pressure side of the PHES system and the low-pressure side of the
PHES system by
one or more valves (e.g., the three-way valve).
[0217] Fluid connections may be arranged differently in a PHES system with
reversible
turbomachines than in a PHES system with traditional compressor(s) and
turbine(s). For
example, in both systems it may be desirable to have heat exchangers 2, 4
configured as
counterflow heat exchangers. In PHES systems with non-reversible
compressors/turbines, if
the working fluid 20 path always enters the same side of the heat exchangers
2, 4, then the hot
side tanks 6, 7 will swap their thermal fluid 21 path connections to the hot
side heat exchanger
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2 between charge and discharge cycles, as illustrated by positions 32, 33
swapping between
FIGs. 2 and 3. This also occurs with cold side tanks 8, 9 and cold side heat
exchanger 4.
[0218] Conversely, in PHES systems with reversible turbomachines, such as
turbomachines
101, 103, the thermal fluid path connections between the tanks 6, 7, 8, 9 do
not need to swap
with their respective heat exchangers 2, 4 during changeover from charge mode
to discharge
mode in order to maintain counterflow because the working fluid is changing
flow direction,
as illustrated between FIGs. 31A and 31B, and also FIGs. 32A and 32B.
Therefore,
beneficially, a PHES system with reversible turbomachines can have simplified
fluid paths for
thermal fluids 21, 22 and potentially avoid the use of valve arrangements for
changing the
thermal fluid path connections between the tanks 6, 7, 8, 9 and their
respective heat exchangers
2, 4 for counterflow heat exchange.
[0219] Drivetrains for reversible turbomachines, such as turbomachines 101,
103, may be
arranged in various manners. For illustration purposes only, the drivetrains
of reversible
turbomachines 101 and 103 are shown in FIGs. 31A, 31B, 32A, 32B, and 33
connected by a
common mechanical shaft 10 and connected to a motor/generator 11. However,
other
drivetrain arrangements are possible. FIG. 34 shows another example drivetrain
arrangement
in which reversible turbomachines 101 and 103 are connected to separate
mechanical shafts
10A and 10B, and connected to separate motor/generators 11A and 11B.
Additionally,
gearboxes, clutches, and other drivetrain components may be included as part
of, or in between
any of, reversible turbomachines 101, 103, their respective shafts 10, 10A,
10B, and/or their
respective motor/generators 11, 11A, 11B. Also, motor/generators 11, 11A, 11B
are each
illustrated as a single component that can handle both driving the
turbomachine(s) as a motor
using external power (e.g., electricity) and being driven by the
turbomachine(s) as a generator
for generating electrical power. However, the motor/generators 11, 11A, 11B
may each include
separate motor and generation components and may further include gearboxes,
clutches, and
other drivetrain components that allow the motor/generators 11, 11A, 11B to
act as motors or
generators depending on the operating state of the PHES system, such as charge
or discharge
mode. The aforementioned drivetrain and motor/generator arrangements may he
implemented
along with the reversible turbomachines in any of the previously described
PHES systems.
IV. Illustrative Reversible Turbomachinery
[0220] Reversible turbomachines (e.g., turbomachines 101, 103) in a PHES
system require
different designs than a dedicated conventional compressor (e.g., compressor
1) or turbine
(e.g., turbine 3). In PHES systems as described herein, reversible
turbomachines may be
symmetrically identical, or symmetrically similar, in operation and design. In
some
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embodiments, the symmetry may not be exact because of, for example, the
different
temperature conditions on the two sides of the system. The reversible
turbomachines herein
differ from conventional turbine/compressor arrangements in at least two
aspects: (i) the
number of turbine stages is large, so that the pressure loading on each is
small and (ii) the
blades are compromise airfoils that are aerodynamically symmetrical. As to the
second aspect,
small loading on the turbine blades allows them to be thin and to have leading
and trailing
edges that are simultaneously tapered and blunt.
[0221] FIG. 35 illustrates simplified representations of a pair of reversible
turbomachines 101,
103 that can be implemented in PHES systems described herein. Turbomachine 101
is
illustrated acting as a turbine, with working fluid entering on the right
through inlet diffuser
101C (shown as dashed lines) and exiting on the left through outlet diffuser
101D (shown as
dashed lines). Turbomachine 103 is illustrated acting as a compressor, with
working fluid
entering on the right through an inlet diffuser (shown as dashed lines) and
exiting on the left
through an outlet diffuser (shown as dashed lines). For turbomachines 101, an
illustrative
simplified cross-section of the rotor 101A and stator 101B is also depicted.
Rotor blades 74A
and stator blades 74B are shown as simplified representations only and may
have different
shapes and quantities in embodiments. The representations of turbomachines
101, 103 are
illustrated in FIG. 35 as mirror images of each other, but they be different
than each other in
some embodiments and/or may have differences not shown in the representations.
Stage Loading
[0222] Conventional turbine/compressor blades are airfoils designed to work in
one way only.
Modern conventional compressor blades are quite thin and so nearly symmetrical
that it is
difficult to distinguish by eye the leading edge from the trailing edge. This
near-symmetry
results from light loading of the compressor stages, a feature required to
suppress surge. A
lightly loaded stage is optimized with the blade that is thin, and a thin
blade must be relatively
symmetric. A relatively sharp leading edge on conventional compressor blades
is also desirable
for operation in the transonic range for the purpose of controlling bow
shocks.
[0223] In contrast to conventional compressor blades, modern conventional
turbine blades are
typically heavily loaded, thick, and asymmetric. This design is favored
because it reduces size
and weight (critically important in an aircraft engine), and also provides
room for internal blade
cooling. However, in a stationary PHES system, there is no need to minimize
weight, and the
operating temperatures can be sufficiently low such that blade cooling is not
required.
Additionally, it can be preferable to operate PHES system turbomachinery
without blade
cooling because the cooling can generate undesirable entropy in the system.
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Nomenclature
[0224] The following nomenclature is used herein with respect to turbomachine
blades.
TABLE I
a blade inlet angle (metal)
13 blade outlet angle (metal)
setting angle
c cord length
spacing (or pitch)
tinaz maximum blade thickness (width)
p pressure
Ap pressure change across a stage
T temperature
AT temperature change across a stage
vs speed of sound
y cr / c,
annular radius of turbomachine
(mean line analysis)
vz axial fluid flow velocity through
turbomachine
w angular rotation frequency
wr azimuthal blade velocity
cp working fluid constant-pressure
heat capacity per mole
cv working fluid constant-volume
heat capacity per mole
mõ working fluid mass per mole
R ideal gas constant
Considerations for Velocity Triangles
[0225] For a rotor with thin blades operating at fixed angular frequency w,
and no significant
external load (i.e., a PHES system with large heat exchangers), both the axial
velocity of the
working fluid vz and compression ratio 1 + Ap/p are fixed by the annular
radius r and the edge
blade angles a and D. A reversible turbomachine will have blade angles ar and
13r for the rotor
and blade angles as and lis for the stator. Blade entrance angles ur and as
may be the same or
different, and blade exit angles lir and s may be the same or different. For
purposes of
simplified illustration, the rotor and stator blades will be shown and
discussed as mirror images
of each other (i.e., a = ar = as and 13 = =
[0226] FIGs. 36-39 illustrate reversible turbomachine blades compared to
conventional turbine
and compressor blades. In each of FIGs. 36-39, representative direction of
fluid flow (at
velocity vz) through the turbomachine is illustrated by arrows. Representative
rotational
direction of rotor blades (at azimuthal blade velocity wr) is shown by an
arrow.
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[0227] As illustrated in FIG. 36, conventional turbine blades 72A (rotor) and
72B (stator) may
be asymmetric and thick, with high camber, blunt tips, and tapered tails. As
illustrated in FIG.
37, conventional compressor blades 73A (rotor) and 73B (stator) may be
asymmetric and thin,
with low camber, blunt tips, and tapered tails. In contrast, reversible
turbomachine blades 74A
(rotor) and 74B (stator) may be symmetric and thin, with low camber, and tips
that range from
tapered to moderately blunt.
[0228] FIG. 40 is a simplified representation of a compressor or turbine
stage, viewed axially.
For example purposes only, rotor 101A is illustrated with a representation of
turbomachine
blades 74A. It is characterized by an angular rotation frequency co and a
radius r, which may
be considered the center in "mean line" analysis. Azimuthal blade velocity or
may be
calculated as the product of those values. In practice, the blade angles
actually twist slightly
between the inner radius (hub) and outer radius (tip) to accommodate slightly
different values
of wr.
[0229] FIG. 41 is a simplified representational diagram illustrating
nomenclature for reversible
turbomachine blades in a blade row. Reversible turbomachine blades 74A are
illustrated as an
example of a symmetrical blade, such as a double-circular arc ("DCA") blade,
with a camber
line 80, a chord line 81, a blade angle (metal) a at a first blade tip, a
blade angle (metal) it at a
second blade tip, a chord length c, a blade spacing (pitch) s, a maximum blade
thickness -max,
and a setting angle ;. In one embodiment, a reversible turbomachine blade
(rotor or stator),
such as reversible rotor blade 74A, may take the form of a double-circular arc
with blade angles
ci = 0 and it = -0.1657t, with a chord length c dependent on overall desired
performance
characteristics, a blade spacing (pitch) s dependent on overall desired
performance
characteristics, a maximum blade thickness t. = 0.05*c, and a solidity c/s =
2. DCA and
DCA-like blades with thin tm and slightly blunted tips are particularly good
shapes for
reversible turbomachine blades, given their symmetry.
Minimal Loss Condition
[0230] At the minimal-loss point, fluid exiting one blade row must match the
entrance angle
of the next blade row downstream in its rest frame. This is achieved when
r = [tan(a) + tan(/3)] (1)
Euler Turbine Equation
[0231] Conservation of energy causes the temperature change AT across a stage
to obey:
AT ¨ +6o2r2 (2)
mv
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where CV and my are the working fluid constant-pressure heat capacity per mole
and mass per
mole, respectively. The plus and minus signs refer to compressor and turbine
mode. From the
thermodynamic relations of the gas working fluid:
y vAT jyRT
c = ________________________________________________ 19 R ________ , =
(3)
p \y¨l/ T \y¨l/ ntv
where vs denotes the speed of sound, giving:
Lp cor )2
¨ = (4)
vs
Half-Reaction Condition
[0232] The exact half-reaction condition, in which the fluid enthalpy
loss/gain occurs half in
the rotor and half in the stator, is:
to2r2 vz 2 Ran2 (ff) s _
tan2 (a)] (5)
Dividing Eq. (5) by Eq. (1) gives:
cor = +12,[tan(a) ¨ tan(f3)] (6)
Combining with Eq. (1) then gives:
a = 0.0 /3=-- (7)
12z
Example Embodiments
[0233] A working rule of thumb in gas turbine design is that compressor
loading should be no
greater than Apip=0.2. Substituting this value into Eq. (4) gives lwri/vs =
0.346 for argon (y =
5/3) and lowl/vs = 0.378 for air (y = 1.4). Setting the total mach number to
just slightly subsonic,
(vz2 2r2)1120),
= 0.7vs, gives vz/vs = 0.608 for argon and vz/vs = 0.563 for air, resulting in
Cl = 0
and J = -0.1657r for argon, and a = 0 and 13 = -0.1887r for air.
Desgin of Reversible Turbomachine Blades as Compared to Conventional Blades
[0234] An axial compressor must have a large number of stages in order to have
high adiabatic
efficiency and to have a safe margin against surging. A conventional axial
turbine does not
have this requirement. Conventional turbine stages are typically loaded more
heavily to save
money and weight, but high loading is not essential for obtaining high
efficiency. Extremely
high efficiencies (0.91 polytropic) are achieved by present-day conventional
multi-stage axial
compressors. As previously indicated, reversible turbine blades are compromise
airfoils that
are aerodynamically symmetrical. However, for a PHES system, in which energy
is both being
generated and stored with reversible turbomachines, high loading is not
essential for obtaining
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high efficiency. Reversible turbomachines (acting as turbines) with
appropriately designed
blades, as disclosed herein, can achieve high efficiencies.
[0235] An adiabatic process is one in which no entropy is created. Entropy is
a measurable
quantity for systems in thermal equilibrium. In the case of a fluid, it is
synonymous with the
equations of state. Thus one measures the entropy of the working fluid before
the
turbomachinery has acted upon it, allows the turbomachinery to act, and then
measures it again,
and the fluid's entropy will have increased. A good design minimizes this
increase. At the limit
where the increase is zero, the motion is perfectly adiabatic. When the
withdrawal or addition
of energy to a system is sufficiently slow that it becomes adiabatic, the
action becomes
reversible.
Blade Shape
[0236] The chief control over entropy creation in reversible turbomachinery is
blade shape.
The base entropy production, due to viscous drag of laminar flow through rotor
and stator
channels, cannot be controlled, but may be too small to be significant. The
more important
contribution comes from boundary layer separation and turbulence in the
negative-pressure
(lift) portions of the blade. This problem, called stall, is potentially
severe.
[0237] A solution is to make the blade a good airfoil. This is illustrated
with respect to FIGs.
42-45. Rather than being a simple plane, a good blade has an upper surface
slightly bowed
upward vis-a-vis the lower surface so as to maximize the Venturi effect and
discourage
boundary layer separation (stall). It is important that this bowing occur
forward in the blade
and that the tailing end of the blade be relatively sharp. The latter
minimizes wake turbulence
by reconnecting the upper and lower air flows smoothly as they leave the
blade.
[0238] FIGs. 43-45 illustrates three blade shapes built on the same blade arc
70 (e.g., camber
line) of FIG. 42. The three shapes are built on the same central arc 70, but
have different airfoil
modifications as described further below. Each blade in FIGs. 43-45 illustrate
turbomachinery
blades shapes with the blade angles a and 13 appearing in Eqns. (8) and (9).
The blade angles a
and
are the same for all 3 cases. As illustrative examples only, the
particular values illustrated
in each of FIGs. 43-45 are a = 0.3n and f = ¨0.1n. The sign convention for
angles is
counterclockwise positive. The turbine and compressor blades in FIGs. 43 and
44, respectively,
have a sharp tailing edge, as required to minimize drag and a blunt leading
edge to allow for
attack angle flexibility. For a PHES system, the bluntness of the front end of
the blade is not
actually crucial. Its chief function is to facilitate a wide range of attack
angles. In an airplane
wing, this is important for maintaining control when the plane's attitude
changes, for example,
at the moment of upward rotation at takeoff. However, inside of a
turbomachine, the blunt front
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end has a less critical function of allowing accommodation to changing loads.
The reversible
blade in FIG. 45 is a compromise that can be made if the attack angle is known
relatively well
in advance and the slight performance degradation of the extra drag due to the
moderately blunt
tail is not prohibitive.
Angles: Axial Velocity and Pressure Ratio
[0239] In axial turbomachincry rotating at fixed frequency co, the axial
velocity of fluid vz and
compression ratio 1 Ap/p are fixed by the annular radius r and the two blade
angles a and II
A general design actually has four angles¨two for the rotor and another two
for the stator¨
but we shall consider only the "half reaction" case in which the rotor and
stator blades are
mirror images of each other, as illustrated in FIGs. 46-49.
[0240] In the half reaction case, FIGs. 46-49 show that the compressor stage
is simply the
turbine stage with time reversed (i.e. with motions of rotor and fluid both
reversed) and with
the airfoil modifications of the blade also reversed. The blade arc and
entrance and exit angles
a and J are exactly the same in both cases. The axial flow velocity vz
direction is represented
by arrows. The azimuthal blade velocity wr direction is represented by an
arrow. In the
example shown, the pitch-to-chord ratio is 0.81. FIGs. 48 and 49 are the same
as FIGs. 46 and
47, respectively, except with the reversible blade of FIG. 45 substituted for
conventional
asymmetric airfoils from FIGs. 43 and 44.
[0241] For all practical purposes, the flow direction of fluid leaving a row
of blades is fixed by
the blade exit angle, but in the rest frame of the blade. The latter requires
transforming in and
out of the rotor' s moving frame, the mathematics of which can be accomplished
and understood
with respect to velocity triangles. It is important that the exit flow
direction from a blade row
is not affected significantly by the entrance angle.
Axial Velocity
[0242] In the special case of no significant fluid resistance outside the
turbomachinery the axial
velocity adjusts itself to satisfy:
wr = 19, Ran(a) + tanC6)] (8)
with sign conventions as discussed herein, for any value of or. The reason for
this may be seen
from FIGs. 46-49. In turbine mode, fluid exiting a stator with azimuthal
velocity vLtan(a) will
appear to the moving rotor to have azimuthal velocity vLtan(a) - wr.
Minimizing turbulent
losses at the interface requires that this equal ¨vztan(p) (with sign
convention that IS < 0).
Similarly, fluid exiting the rotor at azimuthal velocity (in its own frame)
¨vztan(a) appears to
the stator to have azimuthal velocity wr - vz tan(a). Minimizing turbulent
losses at this interface
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requires this to equal vztan(13), which gives the same result. The same
reasoning works in
compressor mode as well and the result is Eqn. (8), with the signs of co and
vz both reversed.
Compression Ratio
[0243] The pressure change across a stage is determined by conservation of
momentum,
conservation of energy, and the working fluid equation of state. This is
succinctly summarized
in the Euler turbine equation, a simple form of which is:
cp
¨ AT = -Ecortz, [tan (a) + tan (f3)] (9)
mv
where cp and my are the gas constant-pressure heat capacity per mole and mass
per mole,
respectively, and AT is the temperature change across the stage. The plus and
minus signs refer
to compressor and turbine mode, respectively. Thus, in turbine mode, the
temperature drops
across the stage, whereas in compressor mode, it increases. The reasoning
leading to the Euler
turbine equation is similar to that of velocity matching. In turbine mode, iz
moles per second
passing through the rotor result in a rate of angular momentum increase in the
rotor of
L = mv-i7rvz[tan(a) + tan (a)]. Since this must equal the torque exerted by
the turbine on the
crankshaft, the power delivered is wt. This must equal the power extracted
from the working
fluid --bcpAT.
[0244] From the thermodynamic relations of a gas working fluid:
Ap _ ( y )61 ( y
cP ¨ _______________________________________________ 1)R (10)
¨
Where y = epic, (c, is the constant-volume heat capacity) and cp ¨ cv = R is
the ideal gas
constant, we then obtain:
,Ap 177. W7-1,1
v z
¨p= + RT [tan(a) + tan(/3)] (11)
Stage Loading (Revisited)
[0245] Because Eqns. (8) and (11) depend on angles solely, and are independent
of blade
shaping, the chief (but not sole) requirement for making a turbine-compressor
pair reversible
is to reduce the loading of the turbine stages to match that of the compressor
stages. A working
rule of thumb in gas turbine design is that compressor loading should be no
greater than Ap/p
= 0.2. As previously noted, the speed of sound is:
1yRT
vs = ¨ ,
mv (12)
Combining Eqns. (8) and (11) provides:
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Lp ciry
(13)
vs )
or korl/vs = 0.38 for room-temperature air (7 = 1.4). For the specific angles
ci = 0.37c and D =
¨0.17E shown FIGs. 42-47, this then gives:
tor
Ivzi = 1 ________ (14)
tanot) + tan(8)
or Ivzi/vs = 0.36. This is about half the design rule of thumb Ivzi/vs = 0.7,
so the magnitudes of
u and J in Figs. 42 - 47 are about twice that.
Symmetrical Blade Shapes
[0246] In addition to decreasing the load on the turbine stages and increasing
their number, a
reversible turbine/compressor also requires a modified blade shape that is
more symmetrical.
This is shown, e.g., in FIGs. 38-39, 41, 45, and 48-49. With symmetrical blade
shapes for the
rotor and stator blades, the turbine and compressor become time-reverses of
each other, as
shown in FIGS. 38-39 and 48-49. Neither of the two key compromises required to
make the
blade symmetrical is extraordinary in compressor design. The first, a slightly
blunted trailing
edge, is actually implemented to some extent in all compressor blades because
an infinitely
thin blade is structurally impossible. The second, a more tapered leading
edge, actually
provides efficiency advantages in thin blades.
V. Conclusion
[0247] While various aspects and embodiments have been disclosed herein, other
aspects and
embodiments will be apparent to those skilled in the art. The various aspects
and embodiments
disclosed herein are for purposes of illustration and are not intended to be
limiting, with the
true scope and spirit being indicated by the following claims.
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