U.S. patent application number 14/748130 was filed with the patent office on 2015-12-03 for compressed gas energy storage system.
The applicant listed for this patent is LightSail Energy, Inc.. Invention is credited to Edwin P. BERLIN, JR., Stephen E. CRANE, Danielle A. FONG, AmirHossein POURMOUSA ABKENAR.
Application Number | 20150345522 14/748130 |
Document ID | / |
Family ID | 53785936 |
Filed Date | 2015-12-03 |
United States Patent
Application |
20150345522 |
Kind Code |
A1 |
FONG; Danielle A. ; et
al. |
December 3, 2015 |
COMPRESSED GAS ENERGY STORAGE SYSTEM
Abstract
Embodiments of the present invention use compressed air to store
and deliver electrical, mechanical, and/or thermal power with high
round-trip efficiency. Various embodiments may be scalable for use
in a variety of environments--from wind farms to power plants to
motor vehicles. An energy storage system according to the present
invention can operate as a stand-alone storage system that connects
electrically to the grid, it can be tightly integrated with a wind
turbine, and/or it can be co-located with a thermal power
generation facility and operate with even higher efficiency by
scavenging low-grade waste heat.
Inventors: |
FONG; Danielle A.; (Oakland,
CA) ; CRANE; Stephen E.; (Santa Rosa, CA) ;
BERLIN, JR.; Edwin P.; (Oakland, CA) ; POURMOUSA
ABKENAR; AmirHossein; (Walnut Creek, CA) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
LightSail Energy, Inc. |
Berkeley |
CA |
US |
|
|
Family ID: |
53785936 |
Appl. No.: |
14/748130 |
Filed: |
June 23, 2015 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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13411484 |
Mar 2, 2012 |
9109614 |
|
|
14748130 |
|
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|
61449403 |
Mar 4, 2011 |
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Current U.S.
Class: |
60/659 |
Current CPC
Class: |
F15B 2211/527 20130101;
F04B 41/02 20130101; F15B 2211/625 20130101; F15B 21/042 20130101;
Y02E 60/16 20130101; F28D 2020/0082 20130101; Y02E 60/14 20130101;
F15B 21/048 20130101; F28D 20/0034 20130101; F15B 21/14 20130101;
F15B 1/265 20130101; F15B 1/024 20130101; F15B 2201/42 20130101;
H02J 15/006 20130101 |
International
Class: |
F15B 21/04 20060101
F15B021/04; F15B 1/26 20060101 F15B001/26 |
Claims
1. An apparatus comprising: an expander comprising an inlet valve
configured to receive a pressurized gas, the expander further
comprising an element configured to introduce liquid to effect
gas-liquid heat exchange with gas being expanded; a
proportional-integral-derivative (PID) controller configured to
adjust a timing of the inlet valve to close when a percentage of
the expander filled is a quotient of downstream and upstream
pressures; a gas-liquid separator configured to separate the liquid
from a gas-liquid mixture received from the expander; and an
insulated thermal storage unit in thermal communication with the
separated liquid.
2. An apparatus as in claim 1 wherein the insulated thermal storage
unit is configured to store the separated liquid.
3. An apparatus as in claim 1 wherein the insulated thermal storage
unit is configured to store a second liquid in thermal
communication with the separated liquid through a heat
exchanger.
4. An apparatus as in claim 3 wherein the second liquid is
maintained at a lower pressure than the gas-liquid mixture.
5. An apparatus as in claim 3 wherein the separated liquid and the
second liquid comprise water.
6. An apparatus as in claim 3 wherein the separated liquid
comprises water and the second liquid comprises an oil.
7. An apparatus as in claim 1 wherein the inlet valve is configured
to receive the pressurized gas from a compressed gas storage
unit.
8. An apparatus as in claim 7 wherein inlet valve comprises an
active valve having a valve timing changed as the compressed gas
storage unit depletes.
9. An apparatus as in claim 1 wherein the inlet valve is configured
to receive the pressurized gas from a higher pressure expansion
stage.
10. An apparatus as in claim 1 wherein the expander comprises a
reversible compressor/expander.
11. An apparatus as in claim 10 wherein the insulated thermal
storage unit is in thermal communication with a heat source when
the reversible compressor/expander is operating as an expander.
12. An apparatus as in claim 11 wherein the heat source is
internal.
13. An apparatus as in claim 11 wherein the heat source is
external.
14. An apparatus as in claim 10 further comprising a second
insulated thermal storage unit configured to be in thermal
communication with the separated liquid when the reversible
compressor/expander is operating as a compressor.
15. An apparatus as in claim 14 wherein the second insulated
thermal storage unit is in thermal communication with a heat
source.
16. An apparatus as in claim 10 further comprising a second
insulated thermal storage unit configured to be in thermal
communication with liquid separated by a second gas-liquid
separator when the reversible compressor/expander is operating as a
compressor.
17. An apparatus as in claim 10 wherein the insulated thermal
storage unit comprises a moveable partition defining, a first
chamber receiving the separated liquid when the reversible
compressor/expander is operating as a compressor, and a second
chamber receiving the separated liquid when the reversible
compressor/expander is operating as an expander.
18. An apparatus as in claim 10 wherein the reversible
compressor/expander is in selective communication with a Variable
Frequency Drive (VFD).
19. An apparatus as in claim 1 wherein a majority of the gas-liquid
heat exchange is sensible heat.
20. An apparatus as in claim 1 wherein the element comprises a
pre-mixing chamber.
Description
CROSS-REFERENCE TO RELATED APPLICATION
[0001] The instant nonprovisional is a continuation of U.S. patent
application Ser. No. 13/411,484, filed Mar. 2, 2012, which claims
priority to U.S. Provisional Patent Application No. 61/449,403
filed Mar. 4, 2011, which is incorporated by reference in its
entirety herein for all purposes.
BACKGROUND
[0002] The availability of low-cost energy storage at utility
scales would address a number of issues relating to the energy
grid. In particular, the deployment of renewable power generators,
such as wind and solar, may be limited by a difficulty in these
technologies providing a reliable supply of power at predictable
times or during periods of high demand.
SUMMARY
[0003] Embodiments of the present invention use compressed air to
store and deliver electrical, mechanical, and/or thermal power with
high round-trip efficiency. Various embodiments may be scalable for
use in a variety of environments--from wind farms to power plants
to motor vehicles. An energy storage system according to the
present invention can operate as a stand-alone storage system that
connects electrically to the grid, it can be tightly integrated
with a wind turbine, and/or it can be co-located with a thermal
power generation facility and operate with even higher efficiency
by scavenging low-grade waste heat.
BRIEF DESCRIPTION OF THE DRAWINGS
[0004] The patent or application file contains at least one drawing
executed in color. Copies of this patent or patent application
publication with color drawing(s) will be provided by the Office
upon request and payment of the necessary fee.
[0005] FIG. 1A shows droplet distribution (colored by particle
velocity) modeled during compression for 200 micron droplets in a
cylinder. FIG. 1B shows the modeled droplet distribution for 50
micron droplets.
[0006] FIG. 2A shows the temperature distribution within the
cylinder for 200 micron droplets. FIG. 2B shows the temperature
distribution within the cylinder for 50 micron droplets.
[0007] FIG. 3 provides the P-V curve at various particle sizes.
[0008] FIG. 4 shows the spatially average temperature inside the
cylinder.
[0009] FIG. 5 shows the particle distribution when the nozzles are
present in the center of the valve pockets.
[0010] FIG. 6 shows the temperature distribution inside the
cylinder.
[0011] FIG. 7 shows P-V curve at various particle diameter when the
nozzles are present in valve pocket.
[0012] FIG. 8 shows a simplified view of an embodiment of an energy
storage system.
[0013] FIG. 9 shows a simplified view of an alternative embodiment
of an energy storage system.
[0014] FIG. 9A shows various basic operational modes of the system
of FIG. 9.
[0015] FIGS. 9BA-BF show simplified views of the gas flow paths in
various operational modes of the system of FIG. 9.
[0016] FIG. 10 plots cost of various energy storage approaches.
[0017] FIG. 11 shows one embodiment of a single-stage configuration
of an energy storage system.
[0018] FIG. 12 plots droplet velocity versus time.
[0019] FIG. 13 plots heat transfer coefficient versus droplet
diameter.
[0020] FIG. 14 plots temperature versus initial volume proportion
of water in air.
[0021] FIG. 15 plots efficiency versus water volume fraction.
[0022] FIG. 16A shows the internal geometry of a nozzle
embodiment.
[0023] FIG. 16B shows a simulation of the velocity droplets
emerging from the nozzle of FIG. 16A
[0024] FIG. 17 shows a swirl nozzle.
[0025] FIGS. 18A-F shows liquid injection into a piston.
[0026] FIG. 19 shows schematically where the energy goes during the
complete compression/expansion cycle.
[0027] FIG. 20 shows a photograph of a compression cylinder.
[0028] FIG. 21 shows a cross-section.
[0029] FIG. 22 shows the position of various sensors in the valve
covers and in the intake and exhaust manifolds.
[0030] FIG. 23 shows the cylinder with the cover removed,
indicating how the water is distributed to the nozzle
manifolds.
[0031] FIG. 24 shows one configuration of the outboard cylinder
head.
[0032] FIG. 25 shows the above head configuration running at about
0.3 liters/sec total flow rate and 100 psi pump pressure.
[0033] FIG. 26 shows display of data from about 50
sensors--pressure, temperature, air and water flow rates,
electrical current, crank angle, and shaft torque.
[0034] FIG. 27 shows a PV diagram of a typical run.
[0035] FIGS. 28-30 show views of an embodiment of an active
valve.
[0036] FIGS. 31-31D are schematics for one embodiment of a
system.
[0037] FIG. 32 plots the efficiency as a function of the polytropic
index.
[0038] FIG. 33 shows an embodiment of an energy storage system
utilizing separate insulated thermal storage tanks for hot and cold
liquid.
[0039] FIG. 34 shows an embodiment of an energy storage system
utilizing a single insulated thermal storage tank having a moveable
partition.
[0040] FIG. 35 shows an embodiment of an energy storage system
utilizing a liquid stored at a lower pressure.
[0041] FIG. 36A shows an embodiment of an energy storage system
utilizing heat exchange between a separated liquid and a second
liquid stored at a lower pressure.
[0042] FIG. 36B shows an alternative embodiment of an energy
storage system utilizing heat exchange between a separated liquid
and a second liquid stored at a lower pressure.
[0043] FIG. 37A shows internal energy flows in one embodiment of a
compression cycle.
[0044] FIG. 37B shows internal energy flows in one embodiment of an
expansion cycle.
DESCRIPTION
[0045] Incorporated by reference herein for all purposes is the
U.S. Patent Publication 2011/0115223. Compressed air is capable of
storing energy at densities comparable to lead-acid batteries--but
without a battery's limited lifetime or materials availability or
disposal issues.
TABLE-US-00001 Energy Energy Cycle Lifetime Density Density
Efficiency (80% DoD Storage Medium (Wh/kg) (Wh/liter) (%) cycles)
Lead-acid batteries 30 to 60 80 to 150 75 <1000 Lithium-ion
batteries 160 270 90 ~1600 to 3000 Redox flow batteries 10 to 20 15
to 25 70+ 10000? High-speed flywheel 4.3 9.8 85 >10000 Pumped
hydro 0.55 0.55 75 to 85 >10000 (head = 200 m) Compressed air
142 38.9 70+ (See >10000 (LSE tech at 300 atm) below)
[0046] The energy density of compressed air is a function of its
pressure. Off-the-shelf technology exists to store and handle air
at 200 to 300 atmospheres (i.e. the working pressure of scuba
tanks), which may be an operational pressure range for embodiments
of the systems.
[0047] A useful number to keep in mind is that about 140 kJ of
energy (=38.9 watt-hours) is stored in one liter of air at 300
atmospheres pressure. That is, 25.7 liters of 300 atmosphere air
will store one kilowatt-hour. Typical laboratory compressed gas
cylinders hold 80 liters, or just over 3 kWh.
[0048] To retrieve the entire amount of the energy, the expansion
process must be 100% efficient. Development efforts focus upon
coming as close that that number as possible.
[0049] One advantage of compressed air energy storage is its low
capital cost. The source for FIG. 10 is a Department of Energy
(DOE) Energy Advisory Board report on energy storage technologies.
Though each bar comes with a number of caveats, other reports come
to similar conclusions. Costs are discussed in more detail
below.
[0050] Compressed air energy storage (CAES) can be a low-cost
solution for bulk energy storage. However, known CAES technology
losses are supplemented by burning fossil fuels and, even then,
delivering only half the energy stored. Embodiments of the present
invention seek to improve round-trip efficiency and permit CAES to
be used cost-effectively for a broad variety of energy storage
applications.
[0051] Thermodynamics of Compression and Expansion
[0052] Thermodynamics plays a role in making compressing and
expanding air inefficient in most applications. When air is
compressed, it heats up. The heated air wants to expand, resisting
further compression. As a result, extra work may need to be done to
obtain a high compression ratio. Unless all the heat generated is
kept in perfectly insulated thermal storage, the lost heat is
irrecoverable, leading to a low efficiency.
[0053] The reverse situation applies during expansion.
Specifically, the expanding air cools, preventing it from expanding
as much as it would have had it stayed at its original temperature.
Less power is produced as a result.
[0054] Also, compressing to high pressures can heat the air to very
high temperatures. Adiabatic compression (that is, a compression
process during which no heat leaves the compression chamber) from
one atmosphere pressure to 300 atmospheres will heat the air being
compressed to about 1500.degree. C. That creates quite a design
challenge for the compression apparatus--most common materials
won't handle those temperatures, and the thermal stress on the
system is extreme. Similarly, expanding air that has any moisture
in it will result in condensation and freezing of the condensed
water, likely clogging valves and pipes.
[0055] A solution is to remove the heat of compression from the
compression chamber as it's being created--and/or to add heat to
the chamber during expansion. If all the heat generated during
compression is removed as quickly as it's being created and then
added back during expansion, both processes will occur at a
constant temperature. That is, compression and expansion will
proceed isothermally.
[0056] Accordingly, certain embodiments of the present invention
may perform gas expansion or compression in a manner in which a
temperature change experienced by the gas is controlled to be
within a predetermined range. Where that range is relatively
narrow, the expansion or compression can be said to occur in a
manner that is isothermal or near-isothermal.
[0057] In certain embodiments of the present invention, gas may be
compressed or expanded cumulatively across multiple stages. In some
embodiments a temperature change experienced by each stage may be
substantially equivalent (whether the process comprises gas
compression or gas expansion). As referenced herein, the term
"substantially equivalent" refers to a temperature change that
differs by 500.degree. C. or less, by 300.degree. C. or less, by
100.degree. C. or less, by 75.degree. C. or less, by 50.degree. C.
or less, by 25.degree. C. or less by 20.degree. C. or less, by
15.degree. C. or less, by 10.degree. C. or less, or by 5.degree. C.
or less. The temperature change experienced by one or more
particular stages, may be controlled according to embodiments of
the present invention utilizing one or more techniques applied
alone or in combination.
[0058] Isothermal compression and expansion are, in theory,
completely reversible. That is, entropy is not created, and the
processes are 100% efficient--you can get exactly as much energy
back out as you put in.
[0059] In reality however, there are many other processes that
drain energy from the system, including but not limited to:
friction, leakage, pressure drops, etc. But, for high-pressure
systems, thermodynamic inefficiency is by far the biggest loss.
[0060] Note that, if more heat is added during expansion than is
necessary to keep the process isothermal, it is possible to get
more energy out than was originally stored in the form of
compressed air. That is, the cycle can be more than 100%
efficient.
[0061] Of course, the extra energy delivered was obtained from the
extra heat added during expansion. But, if a source of even
low-grade heat is available that would otherwise be wasted--warm
water from a power plant, say, or water heated by the sun--a
compressed air energy storage system can be a net generator of
energy.
[0062] Spray-Mediated Heat Transfer
[0063] In order to maintain near-isothermal conditions, a liquid
(typically water with some additive) may be sprayed into the
compression and/or expansion chambers in order to minimize the
change in temperature. For example, water has a heat capacity of
about 3200 times that of air by mass. So, if there is an equal mass
of air and water, about 99.97% of the heat ends up in the water.
Or, considered another way, instead of rising hundreds of degrees,
the temperature of the compressed air/water droplet aerosol may
rise only a few degrees.
[0064] Note that the presence of liquid doesn't change the amount
of heat that has to be transferred from the air. That heat is still
collected (almost entirely by the water) and transmitted out of the
compression chamber, then transferred to a thermal store of some
kind. However, the temperature of the process stays nearly
constant, allowing efficient operation and making the heat transfer
problem much easier to deal with.
[0065] A factor in allowing rapid heat exchange between the water
introduced into the compression or expansion chamber and the air
being compressed or expanded, is to have a large surface area where
heat exchange can take place. Such a large surface area is
characteristic of uniformly distributed, small liquid droplets.
[0066] There are many considerations in designing a spray system
capable of supporting adequate heat exchange. The total air-liquid
surface area may be above some minimum level in the
compression/expansion chamber; the trajectory and velocity of the
droplets may such that they have time to reach thermal equilibrium
with the air before they hit a surface or body of liquid; the
distribution of the droplets may need to be uniform so as to avoid
hot/cold spots, etc.
[0067] Note that unlike a conventional refrigeration cycle, very
little of the liquid sprayed into the compression chamber
evaporates. Instead, the temperature typically rises by only a few
degrees. So the great majority of the heat transfer occurs in the
form of sensible heat--that is, by warming the water (or by cooling
it in the case of expansion).
[0068] Certain compressor technologies may take some advantage of
this principle. Some screw compressors, for example, are flooded
with oil or water. Oil is sprayed into the inlets of some
compressors. In a liquid ring compressor, compression vanes are
surrounded by a ring of water formed by centrifugal force.
[0069] In none of these known cases, however, is enough liquid
mixed with the air to effect substantially isothermal compression.
In fact, water inside a compression chamber can create a number of
issues.
[0070] One issue is corrosion. Most conventional compressor
components are made of carbon steel, which is susceptible to rust.
In addition, turbine compressors avoid water because water droplets
can damage the high-speed turbine blades.
[0071] In a reciprocating compressor, the danger is that the
presence of water in the compression chamber will result in
hydro-lock. That is, the incompressible water will be unable to
escape when the piston approaches top dead center, resulting in a
blown head gasket, broken valve plate, bent connecting rod, or
other catastrophic result.
[0072] Embodiments of systems of the present invention may be
comprised of five components:
[0073] an air compression/expansion mechanism;
[0074] a control system;
[0075] compressed air storage;
[0076] heat exchanger(s) or thermal storage; and
[0077] a motor/generator (if the system is electrically
coupled).
[0078] The Compression/Expansion Mechanism
[0079] There are many ways to compress air. There are centrifugal
compressors, lobe compressors, screw compressors, axial turbine
compressors, etc. Any of these approaches may offer viable
technologies for compressed air energy storage according to
embodiments of the present invention.
[0080] Embodiments of energy storage according to the present
invention may exhibit one or more of the following
characteristics:
[0081] compatibility with water sprays;
[0082] mechanically efficiency:
[0083] capability of handling high volumes of air;
[0084] high power density;
[0085] scalability across a broad power range;
[0086] low cost;
[0087] high reliability.
[0088] For very high power requirements--100 MW or
greater--turbines may be well suited. However, spraying liquid into
turbines can damage conventional high-speed rotor blades.
Nevertheless, embodiments of the present invention may utilize
liquid injection into a turbine.
[0089] Other CAES systems have used turbine compressors (without
liquid spray) or hydraulic piston compressors. However, for powers
ranging from hundreds of watts to megawatts, a technology that
embraces many of the above characteristics is the one that is
perhaps the oldest, most highly refined and definitely the most
ubiquitous--the reciprocating piston.
[0090] As a result, certain embodiments of the present invention
exhibit similarities with a conventional multi-stage reciprocating
air compressor. That is, we will use (in most configurations) a
cascade of pistons and cylinders may compress/decompress the air in
stages.
[0091] Single-Stage System Design
[0092] FIG. 11 shows one embodiment of a single-stage
configuration. Some real-world systems will use multiple
stages.
[0093] Compression may occur as follows, as summarized in the
following table:
1. Air is drawn in through the air filter and enters the cylinder
via the suction valve (Step 1). 2. Liquid is sprayed into the
cylinder (or into a pre-mixing chamber upstream of the cylinder) 3.
The resulting air-liquid aerosol is compressed in the cylinder by
upward motion of the piston. The liquid absorbs the great majority
of the heat of compression (Step 2). 4. The compressed mixture is
exhausted from the cylinder via the discharge valve into the
air-liquid separator (Step 3) 5. The compressed air flows into the
compressed air storage tank via the four-way valve, which is in
position 1 6. The separated liquid is pumped through a heat
exchanger, which exchanges heat with an external thermal reservoir
(e.g. the atmosphere, a water tank) in order to return its
temperature to near-ambient--ready for the next spray cycle
TABLE-US-00002 Step 1 Step 3 Refill Step 2 Move compressed cylinder
Compress air to separator Four-way valve Position 1 Position 1
Position 1 Suction valve Open Closed Closed Discharge valve Closed
Closed Open Piston At TDC at At BDC at Between BDC start of step
start of step and TDC
[0094] During step 1, the piston is driven upward by a crankshaft,
by hydraulic pressure, or by some other mechanical means,
compressing the air and liquid mist contained in the cylinder. Step
2 begins when the air pressure inside the cylinder reaches the
pressure inside the pressure cell, at which point the discharge
valve opens, allowing compressed air to flow from the cylinder to
the pressure cell. During step 3, the piston is pulled down by a
flywheel or other, out-of-phase pistons, allowing low-pressure air
to refill the cylinder.
[0095] Expansion may happen similarly, as described below and in
the following table:
1. Compressed air leaves the air tank via the four-way valve, which
is in position 2, and enters the cylinder via the suction valve
(Step 1) 2. Liquid is sprayed into the cylinder (or into a
pre-mixing chamber upstream of the cylinder). In the case of
expansion, the liquid spray adds heat to the process. 3. The
resulting air-liquid aerosol expands nearly isothermally (Step 2)
4. The air-liquid mixture, now at atmospheric pressure, is
exhausted from the cylinder into the air-liquid separator (Step 3)
5. The air is exhausted to the atmosphere 6. The separated liquid
is pumped through a heat exchanger, which exchanges heat with an
external thermal reservoir (e.g. the atmosphere, a chiller) in
order to return its temperature to near-ambient--ready for the next
spray cycle
TABLE-US-00003 Step 1 Step 3 Add compressed Step 2 Move expanded
air to cylinder Expand air to separator Four-way valve Position 2
Position 2 Position 2 Suction valve Open Closed Closed Discharge
valve Closed Closed Open Piston At TDC at start Between TDC At BDC
at start of step and BDC of step
[0096] There are three process details worth noting here:
[0097] 1. One potential benefit of a reciprocating piston
compressor, is that the same system (with some minor modifications)
can be used for both compression and expansion. This saves capital
cost.
[0098] 2. During expansion, only a pre-determined amount of air may
be admitted into the cylinder at the beginning of each stroke. This
amount is the volume of air at the current pressure in the storage
tank, that will fill the total volume of the cylinder when it has
fully expanded.
[0099] For example, in a single-stage system with a one liter
cylinder and 20 atmospheres of pressure in the storage tank, the
amount of compressed air that needs to be introduced into the
cylinder is one twentieth of a liter, or 50 cc's. When the piston
is at the bottom of its stroke, the air will have expanded to 20
times its compressed volume; that is, to one liter. At that point
it will be at one atmosphere pressure. This approach of letting a
pre-determined pulse of air into a cylinder, then letting it expand
fully with the inlet and exhaust valves closed, allows efficient
extraction of substantially the full amount of energy available in
that air.
[0100] As the storage tank depletes, the inlet valve timing may be
changed. The air pressure in the tank drops, and the volume of air
introduced into the cylinder can be increased to maintain a steady
power level. As a consequence, the valve may stay open longer,
helping to maximize efficiency.
[0101] 3. Control over the amount of air let into the cylinder
allows varying the power output by the system. By allowing more air
to enter the cylinder, the system can generate more power--at the
cost of efficiency. If the system has to follow a varying load, the
control system for the valves can adapt to very quickly.
[0102] Multi-Stage System Design
[0103] Now, let's say we are using N stages. We therefore have N
cylinders. For simplicity, let's say that we want the expansion
ratio, r, in each cylinder to be the same. r is therefore the
n.sup.th root of the overall expansion ratio R:
r=.sup.N {square root over (R)}
[0104] The displacement in each successive cylinder increases
exponentially. If V.sub.i is the volume of the i.sup.th cylinder,
and V.sub.f is the total displacement of the system (that is, the
sum of the displacements of all of the cylinders), then:
V i = V f r i j = 1 N r j ##EQU00001##
[0105] As an example, suppose that the total displacement of a
three-stage system is one liter. If the stroke length of each
piston is the same and equal to the bore (diameter) of the final
cylinder, then the volumes of the three cylinders are 19 cm.sup.3,
127 cm.sup.3, and 854 cm.sup.3. The bores are 1.54 cm, 3.96 cm, and
10.3 cm, with a stroke length of 10.3 cm for all three.
[0106] Using the same compression ratio in each stage may offer an
advantage in that each stage generates roughly the same amount of
work, leading to a balanced load on the crankshaft. However, such
an approach can also pose certain design challenges.
[0107] For example, roughly the same volume of water is sprayed
into each cylinder, regardless of its diameter. Specifically, a
same amount of power is put into each stage and, therefore, a same
amount of heat has to be absorbed.
[0108] In practice, there is a limit to the volume fraction of
water than can be introduced. This limits the compression ratio for
the higher-pressure cylinders, assuming that the same .DELTA.T is
wanted in each cylinder. This is typically a valid assumption,
otherwise, the cylinders with higher .DELTA.T's would operate at
lower efficiencies.
[0109] The Spray System
[0110] Although the idea of spraying water into a compression
cylinder in order to absorb the heat of compression (and add heat
during expansion) is conceptually simple, implementing such a spray
system presents certain challenges.
[0111] One challenge is to provide enough thermal mass of water to
absorb all the heat generated. This determines the minimum mass
fraction of water.
[0112] Another challenge is to provide enough surface area to
permit the heat exchange to occur in the time available (about 10
to 15 milliseconds at 1200 RPM). This determines the minimum
droplet size.
[0113] A further challenge is to distribute water droplets
uniformly throughout the chamber. "Dead spots" not only reduce
efficiency but can subject those areas of the cylinder to high
thermal stresses. As is shown and discussed below in connection
with FIG. 5, sprayers can be configured to introduce liquid
droplets to valve pockets to help enhance a uniformity of spray
distribution.
[0114] Still another challenge is to allow droplets to reach
thermal equilibrium between the time they exit the nozzle, and when
they strike a cylinder surface or coalesce with other droplets. The
mean free path/mean flight time is a complex function of nozzle
pressure, volume fraction of water, droplet size, spray
distribution, and other factors.
[0115] Yet another challenge is posed by having spray exiting the
nozzle, break up into droplets quickly. In addition, the pumping
power consumed to force the water through the nozzles must not be
too great, as it represents a parasitic loss.
[0116] Theory
[0117] One size range of interest is for droplet size on the order
of 100 microns. In this regime the droplets reach terminal
velocity, and are therefore entrained in the air within a few
milliseconds. This maximizes their ability to exchange heat with
the air before they hit a wall and coalesce.
[0118] FIG. 12 shows that 100 micron droplets sprayed into the
compression chamber at 20 meters per second decelerate rapidly and
are effectively at zero velocity at 10 milliseconds. If that time
is, let's say, to be on the order of one fifth of the duration of
the compression stroke (which, in turn, is half of the complete
cycle), then the maximum speed of the system is 600 RPM. FIG. 13
indicates that the ability of a droplet to transfer heat is, as
expected, dependent on its diameter.
[0119] One quantity for determination is the amount of water to be
sprayed into the cylinder to achieve a particular .DELTA.T (and,
therefore, a particular thermodynamic efficiency) at a given
pressure. FIG. 14 shows that, in order to keep the temperature drop
to 10.degree. C. as air expands from 300 atm to 150 atm, it is
useful to begin the expansion with a very high volume fraction of
50 micron water droplets--about 20%.
[0120] The volume fraction needed becomes higher still at higher
expansion (or compression) ratios. For this reason, it may be
desirable to have smaller ratios for higher-pressure stages. At
sufficiently high volume fractions, very high efficiencies are
achievable with reasonably-sized droplets, as can be seen in FIG.
15.
[0121] Nozzle Design
[0122] It is desirable that the spray system till the
compression/expansion cylinder as uniformly as possible throughout
the stroke, with a high density of droplets of the desired size,
and with a minimum of pumping losses. Achieving such uniformity
(that is, volume-filling) property has been challenging.
[0123] Hollow-cone nozzles have been developed with a very wide
cone angle (about 150.degree.) to spread the spray across the full
width of the cylinder bore even when the piston is near top dead
center. The technique for achieving the wide angle is to swirl the
spray as it emerges, then deflect it of a small central plate.
FIGS. 16A-B shows the internal geometry of the nozzle and a
simulation of the velocity of the emerging droplets,
respectively.
[0124] When the piston is closer to bottom dead center, nozzles
that are more volume-filling may be used. Certain "shower-head"
type nozzles for that purpose. In some cases, nozzles designed
specifically for the geometry of a particular section of the
cylinder (e.g. the valve pockets) may be used. The effect of swirl
is shown in FIG. 17, along with the droplet breakup. The absence of
a deflector plate produces a narrower cone.
[0125] Spray System Design
[0126] To achieve uniform volumetric distribution, nozzles can be
positioned inside the compression cylinder with locations and spray
patterns such that sufficient spray finds its way into all the
nooks and crannies of the cylinder. An analogy exists with direct
fuel injection in internal combustion engines. Such an approach is
shown in FIGS. 18A-F.
[0127] Alternatively or in conjunction with direct injection,
pre-mixing can be used upstream of the intake valves that entrain
droplets in the incoming air stream. A pre-mixing chamber may have
the velocity of the droplets slow down to match that of the air
before they strike a wall.
[0128] The spray nozzles must be distributed such the sprays don't
overly interfere with each other, which can lead to droplet
coalescence. The droplets may be small enough to remain in the air
flow as it moves through the intake manifold and the intake valves.
Use of a pre-mixing chamber may call for a valve design allowing
straight-through flow of the air-water droplet aerosol created
within the chamber.
[0129] For the highest-pressure cylinder, both a pre-mixing chamber
and in-cylinder nozzles may be used to achieve desired density of
spray.
[0130] Valve Design
[0131] Large gas compressors typically use large-aperture, passive
valves, such as plate valves for gas intake and discharge. These
valves can operate quickly, have high effective flow areas, and are
easy to maintain (the plates are typically hard plastic and can be
replaced in a few minutes if they warp or crack). Passive valves
are held closed via springs, and they open when there is a pressure
differential across them sufficient to overcome the spring force
(typically, very small).
[0132] In choosing the operating parameters of the valve, one
consideration is that the effective flow area (C.sub.v) is high
enough to allow the water introduced by the spray system to exit
during the exhaust stroke. Otherwise, there is the danger of
hydro-lock, discussed above. In practice, the plate valves (and
similar technologies such as poppet and ring valves) work well with
no modifications, even at high compression ratios and high spray
volumes.
[0133] While passive valves have the virtue of simplicity, they can
flutter if the air flow has pulsations, resulting in leakage. In
addition, it may not be easy to directly control when they open or
close. This is acceptable for compression, but not helpful for
expansion.
[0134] In order to allow a certain amount of air into the cylinder
during expansion, and then close the valve, an active valve may be
used (at least for the intake).
[0135] Active valves are more complex, in that they require some
actuation mechanism, and some fail-safe device to handle the
situation in which the valve doesn't open when it should.
Otherwise, hydro-lock is a real danger.
[0136] However, active valves offer the advantage of direct
control. This allows us to experiment with the timing--opening a
little early, for example--in order to improve efficiency. This has
been a component of performance tuning of internal combustion
engines (another type of engine operated using active valves)
throughout the long history of that technology. However, it is not
typically used on air compressors.
[0137] If a valve is used in conjunction with a pre-mixing chamber,
the flow path through the valve, when open, should not overly
obstruct the passage of the droplet-laden air. Plate valves
typically exhibit circuitous paths that are difficult for water
droplets larger than a few microns to navigate successfully.
[0138] To address this problem, we have designed rotary valves with
a large, unobstructed opening(s). One example is illustrated and
described below in connection with FIGS. 28-30.
[0139] Materials Issues
[0140] Pressure vessels, cylinders, pulsation bottles and other
system components of big compressors are typically made of carbon
steel, and are therefore susceptible to corrosion when in regular
contact with water. One approach, of course, is to use materials or
liners that are less susceptible to corrosion. Examples include but
are not limited to aluminum, brass, PVC, etc.
[0141] For large compression cylinders, however, steel is likely to
be the most cost-effective and practical material. To use it with
water, some form of anti-corrosion coating is required. A
nickel-polymer coating protecting against both against corrosion
and wear, may be useful.
[0142] Large compressors typically use oil-lubricated steel piston
rings. However, the presence of oil in water may result in
undesirable emulsification.
[0143] Non-lubricated versions of compressors are also common, and
such systems typically use piston rings made of Teflon or PEEK,
impregnated with materials such as molybdenum disulfide, carbon, or
brass that act as solid lubricants. Embodiments of the present
invention may employ cylinder and rings designed for non-lubricated
operation. Examples include MoS.sub.2 and brass-impregnated Teflon
rings that have exhibited low wear and low leakage.
[0144] Control System
[0145] Unlike conventional engines and compressors, the timing and
control of embodiments of systems according to the present
invention, may be managed electronically. This allows, for example,
maximizing expansion efficiency by varying V.sub.0 as the pressure
in the storage tank changes, as discussed above. The high-pressure
inlet valve is simply opened for a longer time as the tank
depletes.
[0146] Embodiments of rotary valves may use stepper motors that are
phase-locked to the crankshaft. Pressure, temperature, humidity,
and/or torque may be monitored during operation, and fine timing
adjustments made as required to maximize efficiency.
[0147] The operating characteristics of the system--for example,
the power output--can be determined by timing of the valves and
their flow areas. Balancing pressure ratios among multiple stages
can be effected by changes in the valve parameters as well.
[0148] To the first order, these parameters can be calculated based
on the system simulation and empirical performance characteristics.
However, real-time adjustments may be needed because of changes in
the tank pressure during operation and the ambient conditions.
[0149] The basic approach for controlling an individual valve is to
monitor three pressures: the pressure upstream of the cylinder, the
pressure downstream the cylinder, and the pressure in the cylinder.
For compression, the first-order timing computation uses only two
pressure values are needed. The valve is closed when there is a
pressure difference across it. The valve opens when the pressure on
either side approaches some small value (a fraction of a psi,
typically). The first-order timing may be adjusted to maximize
efficiency, for example by advancing the opening of the inlet
valves.
[0150] The exhaust valve timing for expansion is similar to that
for compression. The inlet valve will open, again, when the
.DELTA.p across the valve is near zero. However, the inlet may
close when the crank angle is such that the percentage of the
cylinder filled is the quotient of the downstream and upstream
pressures (taking the dead volume into account).
[0151] A processor may perform these computations, and the inlet
valve timing for expansion, and possibly others, can be adjusted
using a PID (proportional-integral-derivative) controller to
fine-tune performance. This is particularly true in multi-stage
systems.
[0152] Another example of the possible use of a PID controller is
for liquid pumps. For example, FIG. 31A shows an embodiment of a
compression system including a non-positive displacement
(centrifugal) transfer pump in fluid communication with a positive
displacement multi-stage water pump. Flows of liquid from the
transfer pump to the multi-stage water pump utilize a
Proportional-Integral-Derivative (PID) loop around the transfer
pump as shown. The PID loop is configured to maintain a target
pressure (or other parameter such as flow rate) into the
multi-stage water pump.
[0153] Compressed Air Storage
[0154] Typically, high-pressure air is stored in seamless metal
cylinders no larger than about 100 liters. Such cylinders will be
fine for small facilities (e.g. storage for a residential-scale
solar array). However, for utility-scale volumes larger cylinders
may be fabricated.
[0155] One low-cost solution is to utilize seamless steel pipe of
the sort used for compressed natural gas pipelines. Such pipe (X60
grade is a good choice) is available inexpensively, and it's
relatively easy to machine. It would be spun, heated and necked to
form long tanks; or it could be welded on-site into a very long
pipe. Tanks with an outside diameter of 75 cm, a length of 12
meters and 20 mm wall thickness can hold 300 atmospheres (4500 psi)
safely. A megawatt-hour would require six such tanks, which would
cost about $50K. Costs of manufacturing, coating, valves, a
manifold, and a container--may increase that cost.
[0156] In the US, power plant pressure vessels need to follow the
ASME section VII standard if they're installed permanently. Air
cylinders designed to be transportable must follow Department of
Transportation codes. Various ISO pressure vessel codes tend to be
used overseas.
[0157] External Heat Exchange
[0158] The design of heat exchangers can vary with the
installation. If the only available thermal reservoir is the
atmosphere, a conventional air-cooled heat exchanger or cooling
tower will typically be appropriate. The same heat exchanger can
generally be used both to remove heat from the spray water during
compression or to add it during expansion.
[0159] Note that at least some of the water passing through the
heat exchanger will be at a pressure higher than ambient (an
exception is the water used in the final expansion stage).
Embodiments of the present invention may keep the water supplies
for each stage separate, so that pipes carrying the water removed
by the air-water separator in each stage will be at different
pressures.
[0160] This may rely upon a two-step heat exchange process. First,
the pressurized water passes through a shell-and-tube heat
exchanger, transferring the heat to water (or other heat exchange
fluid) in the shell, which is at atmospheric pressure. That fluid
traverses the shell-and-tube heat exchangers for all the stages,
then travels to an air-cooled heat exchanger to release the heat to
the atmosphere.
[0161] The .DELTA.T is small (in order to run close to isothermal),
so that the heat exchanger's job is relatively easy, provided that
the exit temperature is some distance away from ambient. This
implies that the fluid in the system should be circulated for some
cycles until it reaches a set temperature above ambient, at which
point it is shunted through the air-cooled heat exchanger. This is
the same approach used in a car radiator.
[0162] The reverse is true for expansion. However, the operating
temperature must not be below the freezing point of the water or
the heat exchange fluid.
[0163] A variant on this approach is to replace the air-cooled heat
exchanger or cooling tower with an insulated water storage tank.
The idea is that the energy in the heated water can be recovered
during expansion. Of course, this approach only makes sense if the
interval between energy storage and delivery isn't too long. A
storage tank is likely a lower-cost solution than a heat exchanger,
but it will have a larger footprint.
[0164] For example, FIG. 11 shows an embodiment wherein thermal
energy may be communicated to the system through a heat exchanger
in direct thermal communication with the volume of separated
liquid. Such direct thermal communication between the injected
liquid and a heat source/heat sink is not required, however.
[0165] According to alternative embodiments, thermal storage could
take the form of a second stored liquid in thermal communication
with the injected/separated liquid. FIG. 33 illustrates one
embodiment of a system utilizing insulated storage tanks for
liquid.
[0166] In particular, system 3300 is similar to that shown in FIG.
11, comprising tube-in-shell heat exchanger 3301 having tube 3303
configured to receive separated liquid from air-water separator
3305. Hot insulated thermal storage tank 3302 and cold insulated
thermal storage tank 3304 are in selective fluid communication with
shell 3306 of heat exchanger 3301 through transfer pump 3314 and
3-way valves 3310, 3312.
[0167] In the compression mode of operation, heat from the
separated liquid is pumped to hot insulated thermal storage tank
3302 for storage and later reuse in expansion mode (described
below). Cold liquid is replenished in the shell by drawing from the
cold insulated storage tank.
[0168] In the expansion mode of operation, hot liquid is
replenished in the shell by pumping from the hot insulated storage
tank 3302. Coolness from the separated liquid flows to the cold
insulated thermal storage tank 3304 for storage and later reuse in
compression mode.
[0169] Effective for improving the efficiency of air expansion, is
access to a source of waste heat, even low-grade heat such as hot
water from a steam condenser or liquid warmed by a solar collector.
High temperatures are not necessary. Even 80.degree. C. water will
increase the work done by a factor of about 1.2.
[0170] In applications where storage is located near the point of
demand, the thermal characteristics of the system can be leveraged.
Heat is generated during compression that can be used for any
suitable low-temperature heat application (e.g. space or water
heating). During power delivery, cool air and cool water are
generated that can be used to supply cooling by coupling with an
HVAC system or chiller. Air or water of any desired temperature can
be supplied by suitable adjustments to the spray and other
subsystems.
[0171] A larger .DELTA.T will reduce efficiency slightly. However,
this will often be a favorable trade-off for a reduction in heating
or cooling load--which would otherwise consume electricity.
[0172] Thus FIG. 33 shows an embodiment wherein the stored liquid
may be placed into thermal communication with an appropriate heat
source (H.S.) or cool source (C.S.) in order to maintain or even
further change its temperature. Liquid stored in a cold insulated
thermal storage tank could be in communication with a heat sink
such as a naturally-occurring body of water (e.g. lake, river, or
ocean), or an artificial heat sink such as a cooling tower.
[0173] Liquid stored in a hot insulated thermal storage tank could
be in communication with a heat source, which could be naturally
occurring (e.g. solar or geothermal) or artificial (e.g. an
industrial process, building environment, and/or internal heat from
elements of the system itself). For example, heat generated
internally by the energy storage system may also comprise a source
of thermal energy that may be captured and stored.
[0174] FIG. 37A thus shows a number of thermal sources arising from
the compression of gas for storage in one possible embodiment, and
the flows of energy from these thermal sources to thermal storage.
In the particular embodiment of FIG. 37A, two separate thermal
storage units are employed. These units are maintained at different
temperatures for purposes of efficient storage, but this is not
required and in some embodiments only one thermal storage unit may
be used. FIG. 37B shows a number of thermal sources arising from
the expansion of compressed gas, and the flows of energy from these
thermal sources to storage (again a high temperature storage and a
low temperature storage). Such collection of internal heat for
storage may be effected through a circuit employing a circulating
heat-absorbing fluid such as oil, water, or air. The heat-absorbing
fluid could be circulated by a pump.
[0175] Returning to FIG. 33, while the particular system shown in
that figure features a tube-in-shell heat exchanger, this is not
required. Alternative embodiments could utilize other forms of heat
exchangers. For example, a counter-flow heat exchanger could be
employed to efficiently transfer heat between the separated liquid
flowed for spraying, and the circulated liquid being flowed to
thermal storage.
[0176] And while FIG. 33 shows the use of separate thermal
insulation tanks for hot and cold liquids, this is not required.
FIG. 34 shows an alternative embodiment of an energy storage system
3400 comprising a single thermal tank 3402 having an insulated
partition 3404 moveable to define chambers 3406 and 3408 for
containing hot and cold liquid respectively.
[0177] In the compression mode, the partition moves to the right as
hot liquid accumulates and cold liquid is depleted. In the
expansion mode, the partition moves to the left as cold liquid
accumulates and hot liquid is depleted.
[0178] The thermal storage approaches of FIGS. 33 and 34 offer a
potential benefit, in the ability to retain and utilize thermal
energy available from heat exchange with expanding gas or gas being
compressed. These thermal storage approaches also substantially
reduces the volume of liquid maintained at high pressure.
[0179] In particular, the separated liquid is elevated to a high
pressure by virtue of any compression process. Rather than
incurring the expense of storing this liquid at an elevated
pressure, thermal storage can instead be achieved by storing the
thermal energy of a second liquid at a lower pressure.
[0180] FIG. 35 shows another embodiment which also realizes the
advantage of reduced volumes of liquid that are required to be
maintained at high pressure. The system 3500 does not necessarily
employ thermal storage, but does utilize heat exchange between a
pressurized separated liquid, and a second liquid in thermal
communication with the separated liquid but which may be stored at
a lower pressure. Depending upon the state of the liquid
circulation system 3502, the second liquid may be in thermal
communication with a heat source (H.S.) 3504 or a cool source (C.S)
3506 through respective heat exchangers (which may be of the
tube-in-shell type).
[0181] FIGS. 33-35 depict embodiments employing gas flow valves
dedicated to suction and discharge roles in both the compression
and expansion cases. However this is not required, and alternative
embodiments could employ configurations having gas flow valves
dedicated to high- and low-pressure sides, and whose role (e.g.
suction or discharge) changes between compression and
expansion.
[0182] Accordingly, FIG. 36A depicts an alternative embodiment
wherein thermal energy from separated liquid may be transferred to
a second liquid, whose storage at low pressures may afford
increased efficiency in operation. Energy storage system 3600
comprises reversible compressor/expander 3602 comprising piston
3604 reciprocating within chamber 3606.
[0183] A first gas-liquid separator 3610 is configured to receive a
compressed gas-liquid mixture flowed from reversible
compressor/expander 3602 through dedicated high pressure side valve
3612. A second gas-liquid separator 3614 is configured to receive
expanded gas-liquid mixture flowed from reversible
compressor/expander 3602 through dedicated low pressure side valve
3616.
[0184] Liquid heated by compression and separated by gas-liquid
separator 3610 on the high pressure side, is flowed through
three-way valve 3611 to heat exchanger 3630, where it exchanges
heat with second liquid circulated by pump 3632 from cool liquid
tank 3634 to hot tank 3638. Liquid cooled by expansion and
separated by gas-liquid separator 3614 on the low pressure side, is
flowed through three-way valve 3611 to heat exchanger 3630 where it
exchanges heat with second liquid circulated by pump 3636 from hot
liquid tank 3638 to cool liquid tank 3634. The hot/cold liquid
tanks could be in thermal communication with a heat source/heat
sink to maintain their temperatures or even to increase/decrease
their liquid temperatures, respectively.
[0185] In the specific embodiments of FIG. 36A, the
compressor/expander stage comprises valves 3616 and 3612 that are
dedicated to being proximate to the low- and high-pressure sides
respectively. That is, one side of the valve 3616 consistently
experiences a relatively low pressure, while one side of the valve
3612 consistently experiences a relatively high pressure. Thus
depending upon the particular function (compression or expansion),
the valves 3616, 3612 may serve to intake gas or exhaust a
gas-liquid mixture from the chamber.
[0186] The energy storage embodiment of FIG. 36A may offer certain
benefits in terms of valve design and performance. Specifically, in
the particular embodiment shown in this figure, the valve dedicated
proximate to the high pressure side, may have an opening that is
smaller than the valve that is dedicated proximate to the low
pressure side. This is because the increased density of the higher
pressure gas expected to flow through the valve 3612, may call for
a smaller valve area than valve 3616 flowing a less dense gas a
lower pressures.
[0187] In other embodiments the high pressure side valve could have
a larger relative area. Overall, the valve areas may be sized to
reduce overall pumping losses attributable to pressure drops across
valves. The relative valve sizes (that is the ratio of effective
valve areas) may be greater or less than one in order to achieve
this goal.
[0188] Using dedicated valves with different valve areas can allow
the valve to the high-pressure side to be more compact. Such a
scheme can also enhance efficiency by reducing valve losses
attributable to effects such as free expansion, because the area of
the valve is designed to match the expected pressure of the gas
flowing therethrough.
[0189] A possible benefit of the approach of the embodiment of FIG.
36A, is that a valve always experiences a pressure differential in
the same direction. That is, one side of the valve is at high
pressure relative to the other side of the valve, simplifying
design and construction of the mechanism used to actuate the valve
as appropriate during compression or expansion.
[0190] FIG. 36B depicts an alternative embodiment utilizing a
thermal sink and thermal source. This embodiment 3690 is similar to
that of FIG. 36A, except there is essentially an unlimited amount
of heat (as with geothermal) or cooling (as with a body of water)
available. In that case, the hot source would replace 3638 and the
cold source would replace 3634. After passing through the heat
exchanger, water would be ejected elsewhere (perhaps in a lake far
from the point where water is taken from).
[0191] Round-Trip Efficiency
[0192] FIG. 19 shows schematically where the energy goes during the
complete compression/expansion cycle. The full-cycle, or
round-trip, efficiency depends in part on the way in which the
system is coupled to the grid.
[0193] In the chart, it is assumed that mechanical power is taken
directly from a rotating shaft. If the source of energy is
electricity, a motor would be used to convert that electricity to
mechanical energy. That would reduce the round-trip efficiency by a
factor of about 0.95 (just as a motor at the end of the pipeline
reduces the overall efficiency by the same factor).
[0194] As mentioned above, a boost in round-trip efficiency can be
achieved by scavenging waste heat from some other process (e.g.
solar thermal) during the expansion phase. With enough waste heat
energy, it's possible to overcome all the other drains on the
full-cycle efficiency and to deliver more work than was originally
stored. Of course, energy is still being conserved: the energy
originally stored in the compressed air is being augmented with
some of the waste heat's energy.
[0195] Certain system design trade-offs may affect efficiency. A
smaller .DELTA.T improves efficiency, but it also means that a
larger heat exchanger will be required, increasing the cost of the
system. Similarly, running the system at higher speed will produce
higher power but will require faster heat transfer, adding to
system cost.
[0196] For some applications--a UPS, for example--we'll be less
concerned with efficiency than with capital cost. Other
applications may most often require high efficiency, but, at times
of peak demand, need the highest possible power output. Embodiments
of the present invention offer an ability to control temperature by
varying the amount of fluid sprayed into the cylinders during
operation, and also to vary the power output by adjusting valve
timing, thereby allowing the efficiency/power trade-off to be
adjusted in real time.
[0197] In summary, embodiments according to the present invention
may share one or more of the following features.
[0198] The system may use a temperature-controlled cycle utilizing
a reversible compression/expansion mechanism. Near-isothermal
operation addresses the low-efficiency problem that has limited the
broad adoption of compressed air energy storage. Near isothermal
operation, along with control of other parasitic losses, can
provide system efficiencies comparable to large-scale batteries at
lower cost and with a much longer lifetime.
[0199] Liquid may be injected into the cylinder in the form of
small droplets, which mixes with the air during both compression
and expansion. This facilitates heat exchange and enables
near-isothermal operation.
[0200] Precision electronic valve timing may manage the trade-off
of efficiency for power in real time. This allows the system
operation to match demand and to extract useful power over a wide
of tank pressures.
[0201] Two-step expansion may be used. A pre-determined pulse of
air, V.sub.0, is let into the cylinder, at which point the inlet
and exhaust valves are closed and the air is allowed to expand.
This allows maximum possible work from compressed air, and it also
allows varying the power output by the system.
[0202] Waste heat added during expansion can be used to
substantially improve efficiency.
[0203] Current Development
[0204] A prototype capable of storing and recovering about 30
minutes of energy at about 60 kW is being developed. Tasks
associated with that project include but are not limited to:
1. Development, testing, and optimization of a single-stage
compression/expansion mechanism 2. Maximizing heat exchange inside
the cylinder (that is, mixing the incoming air optimally with the
liquid inside the pressure cells used to facilitate heat exchange)
3. Control and monitoring (valve timing, optimization of cycling
time, logging of system behavior)
[0205] Single-Stage System Configuration
[0206] In an embodiment, the system includes a 200 horsepower motor
(which acts as a generator during expansion), an air intake line to
supply ambient air to the compression cylinder, a water line to
supply water to the spray nozzles, and an air-water separator. FIG.
31 shows a schematic of an embodiment of a system.
[0207] FIG. 20 shows the compression cylinder is 15.25'' in
diameter with a 3.5'' stroke that is double-acting. Eight spray
manifolds are arranged in a circle on each head. Different nozzles
can be swapped in and out each manifold. The hoses, shown in
magenta in FIG. 21, supply water to the nozzle manifolds.
[0208] The compression cylinder is coated with nickel-polymer to
prevent corrosion. FIG. 22 shows that various sensors can be seen
in the valve covers and in the intake and exhaust manifolds.
[0209] The Spray System
[0210] FIG. 23 shows the cylinder with the cover removed,
indicating how the water is distributed to the nozzle manifolds.
FIG. 24 shows one configuration of the outboard cylinder head. Five
of the manifolds have three nozzles each, one has a single nozzle,
another holds a pressure sensor, and one has a rupture disk as
fail-safe to protect the cylinder if a pressure spike occurs. FIG.
25 shows the above head configuration running at about 0.3
liters/sec total flow rate and 100 psi pump pressure.
[0211] Control and Data Logging
[0212] Data from about 50 sensors--pressure, temperature, air and
water flow rates, electrical current, crank angle, and shaft
torque--can be displayed as shown in FIG. 26. FIG. 31 indicates
where the sensors are positioned in that system embodiment. There
is a safety system that monitors a number of critical values (such
as motor vibration) and can shut the motor down automatically when
a problem occurs. Data can be logged and stored at 4 kHz.
[0213] A variable-frequency drive (VFD) allows control over the
motor speed. Our experiments to date have varied the speed from 300
RPM to 1200 RPM.
[0214] Spray nozzles can be turned on and off individually, and the
pump pressure can be set to any value below 500 psi. Water flow can
be set to be straight through (that is, pumped from a storage tank,
through the cylinder and out to a drain) or re-circulated through
an air-cooled heat exchanger and back into the spray pump.
[0215] Other parameters that can be varied are the upstream and
downstream pressures. That is, the incoming pressure to the
cylinder can be set when performing expansion as well as the
outgoing pressure (vented to the atmosphere through a muffler)
during compression. This has allowed us to perform tests at
compression ratios ranging from less than 2:1 to about 6:1. Because
of the high "dead volume" (that is, the un-swept volume remaining
in the cylinder when the piston is at top dead center), higher
compression ratios may need to be tested by another cylinder.
[0216] Results
[0217] The PV diagram of FIG. 27 shows the results of a typical
run. The speed is 900 RPM and the compression ratio is about 4.5:1
(53 psi). Water is being sprayed in at about 0.3 liters/second. The
data is displayed on a log-log scale so that the polytropic index
can be measured as the slope of the diagonal lines.
[0218] Briefly, the polytropic index is a measure of how close to
isothermal the process is. The work done by the expanding air is
derived by integrating the expression W=-.intg.p dV over all the
stages. When the valve is first opened, neglecting some small
transients, the pressure is constant in the limit of a large tank.
Hence the work done in this filling stage is:
W.sub.fill=(p.sub.0-p.sub.back)V.sub.0,
where: V.sub.0 is the volume of the expansion chamber at the moment
the valve is closed, and p.sub.back is the back pressure.
[0219] Afterward, the expansion is polytropic (that is,
pv.sup.k=const and k is called the polytropic index), through N
stages.
[0220] Neglecting back pressure, by integration, the work done in
each stage i is:
W i = p 0 V 0 k - 1 [ 1 - ( 1 r ) k - 1 ] ##EQU00002##
Hence over N stages, it is:
W N = N p 0 v 0 k - 1 [ 1 - ( 1 r ) k - 1 N ] ##EQU00003##
[0221] However, back pressure is to be accounted for. Working
against the back pressure during polytropic expansion requires
additional work:
W.sub.back=p.sub.backV.sub.0(R-1)
That is, simply the back pressure multiplied by the swept
volume.
[0222] Adding these terms gives the work done per `stroke`:
W total = ( p 0 - p back ) V 0 + N p 0 V 0 k - 1 [ 1 - ( 1 r ) k -
1 N ] - p back V 0 ( R - 1 ) ##EQU00004##
[0223] Running the system close to isothermal makes a difference in
the amount of work is required to compress a volume of air to a
pressure p.sub.0 (or that can be gotten out of air stored at a
given p.sub.0).
[0224] In the polytropic model (which is the usual one for engine
design):
pv.sup.k=const,
where: p=pressure, v=volume, and k is the polytropic index, a
measure of how much heat is transferred in or out of the system as
it operates.
[0225] k=1 for isothermal compression and expansion (that is,
operation during which the system remains at a constant
temperature). The polytropic index for pure adiabatic
compression/expansion is about 1.4.
[0226] The work done per piston stroke for a system with N stages,
with expansion at constant pressure during the filling stage
is:
W total = ( p 0 - p back ) V 0 + N p 0 V 0 k - 1 [ 1 - ( 1 r ) k -
1 N ] - p back V 0 ( R - 1 ) , ##EQU00005##
where: r is the expansion ratio in a given stage; R is the total
expansion ratio; p.sub.back is the back pressure; and V.sub.0 is
the initial volume.
[0227] Taking the ratio of computed work output to theoretical work
output, we can plot the efficiency of the system as a function of
the polytropic index. This is shown in FIG. 32.
[0228] In a multi-stage compressor, the discharged air is assumed
to be cooled back to ambient temperature between stages, improving
efficiency. In the limit of an infinite number of stages,
compression would occur isothermally.
[0229] Returning to FIG. 27, the compression cycle is represented
in a PV diagram by a loop (here dashed). The bottom, relatively
horizontal, line is the piston moving down with the intake valve
open, drawing in atmospheric air.
[0230] At the lower rightmost point, the piston starts to move
upwards and the intake valve closes. Air is compressed nearly
isothermally--a slope of 1.0 would represent isothermal
compression--the slope here is 1.1.
[0231] When the pressure reaches that of the air tank, in this case
about 53 psig, the exhaust valve opens (upper right-hand corner of
the diagram). The pressure stays constant as the compressed air is
pushed out of the cylinder, until the piston reaches top dead
center (upper left-hand corner).
[0232] There is still compressed air in the cylinder, filling the
dead volume. This air drops in pressure, again nearly isothermally,
as the piston descends. Once the air in the cylinder reaches
atmospheric pressure, the intake valve opens, completing the
cycle.
[0233] Active Valve Development
[0234] Compression experiments were performed using passive
valves--plate and ring valves, specifically. However, for expansion
experiments, these may be replaced with an active valve developed
in-house
[0235] One embodiment of an active valve is shown in various views
in FIGS. 28-30. The custom valve uses a stepper motor to spin a
disk with pie-shaped ports. The rotation of the disk is
phase-locked to the compressor shaft rotation. A second plate can
be rotated using a second, smaller motor to close off part of the
aperture.
[0236] By changing the phase of the rotating disk relative to the
crankshaft, the valve can be made to open earlier or later in the
cycle. The position of the aperture plate then determines how long
the valve will be open.
[0237] Both the phase and the aperture opening are electronically
adjustable. This allows modification of the valve timing to
maximize efficiency as well as to operate at different tank
pressures.
[0238] Development
[0239] Successive prototypes may add stages. That is, they operate
at higher pressures.
[0240] The single-stage system will compresses and stores air at
about 6 atmospheres (90 psig). This relatively low number is
because the off-the-shelf cylinder being used for the single-stage
prototype has a high "dead volume". That is, there is still a
volume of air in the cylinder even when the piston is at top dead
center--about 20% of the total volume isn't swept by the
piston.
[0241] Some of that dead volume is filled by liquid water which
collects after the first few strokes and remains in the cylinder.
The dead volume that remains, limits the compression ratio.
[0242] Certain units may use custom-designed cylinders with much
lower dead volume, perhaps 5%. This will allow us to operate a
higher compression ratio and still get high air flow. One
embodiment may achieve compression ratios of about 10 for the first
stage, 6 for the second stage, and 4 for the third stage, resulting
in a final compression ratio of 240, or 3600 psi.
[0243] The table below summarizes or goals for upcoming milestones.
The numbers shown are percentages of the input electrical power
lost via various mechanisms--waste heat, leaks, friction, etc. The
losses shown are one-way. The losses are multiplicative. That is,
our one-stage system loses 10% of the input power as heat, then 5%
of the remaining 90% (that is, 4.5% of the input power) via piston
ring leakage. Similar losses are expected for compression and
expansion, so that each loss has to be counted twice. The product
of all the losses is therefore squared to compute the round-trip
efficiency. The "Typical" column represents losses for a
conventional (non spray-cooled) air compressor.
TABLE-US-00004 Category Loss Channel Typical 1 Stage 2 Stage Alpha
Production Notes Thermodynamic Thermal 40% 10% 4% 3% 2% 1 loss
Leaks Piston rings 2 5 3 2 1.5 Packing 0.5 0.5 0.5 0.5 0.5 Valves 3
4 3 2 1 2 Pressure drops Valves 1.5 1.5 1.5 1 1 2 Heat exchanger
0.5 1 0.5 0.5 0.5 Air-water 0 0.5 0.5 0.5 0.5 3 separator Filters
0.5 0.5 0.5 0.5 0.5 Nozzles 0 1.5 1.5 0.5 0.5 3 System piping 0.5 1
1 0.5 0.5 Friction Frame 1 5 2.5 1 1 4 Oil pump 1 2 2 1 1 Piston
rings 3 3 3 2 1.5 5 Spray Pumps 0 2.5 1.5 1 0.5 3 Piping 0 1 1 0.5
0.5 3 Electrical Motor/generator 4 8 6 5 4 6 Valve actuation 0 2 1
1 0.5 2 Power electronics 1.5 6 5 2 1 6 Heat exchanger 1 1 1 1 1
fan Control system 0.5 1 0.5 0.5 0.5 Miscellaneous System 1 2 1 1
0.5 One-way 39% 55% 63% 74% 84% Efficiency Round-trip 15% 30% 40%
55% 70% Efficiency Note 1: The thermal loss shown in the "Typical"
column is for an air compressor with a 6:1 compression ratio -
about what the one-stage prototype can do. Later prototypes will
have higher compression ratios, so this loss would get even worse
in the absence of spray cooling. Note 2: Conventional air
compressors use passive plate valves that leak due to fluttering
and incomplete seating of the valve plate. Sealing of our active
rotary valves is accomplished via a mechanism similar to piston
ring sealing. This should, ultimately, minimize leaks. Note 3: The
water spray system is absent in a conventional compressor. Losses
associated therewith can be reduced over development generations.
Note 4: Frame friction is high for the one and two-stage systems
because only two cylinders are mounted on a frame designed for four
cylinders. Note 5: Ring leakage and friction represents an area of
research, with relatively little investigation by previous
researchers. Note 6: Prototypes use a VFD to allow the motor to run
at different speeds. Production units may be optimized to run at a
single speed. This will reduce the cost of the power electronics
and permit the motor and frame to operate at their optimal
performance points, maximizing efficiency.
[0244] Costs
[0245] Embodiments of compressed air energy storage systems
according to the present invention may comprise at least two major
sub-systems: 1) the air compressor/expander mechanism, and 2) the
air storage sub-system. Because the job of the former is to store
and deliver power at a certain rate, it is best defined in terms of
how many kilowatts (or megawatts) it can handle.
[0246] The air storage system holds energy. So it is best defined
in terms of how many kilowatt-hours or megawatt-hours it can
contain. Therefore, to correctly cost a system, how much power to
be delivered for how long needs to be specified.
[0247] One embodiment, for example, will store and deliver 500 kW.
If it's desired to deliver that rated power for, say, five hours,
it will need 2500 kWhrs of storage. The cost of that system will be
the sum of the costs of a 500 kW compressor/expander and 2500 kWhrs
of air storage.
[0248] The cost per kilowatt-hour is based almost exclusively on
the cost of the air storage tanks. Conventional steel compressed
gas cylinders offer a straight-forward solution for
commercial-scale applications. At 250 atm and an expansion
efficiency of 85%, about 40 liters of compressed air is required to
deliver a kilowatt-hour. High-strength steel tanks that hold 110
liters at 250 atm cost about $430 in quantity. That works out to
about $155 per kWh. To this needs to be added the cost of a
container, mounting brackets, valves, manifolds, gauges, etc.; plus
the costs associated with certification. $200 per kWhr probably
represents a reasonable estimate for this approach.
[0249] Other options exist for storing large quantities of
compressed air. For large-scale applications, underground storage
is likely the least expensive solution--but the exact cost is
highly location-specific. Seamless steel pipes are another approach
for megawatt-scale storage. X65 steel (commonly used for natural
gas pipelines) have roughly half the tensile strength of the CrMo
steel used in the gas storage tanks mentioned above, but costs
proportionately less. Welding X65 (or similar) pipes together to
form a large storage unit, especially if installed underground as
is done with natural gas pipelines, could be a more economical
solution than packing many small air tanks into a shipping
container.
[0250] The cost per kilowatt of power delivered is driven by the
cost of our expansion/compression mechanism. That cost per kilowatt
of power delivered is more complex to compute, as there are a
number of different components. There are also trade-offs to be
made between the triad of cost, efficiency, and power output. For
example, running slower improves the thermal efficiency, but
reduces the power output. A bigger heat exchanger improves
efficiency but costs more.
[0251] Estimate for a typical 500 kW system are shown in the table
below:
TABLE-US-00005 Unit #1 Cost Unit #500 Cost Component ($000) ($000)
Compressor frame 175 100 Custom cylinders 45 N/A Spray system 20 10
Valves 50 25 Heat exchange/Thermal store 50 30 Air-water separator
30 20 Piping, misc. pressure vessels 80 60 Motor-generator 50 40
Control and power electronics 50 30 Balance of system 50 30
Compressor/expander Total 600 ($1200/kW) 345 ($690/kW) Air storage
600 450 (6 hours = 3 MWHrs) System Total 1200 ($400/kWhr) 795
($265/kWhr) Lithium-ion battery system 3000 ($1000/kWhr) 2250
($750/kWhr)
[0252] The cost of units is expected to be cost-reduced in a number
of areas compared with the initial production units. Most
significantly, it will likely be built on a custom compressor frame
that integrates the cylinders into the engine block.
[0253] Note that the "System Total" line is divided by the number
of kilowatt-hours stored to give a system cost per kilowatt-hour
for a six-hour system. This allows a more or less direct comparison
with the cost of a comparable battery system, which is typically
quoted as a price per kilowatt-hour.
[0254] Mechanical Versus Hydraulic Compression
[0255] Other approaches to air compression for energy storage
purposes have been proposed. One such approach that is being
developed is hydraulic or "liquid piston" compression. An
embodiment of such a system is shown in FIG. 3 of U.S.
Nonprovisional patent application Ser. No. 13/010,683 filed Jan.
20, 2011 ("the '683 Application"), incorporated by reference in its
entirety herein for all purposes.
[0256] In a liquid piston system, hydraulic fluid is pumped from
one pressure vessel into another via a hydraulic motor, compressing
the air inside the second vessel. At the end of the compression
stroke, the first vessel will be empty of fluid (but filled with
air drawn in at atmospheric pressure), and the second vessel will
be filled with fluid.
[0257] In the FIG. 3 of the '683 Application, the four-way
hydraulic valve shown at the lower left of changes state, and the
flow direction reverses, compressing the air in the first pressure
vessel. Expansion works in a similar fashion, with the compressed
air in the air storage tank expanding to push fluid through the
hydraulic motor, turning a generator.
[0258] This is a viable approach to air compression, with the
virtue that the same mechanism can be used for expansion as well as
compression. Such approaches, however, may face challenges in
providing sufficient efficiency and power density. For example
hydraulic motors, generally speaking, aren't as efficient as
reciprocating mechanical engines.
[0259] Moreover, columns of fluid can't move as fast as mechanical
linkages. In particular, the force of gravity may impose limits
upon the efficient transfer of power utilizing a hydraulic system.
For example, downward acceleration of a liquid column in excess of
the pull of gravity, can give rise to non-laminar flow in the
column that degrades efficient transfer of power.
[0260] This is the reason mechanical crankshafts and pistons are
used where power density is important (in applications like
automobile engines for example). To deliver a lot of power, a lot
of air needs to be moved. This can present problems for system
unable to move that air quickly.
[0261] To deliver a megawatt, for example, an expander would
exhaust about three cubic meters of air per second. That would
involve moving a same volume of hydraulic fluid from one pressure
vessel to another via the hydraulic motor, likely requiring a
large/cumbersome system.
[0262] Further Description
[0263] Included in the '683 Application is a description of various
types of energy storage systems utilizing compressed gas as a
storage medium. In certain embodiments, energy stored in the
compressed gas is recovered by expansion of compressed gas in a
cylinder device, to drive a moveable member (such as a
reciprocating piston) that is in physical communication with a
generator through a linkage (such as a crankshaft).
[0264] FIG. 8 shows a simplified view of one embodiment of such a
compressed gas energy system. In particular, the system 800
includes a compressor/expander 802 comprising a cylinder 804 having
piston 806 moveably disposed therein. The head 806a of the piston
is in communication with a motor/generator 808 through a piston rod
806b and a linkage 810 (here a crankshaft).
[0265] In a compression mode of operation, the piston may be driven
by the motor/generator 805 acting as a motor to compress gas within
the cylinder. The compressed gas may be flowed to a gas storage
tank 870, or may be flowed to a successive higher-pressure stage
for additional compression.
[0266] In an expansion mode of operation, the piston may be moved
by expanding gas within the cylinder to drive the motor/generator
acting as a generator. The expanded gas may be flowed out of the
system, or flowed to a successive lower-pressure stage for
additional expansion.
[0267] The cylinder is in selective fluid communication with a high
pressure side or a low pressure side through valving 812. In this
particular embodiment, the valving is depicted as a single
multi-way valve. However, the present invention is not limited to
such a configuration, and alternatives are possible.
[0268] For example, in lieu of a single, multi-way valve, some
embodiments of the present invention may include the arrangement of
multiple one-way, two-way, or three-way valves in series. Examples
of valve types which could be suitable for use in accordance with
embodiments of the present invention include, but are not limited
to, spool valves, gate valves, cylindrical valves, needle valves,
pilot valves, rotary valves, poppet valves (including cam operated
poppet valves), hydraulically actuated valves, pneumatically
actuated valves, and electrically actuated valves (including
voice-coil actuated valves).
[0269] When operating in the compression mode, gas from the low
pressure side is first flowed into the cylinder, where it is
compressed by action of the piston. The compressed gas is then
flowed out of the cylinder to the high pressure side.
[0270] When operating in the expansion mode, gas from the high
pressure side is flowed into the cylinder, where its expansion
drives the piston. The expanded gas is subsequently exhausted from
the cylinder to the low pressure side.
[0271] While FIG. 8 shows an apparatus employing a moveable member
configured to reciprocate within a chamber, this is not required.
Alternative embodiments could employ a moveable member that is
configured to undergo a different type of motion. Thus, some
embodiments could employ a moveable member configured to rotate
within a chamber. Examples include but are not limited to screws,
and also rotors featuring blades, lobes, scrolls, or vanes. Some
examples include but are not limited to turbines, gerotors,
quasi-turbines, and roots-type devices.
[0272] While the particular embodiment of FIG. 8 shows power being
communicated to and from the chamber through a mechanical linkage
comprising a piston rod and a crankshaft, a wide variety of
mechanical linkages are possible. Examples include but are not
limited to multi-node gearing systems such as planetary gear
systems. Examples of mechanical linkages include shafts such as
crankshafts, chains, belts, driver-follower linkages, pivot
linkages, Peaucellier-Lipkin linkages, Sarrus linkages, Scott
Russel linkages, Chebyshev linkages, Hoekins linkages, swashplate
or wobble plate linkages, bent axis linkages, Watts linkages, track
follower linkages, and cam linkages. Cam linkages may employ cams
of different shapes, including but not limited to sinusoidal and
other shapes. Various types of mechanical linkages are described in
Jones in "Ingenious Mechanisms for Designers and Inventors, Vols. I
and II", The Industrial Press (New York 1935), which is hereby
incorporated by reference in its entirety herein for all
purposes.
[0273] And while the particular embodiment of FIG. 8 features the
transmittal of power from the chamber by a mechanical linkage, this
is also not necessarily required. Alternative embodiments could
utilize other types of linkages, including but not limited to
hydraulic linkages, magnetic linkages, electro-magnetic linkages,
electric linkages, or pneumatic linkages.
[0274] Moreover, while FIG. 8 shows an apparatus comprising a
single moveable member capable of operating in a reversible manner
to compress gas or to be driven in response to expanding gas, this
is also not required. Alternative embodiments could employ separate
structures dedicated to the compression and expansion functions. In
certain embodiments, these dedicated compression and expansion
structures could be configured to be in selective communication
with one another through a common linkage, which may be mechanical,
hydraulic, pneumatic, magnetic, or electro-magnetic in nature. For
example, particular embodiments may employ selective communication
with a common rotating shaft through clutch mechanisms.
[0275] Embodiments of the present invention utilize heat exchange
between liquid and gas that is undergoing compression or expansion,
in order to achieve certain thermodynamic efficiencies.
Accordingly, the system further includes a liquid flow network 820
that includes pump 834 and valves 836 and 842.
[0276] The liquid flow network is configured to inject liquid into
the cylinder to perform heat exchange with expanding or compressing
gas. In this embodiment, the liquid is introduced through nozzles
822. In other embodiments, a bubbler may be used, with the gas
introduced as bubbles through the liquid.
[0277] The liquid that has been injected into the cylinder to
exchange heat with compressed gas or expanding gas, is later
recovered by gas-liquid separators 824 and 826 located on the low-
and high-pressure sides respectively. Examples of gas-liquid
separator designs include vertical type, horizontal type, and
spherical type. Examples of types of such gas-liquid separators
include, but are not limited to, cyclone separators, centrifugal
separators, gravity separators, and demister separators (utilizing
a mesh type coalescer, a vane pack, or another structure).
[0278] Liquid that has been separated may be stored in a liquid
collector section (824a and 826a respectively). A liquid collector
section of a separator may include elements such as inlet diverters
including diverter baffles, tangential baffles, centrifugal,
elbows, wave breakers, vortex breakers, defoaming plates, stilling
wells, and mist extractors.
[0279] The collected separated liquid is then thermally conditioned
for re-injection. This thermal conditioning may take place
utilizing a thermal network. Examples of components of such a
thermal network include but are not limited to liquid flow
conduits, gas flow conduits, heat pipes, insulated vessels, heat
exchangers (including counterflow heat exchangers), loop heat
pipes, thermosiphons, heat sources, and heat sinks.
[0280] For example, in an operational mode involving gas
compression, the heated liquid collected from gas-liquid separator
826 is flowed through heat exchanger 828 that is in thermal
communication with heat sink 832. The heat sink may take one of
many forms, including an artificial heat sink in the form of a
cooling tower, fan, chiller, or HVAC system, or natural heat sinks
in the form of the environment (particularly at high latitudes or
altitudes) or depth temperature gradients extant in a natural body
of water.
[0281] In an operational mode involving gas expansion, the cooled
liquid collected from gas-liquid separator 824 is flowed through
heat exchanger 852 that is in thermal communication with heat
source 830. Again, the heat source may be artificial, in the form
of heat generated by industrial processes (including combustion) or
other man-made activity (for example as generated by server farms).
Alternatively, the heat source may be natural, for example
geothermal or solar in nature (including as harnessed by thermal
solar systems).
[0282] Flows of liquids and/or gases through the system may occur
utilizing fluidic and/or pneumatic networks. Examples of elements
of fluidic networks include but are not limited to tanks or
reservoirs, liquid flow conduits, gas flow conduits, pumps, vents,
liquid flow valves, gas flow valves, switches, liquid sprayers, gas
spargers, mixers, accumulators, and separators (including
gas-liquid separators and liquid-liquid separators), and
condensers. Examples of elements of pneumatic networks include but
are not limited to pistons, accumulators, gas chambers liquid
chambers, gas conduits, liquid conduits, hydraulic motors,
hydraulic transformers, and pneumatic motors.
[0283] As shown in FIG. 8, the various components of the system are
in electronic communication with a central processor 850 that is in
communication with non-transitory computer-readable storage medium
854, for example relying upon optical, magnetic, or semiconducting
principles. The processor is configured to coordinate operation of
the system elements based upon instructions stored as code within
medium 854.
[0284] The system also includes a plurality of sensors 860
configured to detect various properties within the system,
including but not limited to pressure, temperature, volume,
humidity, and valve state. Coordinated operation of the system
elements by the processor may be based at least in part upon data
gathered from these sensors.
[0285] Liquid that is being introduced for the purpose of heat
exchange with gas undergoing compression or expansion according to
embodiments of the present invention, may exhibit particular
performance characteristics. One performance characteristic is
droplet size.
[0286] Droplet size may be measured using DV50, Sauter mean
diameter (also called SMD, D32, d32 or D[3, 2]), or other measures.
Embodiments of nozzles according to the present invention may
produce liquid droplets having SMD's within a range of between
about 10-300 microns. Examples of droplet sizes produced by
embodiments of nozzles according to the present invention may
include but are not limited to those having a SMD of about 300
microns, 250 microns, 200 microns, 150 microns, 100 microns, 90
microns, 80 microns, 70 microns, 60 microns, 50 microns, 40
microns, 30 microns, 25 microns, and 10 microns.
[0287] Another performance characteristic of liquid spray nozzles
according to embodiments of the present invention, is flow rate.
Embodiments according to the present invention may produce a flow
rate of between about 20 and 0.01 liters per second. Examples of
flow rates of embodiments of nozzles according to the present
invention are 20, 10, 5, 2, 1, 0.5, 0.25, 0.1, 0.05, 0.02, and 0.01
liters per second.
[0288] Another performance characteristic of liquid spray nozzles
according to embodiments of the present invention, is breakup
length. Liquid output by embodiments of nozzles according to the
present invention may exhibit a breakup length of between about
1-100 mm. Examples of breakup lengths of sprays of liquid from
nozzles according to the present invention include 100, 50, 25, 10,
5, 2, and 1 mm.
[0289] Embodiments of nozzles according to the present invention
may produce different types of spray patterns. Examples of spray
patterns which may be produced by nozzle embodiments according to
the present invention include but are not limited to, hollow cone,
solid cone, stream, single fan, and multiple fans.
[0290] Embodiments of nozzles according to the present invention
may produce spray cone angles of between about 20-180 degrees.
Examples of such spray cone angles include but are not limited to
20.degree., 50.degree., 90.degree., 120.degree., 140.degree.,
150.degree., 155.degree., 160.degree., 165.degree., 170.degree.,
175.degree., and 180.degree..
[0291] Droplet spatial distribution represents another performance
characteristic of liquid spray nozzles according to embodiments of
the present invention. One way to measure droplet spatial
distribution is to measure the angle of a sheet or cone
cross-section that includes most of the droplets that deviate from
the sheet. In nozzle designs according to embodiments of the
present invention, this angle may be between 0-90 degrees. Examples
of such angles possibly produced by embodiments of the present
invention include but are not limited to 0.degree., 1.degree.,
2.degree., 3.degree., 4.degree., 5.degree., 7.5.degree.,
10.degree., 15.degree., 20.degree., 25.degree., 30.degree.,
45.degree., 60.degree., 75.degree., or 90.degree..
[0292] According to certain embodiments of the present invention,
it may be important to control the amount of liquid introduced into
the chamber to effect heat exchange. The ideal amount may depends
on a number of factors, including the heat capacities of the gas
and of the liquid, and the desired change in temperature during
compression or expansion.
[0293] The amount of liquid to be introduced may also depend on the
size of droplets formed by the spray nozzle. One measure of the
amount of liquid to be introduced, is a ratio of the total surface
area of all the droplets, to the number of moles of gas in the
chamber. This ratio, in square meters per mole, could range from
about 1 to 250 or more. Examples of this ratio which may be
suitable for use in embodiments of the present invention include 1,
2, 5, 10, 15, 25, 30, 50, 100, 125, 150, 200, or 250.
[0294] Two-Phase Heat Transfer
[0295] As mentioned above, embodiments of the present invention may
utilize heat exchange between liquid and gas that is undergoing
compression or expansion, in order to achieve certain thermodynamic
efficiencies. In certain embodiments this liquid may be introduced
into the cylinder in the form of fine liquid droplets that become
entrained within the gas flow.
[0296] Embodiments in accordance with the present invention are not
limited to injection of liquids in any particular direction
relative to a direction of motion of a moveable member, or to a
direction of an inlet flow of gas. For example, the particular
embodiments of FIGS. 8, 11, and 33-36 feature liquid sprayers
positioned on end walls of a cylinder with valve structures.
[0297] In the configuration of these embodiments, owing to the
location of the sprayers, liquid may be injected into the chamber
in a direction parallel to the movement of the piston. Such an
orientation may promote interaction between the gas and the
injected liquid to form a liquid-gas mixture having the desired
properties.
[0298] In other embodiments, the direction of liquid injection need
not necessarily be substantially coincident with the direction of
inlet of gases through gas flow valves located on the end walls of
the chamber. Such an orientation may promote interaction between
the gas and the injected liquid to form a liquid-gas mixture having
the desired properties.
[0299] For example, the particular embodiment of FIG. 31 shows
sprayers positioned on opposite side walls of the chamber, with the
valve structures positioned on the end walls. Accordingly, a
direction of liquid injection may not necessarily be substantially
parallel to a direction of gas flowed into the chamber (in
compression or expansion mode). Such lack of coincidence between
the direction of liquid injection and directions of inlet gas flow,
may promote gas-liquid mixing and the formation of a liquid-gas
mixture having the desired properties.
[0300] According to still other embodiments, liquid could be
sprayed in directions not corresponding to either the direction of
gas flow or of piston reciprocation. Such lack of coincidence
between the direction of liquid injection and directions of inlet
gas flow or piston movement could further promote gas-liquid mixing
in a manner to promoting heat exchange.
[0301] While the embodiments shown in certain of figures shows
sprayers positioned on one chamber wall, this is not required.
According to alternative embodiments, sprayers could be positioned
on a plurality of walls. Such a configuration may be facilitated by
use of a liquid manifold extending around multiple sides of the
compression or expansion chamber. In some embodiments, the outlets
of the sprayers may be aligned in a uniform or non-uniform manner
relative to each other.
[0302] Certain embodiments may position valves and sprayers
proximate to one another within a within a relatively small region
at the end wall of the chamber. Such a clustering of elements
within a small space may affect design, construction, inspection,
and/or maintenance of the apparatus.
[0303] However, it is typically the orientation of the sprayers
relative to a gas inlet valve, that determines the character of the
liquid-gas mixture. In particular the liquid is injected into the
inlet gas for heat exchange during compression/expansion processes.
Because compression or expansion may be concurrent with inlet gas
flow, it may be desirable to position the sprayers in a manner
promoting rapid interaction between incoming gas and the liquid
spray.
[0304] By contrast, the orientation of the liquid sprayers relative
to the outlet valve may be less important. This is because the
outlet valve is utilized simply to exhaust the liquid-gas mixture
once an exchange of thermal energy during compression or expansion
has already taken place.
[0305] Accordingly, certain embodiments of the present invention
may introduce liquid through sprayers oriented relative to a single
valve dedicated to regulating a flow of gases into the chamber in
compression and/or expansion modes. By careful design of the
sprayers and their position relative to the inlet valve, liquid may
be introduced to the chamber to result in a liquid-gas mixture
possessing the desired characteristics (such as droplet size,
uniformity of droplet distribution, liquid volume fraction,
temperature, and pressure). And because the same valve is used to
admit gas in both the compression and expansion modes, a liquid-gas
mixture having desired properties may be produced in each case.
[0306] The conditions under which liquid is introduced, may be
different in the compression case versus the expansion case. For
example during compression, the liquid will be introduced into a
gas flow having a lower pressure. During expansion, the liquid will
be introduced into a compressed gas flow having a higher
pressure.
[0307] Accordingly, the operational parameters of certain elements
may be controlled to produce a liquid-gas mixture having the
desired properties. One example of a parameter which may be varied
is the velocity at which the liquid is introduced into the chamber.
Such a velocity parameter may be affected by variables such as the
speed of the pump, and/or the dimensions of the sprayer, and/or
characteristics of the conduit leading to the sprayer, such as
bore, length, and number/degree of turns. In certain embodiments,
the sprayer may comprise a nozzle having an orifice with dimensions
adjustable to control a velocity of the liquid. In certain
embodiments, the characteristics of the conduit leading to the
sprayer may be changed (for example by actuation of valving
changing a path of liquid flow).
[0308] In certain embodiments, a pressure of the liquid may be
changed. This may be done, for example, by altering a
characteristic of operation of the pump (for example pump speed).
In certain embodiments, liquid pressure may be changed by
manipulation of a valve to give rise to pressure accumulation that
is periodically relieved by bursts of liquid flows at high
velocities.
[0309] The size of the liquid droplet may also affect its
interaction with gas flows of different pressures. For example, a
liquid droplet of a greater size may be able to penetrate more
deeply into a compressed volume of gas. Thus in certain
embodiments, the sprayer may be designed to produce droplet size
that is different for the compression versus the expansion
case.
Example #1
[0310] A series of simulations were devised to investigate flow
dynamics and heat transfer within gas undergoing compression within
a cylinder device. In particular, the cylinder device housed a
horizontally reciprocating piston, with two gas input ports located
in pockets in the upper portion of the cylinder head, and two gas
output ports located in pockets in the lower portion of the
cylinder head. Four liquid spray nozzles were arranged outside of
the valve pockets in the cylinder head.
[0311] Simulations were carried out using MATLAB AND FLUENT. The
velocities and directions of the gas flows through the inlet valves
were calculated based upon the cylinder size, inlet valve size, and
a piston speed of 300 RPM.
[0312] The simulations utilized Lagrangian particle tracking to
estimate the cooling effect brought about by liquid particles
injected from nozzles. Heat transfer through a splash effect was
not considered. The droplets were terminated upon hitting the wall
of the cylinder device.
[0313] In particular, turbulence effects were accounted for using a
k-epsilon model. A random stochastic tracking approach was used to
account for turbulent dispersion of droplets.
[0314] Simulations were run with three different droplet sizes (200
microns, 100 microns, and 50 microns). The droplet size in a given
simulation was uniform.
[0315] FIG. 1A shows the droplet distribution (colored by droplet
velocity) during the compression part of cycle for 200 micron
droplets. FIG. 1B shows the distribution for 50 micron droplets.
The droplets were injected at a velocity of 20 m/s. The piston
position is shown at a crank angle of 135 degrees. Comparison of
FIGS. 1A-B shows that at a droplet size of 200 micron, there are
less than 1/10 of the number of suspended droplets at 50
micron.
[0316] The simulation reflects that this difference in distribution
of the different sized droplets within the cylinder, is primarily
because the smaller 50 micron droplets slow down more quickly due
to turbulent dispersion, and are thus able to become entrained in
the flow.
[0317] FIGS. 2A-B shows the resulting temperature distribution
within the cylinder, with particle sizes of 200 micron and 50
microns respectively. The particles were injected at a velocity of
100 m/s, and the crank angle is as in FIGS. 1A-B. The 50 micron
particles appear to provide better heat transfer in comparison with
200 micron-sized particles.
[0318] FIG. 3 provides the P-V curve at various particle sizes.
These were calculated utilizing FLUENT and MATLAB. The operation
becomes more isothermal at smaller particle diameter.
[0319] Also, it seems to be easier to achieve isothermal behavior
during compression rather than expansion because the particles have
had greater chance to disperse inside the cylinder before the
beginning of the compression part of the cycle.
[0320] FIG. 4 shows the spatially average temperature inside the
cylinder. This figure indicates that the 50 micron particle
provides improved cooling.
[0321] Returning to FIGS. 2A-B, hot spots were observed in the
inlet valve pocket for both the 200 micron and 50 micron particle
sizes. Accordingly, the potential cooling effect of placing extra
sprays/nozzles at the center of the valve pockets was also
simulated.
[0322] FIG. 5 shows the particle distribution when the nozzles are
present in the center of the valve pockets. The injected particle
size is 50 microns, with a velocity of 30 m/s. The piston is
running at 300 RPM, and the crank angle is 135 degrees.
[0323] Spray can be observed in the valve pockets, and is seen to
cover a considerable region of these pockets. It may be desirable
to have a smaller droplet diameter in this region to further
improve the number of entrained droplets in valve pockets.
[0324] FIG. 6 shows the temperature distribution inside the
cylinder. Significant reduction in temperature inside the valve
pocket is observed. However, regions with high temperature continue
to persist.
[0325] FIG. 7 shows P-V curve at various particle diameter when the
nozzles are present in valve pocket. Again, these were calculated
utilizing FLUENT and MATLAB. The P-V curves for 50 micron and 100
micron particles are slightly closer to the isothermal P-V
curve.
[0326] By contrast, not much impact is observed for 200 micron
particle. This is likely because there is not enough entrainment of
such droplets in the gas flow.
[0327] In conclusion, this CFD simulation example indicates that
relatively few particles of size 200 microns are entrained in the
flow. Instead, most of the 200 micron particles pass straight
through the domain without slowing down and then impact of the
walls of the compressor. Due to low entrainment of the larger
particles, there is less cooling effect.
[0328] At particle sizes of 50 and 100 micron, more droplets become
entrained in the gas flow and the average temperature is
significantly lower. The number of 100 micron droplets suspended in
the cylinder is 5 times the number of droplets of 200 microns. The
number of droplets entrained in 50 microns is roughly 1.8 times the
number of droplets at 100 microns.
[0329] Smaller droplet sizes seem to aid in the transfer of heat
from the compressed gas. Droplets distributed with a Sauter mean
diameter in the range of 100 microns or less may be valuable to
achieve effective heat transfer.
[0330] The above example represents only one possible embodiment,
to which the present invention should not be construed as being
limited. For example, while this example utilizes k-epsilon
modeling of turbulence, this is not required and other turbulence
modeling approaches could be used.
[0331] For example, k-epsilon turbulence modeling is a two-equation
form of a linear eddy viscosity model utilizing the
Reynolds-averaged Navier-stokes (RANS) technique. Near-wall
approaches are another two-equation form of this technique.
[0332] Alternative linear eddy viscosity models include but are not
limited to algebraic models (including the Cebeci-Smith model, the
Baldwin-Lomax model, the Johnson-King model, and
roughness-dependent models), one-equation models (including
Prandtl's one-equation model, the Baldwin-Barth model, and the
Spalart-Allmaras model), k-omega models (including Wilcox's k-omega
model, Wilcox's modified k-omega model, the SST k-omega model, and
Near-wall treatment).
[0333] Nonlinear eddy viscosity models that are based upon RANS
include but are not limited to, explicit nonlinear constitutive
relation (including cubic k-epsilon and Explicit Algebraic Reynolds
Stress Models--EARSM), and .nu..sup.2-f models (including
.nu..sup.2-f and .zeta.-f).
[0334] The Reynolds Stress Model (RSM) is another example of
RANS-based turbulence model. An example of non-RANS based
turbulence modeling is Large Eddy Simulation (LES) (which includes
the Smagorinsky-Lilly model, the Dynamic subgrid-scale model, the
RNG-LES model, the wall-adapting local eddy-viscosity (WALE) model,
the kinetic energy subgrid-scale model, and near-wall treatment for
LES models. Other examples of non-RANS based turbulence modeling
include but are not limited to detached eddy simulation (DES),
direct numerical simulation (DNS), and turbulence near-wall
modeling.
Alternative Embodiments
[0335] The particular system shown in FIG. 8 represents only one
embodiment of the present invention, and alternative embodiments
having other features are possible. For example, while FIG. 8 shows
an embodiment with compression and expansion occurring in the same
cylinder, with the moveable element in communication through a
linkage with a motor/generator, this is not required.
[0336] FIG. 9 shows an alternative embodiment utilizing two
cylinders, which in certain modes of operation may be separately
dedicated for compression and expansion. Embodiments employing such
separate cylinders for expansion and compression may, or may not,
employ utilize a common linkage (here a mechanical linkage in the
form of a crankshaft) with a motor, generator, or
motor/generator.
[0337] For example, FIG. 9A is a table showing four different basic
configurations of the apparatus of FIG. 9. The table of FIG. 9A
further indicates the interaction between system elements and
various thermal nodes 14625, 14528, 14530, 14532, 14534, 14536, and
14540, in the different configurations. Such thermal nodes can
comprise one or more external heat sources, or one or more external
heat sinks, as indicated more fully in that table. Examples of such
possible such external heat sources include but are not limited to,
thermal solar configurations, geothermal phenomena, and proximate
heat-emitting industrial processes. Examples of such possible such
external heat sinks include but are not limited to, the environment
(particularly at high altitudes and/or latitudes), and geothermal
phenomena (such as snow or water depth thermal gradients).
[0338] FIGS. 9BA-9BD are simplified views showing the various basic
operational modes listed in FIG. 9A. The four different basic modes
of operation shown in FIG. 9A may be intermittently switched,
and/or combined to achieve desired results. FIGS. 9BE-BF show
operational modes comprising combinations of the basic operational
modes.
[0339] One possible benefit offered by the embodiment of FIG. 9 is
the ability to provide cooling or heating on demand. Specifically,
the change in temperature experienced by an expanding or compressed
gas, or an injected liquid exchanging heat with such an expanding
or compressed gas, can be used for temperature control purposes.
For example, gas or liquid cooled by expansion could be utilized in
an HVAC system. Conversely, the increase in temperature experienced
by a compressed gas, or a liquid exchanging heat with a compressed
gas, can be used for heating.
[0340] By providing separate, dedicated cylinders for gas
compression or expansion, embodiments according to FIG. 9 may
provide such temperature control on-demand, without reliance upon a
previously stored supply of compressed gas. In particular, the
embodiment of FIG. 9 allows cooling based upon immediate expansion
of gas compressed by the dedicated compressor.
[0341] While FIGS. 8-9 show embodiments involving the movement of a
solid, single-acting piston, this is not required. Alternative
embodiments could utilize other forms of moveable elements.
Examples of such moveable elements include but are not limited to
double-acting solid pistons, liquid pistons, flexible diaphragms,
screws, turbines, quasi-turbines, multi-lobe blowers, gerotors,
vane compressors, scroll compressors, and centrifugal/axial
compressors.
[0342] Moreover, embodiments may communicate with a motor,
generator, or motor/generator, through other than mechanical
linkages. Examples of alternative linkages which may be used
include but are not limited to, hydraulic/pneumatic linkages,
magnetic linkages, electric linkages, and electro-magnetic
linkages.
[0343] While the particular embodiments of FIGS. 8-9 show a solid
piston in communication with a motor generator through a mechanical
linkage in the form of a crankshaft, this is not required.
Alternative embodiments could utilize other forms of mechanical
linkages, including but not limited to gears such as multi-node
gearing systems (including planetary gear systems). Examples of
mechanical linkages which may be used include shafts such as
crankshafts, gears, chains, belts, driver-follower linkages, pivot
linkages, Peaucellier-Lipkin linkages, Sarrus linkages, Scott
Russel linkages, Chebyshev linkages, Hoekins linkages, swashplate
or wobble plate linkages, bent axis linkages, Watts linkages, track
follower linkages, and cam linkages. Cam linkages may employ cams
of different shapes, including but not limited to sinusoidal and
other shapes. Various types of mechanical linkages are described in
Jones in "Ingenious Mechanisms for Designers and Inventors. Vols. I
and II", The Industrial Press (New York 1935), which is hereby
incorporated by reference in its entirety herein for all
purposes.
[0344] 1. A method comprising:
[0345] providing a cylinder device having a moveable member
disposed therein;
[0346] allowing expanding gas within the cylinder device to
accelerate the moveable member at greater than 9.8 m/s (1 g) in an
absence of combustion;
[0347] spraying a liquid through a nozzle to create in the cylinder
device a mist comprising droplets having a mean diameter of 100 um
or less to exchange heat with the expanding gas;
[0348] flowing from the cylinder device, a mixture comprising
expanded gas and the liquid; and
[0349] separating the liquid from the mixture.
[0350] 2. A method as in claim 1 wherein the gas comprises air and
the liquid comprises water.
[0351] 3. A method as in claim 1 wherein the liquid droplet is
introduced at approximately ambient temperature.
[0352] 4. A method as in claim 1 wherein the moveable member is
solid.
[0353] 5. A method as in claim 4 wherein the moveable member
comprises a solid piston.
[0354] 6. A method as in claim 4 further comprising extracting
energy by causing the moveable member to drive a physical
linkage.
[0355] 7. A method as in claim 6 wherein the physical linkage
comprises a crankshaft.
[0356] 8. A method as in claim 6 further comprising generating
electricity from the physical linkage.
[0357] 9. A method as in claim 1 wherein the mist comprises
droplets having a mean diameter of 50 um or less.
[0358] 10. A method as in claim 1 wherein expanding gas within the
cylinder device accelerates the moveable member at greater than 10
g.
[0359] 11. A method as in claim 1 further comprising flowing the
gas to the cylinder device from a compressed gas storage unit.
[0360] 12. A method as in claim 11 further comprising flowing the
gas to the cylinder device from a higher-pressure expansion
stage.
[0361] 13. An apparatus comprising:
[0362] a cylinder device having a moveable member in communication
with a mechanical linkage;
[0363] a valve in electronic communication with a control system to
selectively admit compressed gas through a port of the cylinder
device;
[0364] a nozzle configured to create in the cylinder device a mist
comprising droplets having a mean diameter of 100 um or less to
exchange heat with the compressed gas expanding within the cylinder
in an absence of combustion; and
[0365] an electrical generator in communication with the mechanical
linkage.
[0366] 14. An apparatus as in claim 13 further comprising a
gas-liquid separator configured to receive a gas-liquid mixture
from the cylinder device.
[0367] 15. An apparatus as in claim 13 wherein the gas-liquid
separator is configured to receive the gas-liquid mixture through a
second port of the cylinder device.
[0368] 16. An apparatus as in claim 13 wherein the cylinder device
is in selective fluid communication with a compressed gas storage
unit through the valve.
[0369] 17. An apparatus as in claim 13 wherein the cylinder device
is in selective fluid communication with a higher-pressure
expansion stage through the valve.
[0370] 18. An apparatus as in claim 13 wherein the mechanical
linkage comprises a crankshaft.
[0371] 19. An apparatus as in claim 13 wherein the mechanical
linkage is in selective physical communication with a motor to
cause the moveable member to compress gas within the cylinder
device.
[0372] 20. An apparatus as in claim 19 wherein the motor and the
electrical generator comprise a motor/generator.
[0373] 21. An apparatus as in claim 13 further comprising a second
moveable member in physical communication with the mechanical
linkage to compress gas within a second cylinder device.
* * * * *