U.S. patent application number 11/489304 was filed with the patent office on 2008-01-24 for active gas regenerative liquefier system and method.
Invention is credited to John Arthur Barclay, Michael Arthur Barclay, Agnes Jakobsen, Miroslaw Piotr Skrzypkowski.
Application Number | 20080016907 11/489304 |
Document ID | / |
Family ID | 38970149 |
Filed Date | 2008-01-24 |
United States Patent
Application |
20080016907 |
Kind Code |
A1 |
Barclay; John Arthur ; et
al. |
January 24, 2008 |
Active gas regenerative liquefier system and method
Abstract
The present invention provides an active gas regenerative
liquefier (AGRL) for efficiently cooling and liquefying a process
stream based on the combination of several active gas regenerative
refrigerator (AGRR) stages configured to sequentially cool and
liquefy the process stream, e.g. natural gas or hydrogen. In
specific embodiments, the individual AGRR stages include heat
exchangers, dual active regenerators, and a compressor/expander
assembly, configured to recover a portion of the work of
compression of a refrigerant by simultaneously expanding a
refrigerant in one portion of the device while compressing the
refrigerant in another portion to effect cooling of a heat transfer
fluid, and ultimately the process stream.
Inventors: |
Barclay; John Arthur;
(Seattle, WA) ; Barclay; Michael Arthur;
(Arborfield Cross, GB) ; Skrzypkowski; Miroslaw
Piotr; (Seattle, WA) ; Jakobsen; Agnes;
(Sammamish, WA) |
Correspondence
Address: |
Kevin Steinacker;DICKSON STEINACKER LLP
Ste. 1401, 1201 Pacific Ave.
Tacoma
WA
98402
US
|
Family ID: |
38970149 |
Appl. No.: |
11/489304 |
Filed: |
July 18, 2006 |
Current U.S.
Class: |
62/612 ;
62/6 |
Current CPC
Class: |
F25J 1/0022 20130101;
F25J 1/0227 20130101; F25J 2270/908 20130101; F25J 1/001 20130101;
F25B 9/14 20130101; F25J 1/0015 20130101; F25J 1/0017 20130101 |
Class at
Publication: |
62/612 ;
62/6 |
International
Class: |
F25B 9/00 20060101
F25B009/00; F25J 1/00 20060101 F25J001/00 |
Goverment Interests
FEDERALLY SPONSORED RESEARCH
[0002] The invention was created during a Phase I Small Business
Innovation and Research award from NASA to CryoFuel Systems, Inc.
under contract number NNJ04JC25C completed Jul. 15, 2004, under
which the Government may have certain rights in this invention.
Claims
1. An active gas regenerative liquefier (AGRL), comprising at least
a first active gas regenerative refrigerator (AGRR) stage and a
second AGRR stage, the first AGRR stage configured to receive and
cool a process stream, and deliver the process stream to the second
AGRR stage, wherein the first and second AGRR stages have means for
heat rejection to a common heat sink.
2. The AGRL of claim 1, in which the second AGRR stage and any
subsequent AGRR stage is configured to successively receive and
cool the process stream.
3. The AGRL of claim 2, further comprising at least a third AGRR
stage for the liquefaction of natural gas.
4. The AGRL of claim 2, further comprising at least a third AGRR
stage, wherein the first through third AGRR stages have respective
cold reservoirs at temperatures of about 220 K, 164 K, and 123 K,
respectively.
5. The AGRL of claim 2, further comprising at least a third, a
fourth, a fifth, and a sixth AGRR stage for the liquefaction of
hydrogen.
6. The AGRL of claim 2, further comprising at least a third, a
fourth, a fifth, and a sixth AGRR stage, wherein the first through
sixth AGRR stages have respective cold reservoirs at temperatures
of about 192 K, 120 K, 76 K, 48 K, 32 K, and 20 K,
respectively.
7. The AGRL of claim 2 in which the process stream is hydrogen gas,
further comprising at least one ortho to para converter located
between any two sequential AGRR stages.
8. An active gas regenerative liquefier (AGRL), comprising: at
least one active gas regenerative refrigerator (AGRR) stage, said
AGRR stage comprising at least two active regenerators, means for
heat rejection to a common heat sink near room temperature, and
means to provide work to compress a refrigerant and simultaneously
recover work from expansion of the refrigerant; said AGRR stage
configured to receive and cool a process stream to the point of
liquefaction, and deliver the liquefied process stream to a storage
vessel.
9. The AGRL of claim 2, in which the AGRR stage comprises: a first
active regenerator and a second active regenerator, the first
active regenerator and the second active regenerator each
comprising an array of tubes, wherein at least one passive
micro-regenerator is located at an entrance of each of said tubes,
said tubes containing a refrigerant; a compressor/expander
assembly, said assembly comprising a manifold having a first
portion and a second portion, a first piston in said first portion
of said manifold, and a second piston in said second portion of
said manifold, said array of tubes of the first active regenerator
being connected to said first portion of said manifold and said
array of tubes of the second active regenerator being connected to
said second portion of said manifold, said manifold also containing
the refrigerant, said pistons being configured to separately
periodically compress and expand the refrigerant to thereby
increase or decrease the temperature of the refrigerant in said
tubes; means to drive the pistons such that one piston compresses
the refrigerant in one portion of said manifold while the other
piston expands the refrigerant in the other portion of said
manifold, thereby enabling work recovery from expansion to offset
work required for compression; means to circulate a heat transfer
fluid between the process stream and the heat sink, said heat
transfer fluid circulating past the first active regenerator and
the second active regenerator in order to accept heat from or
transfer heat to said active regenerators; at least a first process
stream heat exchanger for exchanging heat from said process stream
to said heat transfer fluid; and at least a first heat rejection
exchanger for exchanging heat from said heat transfer fluid to said
heat sink.
10. The AGRL of claim 9, in which the means to circulate the heat
transfer fluid is coupled in phase with the means to drive the
pistons, such that the heat transfer fluid circulates along a first
flow path while the refrigerant in the first portion of the
manifold is expanded and along a second flow path while the
refrigerant in the second portion of the manifold is expanded.
11. The AGRL of claim 10, further comprising: a second process
stream heat exchanger, a second heat rejection exchanger, a first
valve connecting the first process stream heat exchanger, the
second process stream heat exchanger, and the process stream, and a
second valve connecting the first heat rejection exchanger, the
second heat rejection exchanger, and the heat sink; where said
means to circulate the heat transfer fluid comprises a circulator
that directs the heat transfer fluid in an oscillatory manner along
the first flow path in which the heat transfer fluid flows
sequentially through the first active regenerator, the first
process stream heat exchanger, the second process stream heat
exchanger, the second active regenerator, and the second heat
rejection exchanger and along the second flow path sequentially
through the second active regenerator, the second process stream
heat exchanger, the first process stream heat exchanger, the first
active regenerator, and the first heat rejection exchanger; said
first valve configured to direct the process stream through the
first process stream heat exchanger when the heat transfer fluid
flows along the first flow path and to direct the process stream
through the second process stream heat exchanger when the heat
transfer fluid flows along the second flow path; said second valve
configured to direct a cooling fluid through the second heat
rejection exchanger when the heat transfer fluid flows along the
first flow path and to direct the cooling fluid through the first
heat rejection exchanger when the heat transfer fluid flows along
the second flow path.
12. The AGRL of claim 10, in which the means to circulate the heat
transfer fluid comprises a plurality of valves configured to direct
the flow of the heat transfer fluid alternately along the first
flow path in which the heat transfer fluid passes sequentially
through the first active regenerator, the process stream heat
exchanger, the second active regenerator, and the heat rejection
exchanger, and the second flow path in which the heat transfer
fluid passes sequentially through the second active regenerator,
the process stream heat exchanger, the first active regenerator,
and the heat rejection exchanger.
13. The AGRL of claim 12, in which a first valve is positioned
between the first active regenerator and the heat rejection
exchanger, a second valve is positioned between the first active
regenerator and the process stream heat exchanger, a third valve is
positioned between the second active regenerator and the heat
rejection exchanger, and a fourth valve is positioned between the
second active regenerator and the process stream heat
exchanger.
14. The AGRL of claim 12, in which the valves maintain continuous
counterflow of the heat transfer fluid in the process stream heat
exchanger and the heat rejection exchanger with periodic heat
transfer fluid flow from the first and the second active
regenerators.
15. The AGRL of claim 9, wherein the tubes of each active
regenerator are arranged in a plurality of layers with temperatures
in the tubes spanning from a cold temperature in a bottom layer to
a hot temperature in a top layer, such that the first piston
performs distributed work as a function of the temperature in the
tubes during compression of the refrigerant and the second piston
simultaneously recovers distributed work as a function of the
temperature in the tubes during expansion of the refrigerant.
16. The AGRL of claim 9, wherein the passive micro-regenerator
comprises spheres with a diameter less than the diameter of the
entrance of the tube and with a thermal mass several times greater
than the thermal mass of the refrigerant.
17. The AGRL of claim 9, wherein the passive micro-regenerator
comprises screens with a diameter approximately equal to the
diameter of the entrance of the tube and with a thermal mass
several times greater than the thermal mass of the refrigerant.
18. The AGRL of claim 9, in which the temperature differences
between the heat transfer fluid and the refrigerant within the
active regenerators are 2 K or less.
19. The AGRL of claim 9, in which the temperature differences
between the heat transfer fluid and a cooling fluid in the heat
rejection exchanger, and between the heat transfer fluid and the
process stream in the process stream heat exchanger are 10's of K
or less.
20. A natural gas liquefier, comprising at least three active gas
regenerative refrigerator (AGRR) stages situated and configured to
receive and sequentially cool a natural gas process stream.
21. A hydrogen liquefier, comprising at least six active gas
regenerative refrigerator (AGRR) stages situated and configured to
receive and sequentially cool a hydrogen process stream and an
ortho to para converter situated between each of the AGRR
stages.
22. A method of liquefying a process stream of gas, comprising the
steps of: A. Cooling a heat transfer fluid by passing it through a
first active regenerator in which a refrigerant has been cooled by
expanding said refrigerant; B. Circulating said heat transfer fluid
through a process stream heat exchanger for exchanging heat from
the process stream to the heat transfer fluid; C. Passing said heat
transfer fluid through a second active regenerator in which a
refrigerant has been heated by compression of said refrigerant; D.
Circulating said heat transfer fluid through a heat rejection
exchanger for exchanging heat from said heat transfer fluid to a
heat sink; E. Repeating steps A-D multiple times in a first stage
of refrigeration to cool said process stream to a first
temperature; and F. Repeating steps A-D multiple times in at least
one subsequent stage of refrigeration to further cool said process
stream to successively lower temperatures until liquefaction
occurs.
23. The method of claim 22, in which expansion of the refrigerant
in step A is simultaneous with compression of the refrigerant in
step C, such that a distributed work of compression of the
refrigerant is offset by a distributed work of expansion of the
refrigerant.
24. A method of making a highly efficient liquefier, comprising the
steps of A. providing at least one stage of refrigeration, said
stage comprising a first active regenerator, a second active
regenerator, and a refrigerant; B. providing means to input work
via compression of the refrigerant, said input work distributed
over a first temperature span in the first active regenerator of
each stage; C. providing means to recover work via expansion of the
refrigerant, said recovered work distributed over a second
temperature span in the second active regenerator of each stage,
said second temperature span being lower than said first
temperature span; D. providing means to couple said means for work
input and said means for work recovery enabling the work input to
be offset by the work recovered; E. maintaining small temperature
differences wherever heat transfer occurs; and F. simultaneously
optimizing heat transfer, pressure drops, and longitudinal
conduction, friction losses, and other parasitic heat leaks in the
liquefier operation.
25. The method of claim 24, in which the first and the second
temperature spans range from near cryogenic temperatures to near
room temperature and in which the means to input and recover work
are accomplished by polytropic temperature changes of the
refrigerant from distributed compression or expansion.
26. The method of claim 25, wherein the distributed compression or
expansion of the refrigerant causes temperature changes of between
15 K and 20 K.
27. The method of claim 24, wherein the means to couple means for
work input and said means for work recovery comprises a resonant
piston compressor/expander in a common cylinder.
Description
PRIORITY
[0001] The applicant claims priority from a Provisional Patent
Application filed on Jul. 15, 2005, under Application No.
60/699,948.
FIELD OF THE INVENTION
[0003] The present invention generally relates to the apparatus and
method for liquefying natural gas, hydrogen, or other cryogenic
fluids using one or more active gas regenerative refrigerators.
BACKGROUND OF THE INVENTION
[0004] Cryogenic liquids such as nitrogen, helium and oxygen are
common forms of important industrial commodities. Similarly, liquid
natural gas and liquid hydrogen provide storage, transport, and
distribution for energy systems. The capital equipment and power
required to make such cryogens are key factors in their use.
[0005] Liquefaction requires first cooling the gas from near room
temperature to its characteristic boiling temperature. At this
temperature, further cooling condenses the gas into liquid.
Cryogenic liquefaction of gases can be accomplished through a
variety of methods developed since about 1900.
[0006] Two general liquefaction techniques have evolved; those with
a combined process and refrigerant stream, and those whose process
and refrigerant streams are separate. The process stream is the gas
to be liquefied and the refrigerant stream is the substance
providing the cooling. In the former case, the Claude, Linde, or
Brayton cycles commonly liquefy gases such as methane (the
predominant component of natural gas), hydrogen, or nitrogen by
processes where the process gas is simultaneously used as the
refrigerant fluid. The cascade, mixed refrigerant, magnetic, or
Stirling cycles are good examples of existing cycles of the latter
case, where methane, hydrogen, or other cryogens are liquefied with
separate process and refrigerant streams. The patent literature
describes many devices covering systems of both types. A pre-cooled
Claude cycle liquefier of the first type is the most common
commercial scale hydrogen liquefier. A mixed refrigerant cycle
liquefier of the second type is the most common commercial scale
natural gas liquefier.
[0007] A refrigerator is a device that transfers or pumps heat from
a specific colder temperature to a specific hotter temperature. A
well-defined amount of work is required to pump a given amount of
heat. An effective liquefier with separate process and refrigerant
streams requires several refrigerators or stages combined
appropriately to liquefy completely the cryogen. In such liquefiers
each refrigerator stage requires work input to pump heat from a
colder temperature to a hotter temperature.
[0008] Each refrigerator stage has a thermodynamic coefficient of
performance (COP) called the Carnot ideal COP. The COP is the ratio
of the heat extracted from a specific colder temperature to the
work required to pump that heat to a specific hotter temperature.
Real refrigerators require more work than the ideal minimum work
due to a variety of well known mechanisms such as friction,
pressure drop, and finite heat transfer. The ratio of ideal COP to
real COP is called the efficiency of the refrigerator. Thus, in a
liquefier comprised of several refrigerator stages, the
efficiencies of the stages are combined to determine the real work
required for liquefaction. The ratio of the ideal minimum work of
liquefaction to the real work of liquefaction is defined as the
Figure of Merit or FOM of the liquefier.
[0009] The ideal rate of work input for a liquefier of a certain
rate of cryogen production depends on the gas to be cooled and
liquefied. For example, starting with a pure process stream at one
atmosphere and near room temperature, it ideally takes about 1090
kJ/kg to liquefy natural gas and about 12,100 kJ/kg to liquefy
hydrogen. As stated above, the ratio of the ideal minimum
liquefaction work to the actual liquefaction work is defined as the
FOM. Existing conventional commercial-scale liquefier technology
for natural gas and hydrogen is limited to a FOM of about
0.35.sup.1, i.e., it requires about 3 times more work than the
ideal to make liquid natural gas (LNG) or liquid hydrogen
(LH.sub.2). The technical literature.sup.2 shows clearly that this
FOM has remained the limit over the past three decades or more of
technology development. .sup.1 Block, D. L., Dutta, S. and
T-Raissi, A., "Hydrogen for Power Applications, Task 2: Storage of
Hydrogen in Solid, Liquid and Gaseous Forms," Contract Report
FSEC-CR-204-88, Florida Solar Energy Center, Cape Canaveral, Fla.
(1988). Reference to several liquefier efficiency papers.sup.2 M.
T. Syed, et al. Intl. Journal of Hydrogen Energy, Vol. 23, p. 565,
1998
[0010] It has been recently suggested that the FOM in conventional
liquefier cycles such as the Claude cycle could approach 60%.
However, this increase from about 35% to about 60% requires very
high performance components in conventional liquefier designs.
These choices translate into very expensive components that
increase the capital cost of the liquefiers. For example, a
relatively inexpensive turbine expander used in small commercial
LNG liquefiers has an isentropic efficiency of about 82%. (Such an
expander is used in a turbo-Brayton/Claude cycle in Prometheus
Energy's stranded-gas-well to LNG commercial liquefiers making
.about.5,000 gallons of LNG/day.) It is possible to increase that
expander efficiency to as high as 92% but only at about double the
cost. The same is true with more effective heat exchangers required
to increase FOMs in conventional liquefiers. More efficient gas
compressors for advanced liquefiers require more stages of
compression with intercoolers that increases prices sharply. These
disadvantages reinforce the need for a breakthrough in liquefaction
technology to increase liquefier FOM from .about.0.35 to
.about.0.60 or higher while simultaneously reducing the capital
cost of the liquefier.
[0011] Several techniques are important in designing more
efficient, cost effective liquefiers. These include: means for
efficient work input and work recovery; using heat exchange that
matches the heat removal from the process stream to minimize heat
transfer across large temperature approaches; elimination of large
temperature differences between refrigerants and the process
stream; reduction of parasitic heat leaks across large temperature
differences; and avoidance of irreversible process changes such as
excess pressure drops in pipes and valves. Each of these factors
affects the FOM of the device. Active regenerative refrigerator
technology has been developed over the past three decades with the
objective of providing such features with substantially less
irreversible entropy production and therefore higher
efficiency.
[0012] An active regenerative refrigerator separates the process
stream from the refrigerant, i.e. a gas or solid, and the heat
transfer fluid. A passive regenerator with periodic heat transfer
can be thought of as a thermal flywheel, i.e., storing thermal
energy in one stage of the cycle and returning thermal energy in
another stage of the cycle. Regenerative heat exchange between the
refrigerant and the heat transfer fluid is periodic rather than
steady state heat exchange. It is well known that high performance
regenerative heat exchangers offer compact highly efficient designs
for a regenerative thermodynamic cycle such as disclosed herein.
The refrigerants provide the means to introduce and recover work in
the cycles, usually with associated temperature changes that cause
heat to flow. The refrigerants also provide the cooling for the
thermodynamic cycles. In some regenerative cycles, such as a
magnetic cycle and the gas cycle disclosed in the present
invention, the regenerator function necessary for the cycle is
simultaneously provided by the refrigerant. In that sense the
refrigerant is "an active regenerator". In known prior art, active
regenerators utilize excellent heat transfer geometries that have
large thermal masses and huge heat transfer surfaces simultaneously
with low pressure drop and low thermal conduction across inherent
thermal gradients required for efficient regenerative cycle
devices.
[0013] The active magnetic regenerative liquefier (AMRL) technology
has been progressing since the mid seventies..sup.3 As described in
U.S. Pat. No. 4,704,871 to Barclay, et al., active magnetic
regenerative refrigerators (AMRRs) employ paramagnetic or
ferromagnetic materials that, when adiabatically passed into or out
of a magnetic field (typically created by a superconducting
magnet), increase or decrease in temperature due to a phenomenon
known as the magnetocaloric effect. The basic regenerative magnetic
cycle consists of: adiabatic temperature increase upon
magnetization; heat transfer to a thermal sink; regenerative heat
transfer to decrease the magnetized magnetic refrigerant average
temperature; adiabatic temperature decrease upon demagnetization;
heat transfer from the thermal load; and regenerative heat transfer
to increase the demagnetized magnetic refrigerant average
temperature back to the starting temperature. .sup.3 See for
example, U.S. Pat. No. 4,332,135 (1982) by W. A. Steyert and J. A.
Barclay, NASA KSC reports, and other more recent patents and
papers.
[0014] Several AMRR stages can be configured as liquefiers to make
an active magnetic regenerative liquefier or AMRL for natural gas
or hydrogen. The AMRL disclosed in U.S. Pat. No. 6,467,274 to
Barclay, et al. achieves high FOMs by using multi-stage AMRRs in a
parallel or a series configuration. Decades of component
development and numerous analyzed designs establish a good basis to
make efficient AMRLs spanning from about 300 K to about 110 K for
LNG or to about 20 K for LH.sub.2. However, there is a substantial
cost in utilizing the AMRL. Combinations of multiple magnetic
materials must be incorporated into highly effective magnetic
regenerators. These regenerators must have a heat transfer fluid
passing periodically through them to couple the solid refrigerant
to the process stream. Superconducting magnets cooled to near
liquid helium temperatures (4 K) are needed to provide magnetic
fields of sufficient strength (5-6 Tesla) to cause temperature
changes in the magnetic refrigerants essential to achieving high
FOMs. The AMRL technology has excellent promise for large-scale
hydrogen or natural gas liquefaction, but has the inherent
disadvantage of being a complex system to design and operate.
[0015] Other novel active regenerative refrigerators using
electrocaloric and elastocaloric effects have also been proposed
and/or demonstrated, as in U.S. Pat. No. 5,465,781, issued to
DeGregoria. None of these have been proposed or used for
liquefaction of natural gas, hydrogen, or other cryogens because of
the inability to be effectively implemented at cryogenic
temperatures.
[0016] U.S. Pat. No. 6,332,323 to Reid, et al., discloses numerous
potential embodiments for an active gas regenerative refrigerator
(AGRR). None of the embodiments of the AGRRs in this patent were
configured as a liquefier for natural gas, hydrogen, or other
cryogens. In addition, none of the AGRRs in this patent allowed for
distributed work input and recovery.
[0017] Thus, there is a need for a liquefier for natural gas,
hydrogen, and other cryogens that operates with a high Figure of
Merit and yet is relatively simple and cost-effective to produce
and operate. There is also a need for more efficient active gas
regenerative refrigerators for separate applications such as
economical cooling high temperature superconductor devices. The
present invention fulfills this and other needs.
SUMMARY OF THE INVENTION
[0018] The present invention claimed in this patent application is
an active gas regenerative liquefier (AGRL) based on the
combination of several active gas regenerative refrigerators
(AGRRs). The AGRR stages are configured to sequentially receive and
cool a process stream and deliver the process stream to the next
colder AGRR stage until the process stream is liquefied.
[0019] Each AGRR stage used in the AGRL may include dual identical
active regenerators also containing a refrigerant, a heat transfer
fluid and a heat transfer fluid circulator that connects the dual
active regenerators to a process stream heat exchanger and a heat
sink exchanger, and a refrigerant compressor/expander assembly as
disclosed and claimed. In one embodiment, each AGRR stage has two
process stream heat exchangers and two heat rejection exchangers,
using one set of exchangers in one half of the cycle and the other
set of exchangers in the other half. In an alternate embodiment,
each AGRR stage incorporates multiple three-way valves to
efficiently control the flow of the heat transfer fluid through a
single process stream heat exchanger and a single heat sink
exchanger during the cycle. Preferably, each AGRR stage will also
have instrumentation for operational control and a cold box for
effective thermal insulation of the cryogenic components from the
environment.
[0020] An efficient AGRL for natural gas is disclosed using at
least three AGRR stages. An AGRL for hydrogen is also disclosed
using at least six AGRR stages. Use of several stages of
refrigeration approximates continuous heat removal from the process
stream at the highest possible temperature, one of the elements of
a highly efficient liquefier design. Using such configured AGRR
stages, the ideal FOM can reach about 0.92 for the three-stage
natural gas liquefier of this invention and about 0.81 for the
six-stage hydrogen liquefier of this invention. Fewer or more AGRR
stages can also be used in other embodiments of an AGRL.
[0021] Work recovery is another feature of a highly efficient
liquefier. Gas refrigerant compression is very work intensive and
is normally done near ambient temperature. Gas refrigerant
expansion is normally done in a device called an expander at
cryogenic temperatures. Besides providing cooling of the
refrigerant, it can be a means to recover a portion of work of
compression.sup.4. The imbalance of work input in the compressor
and recovery in an expander in a conventional gas liquefier is a
fundamental limitation to its efficiency. The present invention
uniquely provides distributed compression in all AGRR stages at
temperatures from near room temperature to near cryogen
liquefaction temperatures and simultaneously recovers work from
distributed expansion at temperatures from near room temperature to
cryogenic temperatures. This feature makes the AGRL design
inherently more efficient than conventional liquefiers. .sup.4
Isentropic expansion of a gaseous refrigerant from high pressure to
low pressure cools the refrigerant and provides work output. This
work can be used to offset the work of compression of the
refrigerant.
[0022] In a parallel-type AGRL each AGRR stage uses a refrigerant
to pump a thermal load from a cold temperature unique to each AGRR
stage to a common heat rejection temperature, e.g., near room
temperature. The final, or coldest, AGRR stage removes primarily
latent heat from a process stream to liquefy it and expels rejected
heat at near room temperature. The previous successively warmer
AGRR stages in a parallel-type AGRL remove primarily sensible heat
from the process stream to cool it and expel rejected heat at near
room temperature. The efficiency of each AGRR stage depends on its
inherent inefficiencies from real heat transfer, fluid flow,
refrigerant compression/expansion, and other processes necessary to
pump heat from a colder to a warmer temperature. Since each AGRR
stage spans a different temperature range in the disclosed AGRL,
each stage can be optimally designed to achieve high efficiency by
choices that minimize irreversible entropy creation in all aspects
of the overall AGRL design. These include minimum temperature
approaches in all heat exchangers, small pressure drops, and
efficient work input and recovery. By using a parallel-type AGRL
configuration with highly efficient AGRR stages as disclosed a FOM
of about 0.60 for liquefaction of hydrogen is achievable at
relatively low cost. The combination of the several real AGRR stage
efficiencies can provide natural gas or hydrogen liquefiers with
FOMs of between 0.52 and 0.69. This performance is a quantum
increase over that of the best conventional liquefiers with FOMs of
about 0.35. This AGRL invention provides a breakthrough in
efficient and cost-effective hydrogen and natural gas
liquefaction.
[0023] In another embodiment of the invention, the AGRL includes
several stages of refrigeration, with each stage including an array
of discrete micro compressor-expander units (MCEU).sup.5 configured
as a high performance active regenerator having excellent heat
transfer, low pressure drop, and low longitudinal conduction with
respect to the heat transfer fluid and regenerator materials. The
compressor-expander units are configured such that the compression
of the refrigerant within a unit is coupled to the simultaneous
expansion of the refrigerant within the other end of the unit,
thereby allowing distributed work input and recovery from near
ambient temperature to cryogenic temperatures necessary for
liquefaction of natural gas or hydrogen. In this embodiment, the
net work input is reduced substantially thus providing very
efficient regenerative refrigeration, no matter what the
temperature span of the liquefier. This input of "distributed net
work" is unique among gas liquefiers. .sup.5 The MCEU or micro
compressor expander unit may consist of a small diameter tube, such
as mm dimensions, with working refrigerant gas that is separately
compressed at one end of the MCEU, and simultaneously separately
expanded on the other end of the MCEU such that the work for
compression of the refrigerant is partially compensated by work
produced by the expansion.
[0024] Another feature of the AGRL is the use of multistage
refrigeration to a sequence of separate process stream heat
exchangers containing the flowing process stream (natural gas or
hydrogen gas) that approximates continuous cooling. In the case of
hydrogen, associated ortho-to-para (o-p) exothermic converters at
each stage enables removal of the o-p heat as the hydrogen is
cooled and liquefied. This multistage AGRR design feature markedly
increases the thermal efficiency of an AGRL compared to
conventional hydrogen liquefiers, e.g. ones based on a Claude
cycle.
[0025] In another embodiment of the disclosed invention,
temperature approaches in between refrigerants in the individual
tubes or MCEUs and heat exchange fluid in the active regenerators
are kept small, thus avoiding the inherent inefficiency of
conventional cycles when heat transfers across larger temperature
spans. Thus, the temperature differences between the array of
distributed MCEUs, between the heat transfer fluid that couples the
refrigerants in the MCEUs to the process stream heat exchangers and
heat sink exchangers of each AGRR, and between the heat exchange
fluid and the process streams can be selected in a manner that
optimizes efficient heat transfer and thereby increases
thermodynamic efficiency.
[0026] The AGRL as disclosed and claimed is relatively simple in
design as compared to other liquefiers, thus requiring fewer
components, less expense, and simpler controls for automatic
operation. These features are important attributes of commercial
liquefiers.
[0027] A method of liquefying a process stream of gas is also
disclosed. The steps include: cooling a heat transfer fluid by
passing it through a first active regenerator in which a
refrigerant has been expanded, circulating the heat transfer fluid
through a process stream heat exchanger to pick up a thermal load
from the process stream, passing the heat transfer fluid through a
second active regenerator in which the refrigerant has been
compressed, thereby heating the heat transfer fluid, and
circulating the heat transfer fluid through a heat rejection
exchanger to expel excess heat to a heat sink. These steps may be
repeated several times in each of several stages of refrigeration,
each stage successively cooling the process stream until
liquefaction occurs.
[0028] The above summary of the present invention is not intended
to represent each embodiment, or every aspect, of the present
invention. The present invention also includes any additional
features and benefits which are apparent from the detailed
description and figures set forth below.
BRIEF DESCRIPTION OF THE FIGURES
[0029] FIG. 1A depicts an AGRL of three AGRR stages for
liquefaction of natural gas.
[0030] FIG. 1B depicts an AGRL of six AGRR stages for liquefaction
of hydrogen.
[0031] FIG. 2 is a schematic of a distributed-tube AGRR with a
common compressor/expander assembly. This AGRR is configured as a
single stage of a multistage AGRL.
[0032] FIG. 3 is a pressure versus temperature diagram of the gas
cycle of the combined refrigerant and a thin-walled tube in a
selected micro compressor-expander unit of an AGRR operating in an
active regenerative cycle near room temperature. This diagram
assumes instantaneous heat transfer between the refrigerant and the
thin-walled tube in the MCEU. The pressures used in this diagram
are illustrative rather than prescriptive.
[0033] FIG. 4A is a side view of a simplified active regenerator
configured as an array of thin-walled tubes with helium gas as the
refrigerant. This active regenerator has a common supply duct from
the compressor/expander and thermally isolates each layer of
thin-walled tubes from room temperature using a passive
micro-regenerator of stainless steel particles.
[0034] FIG. 4B is side view of the active regenerator taken along
the line 4B-4B in FIG. 4A that illustrates the array of thin-walled
tubes comprising the regenerator geometry.
[0035] FIG. 4C is an expanded view of two thin-walled tubes of the
active regenerator, together with the passive micro-regenerators
that thermally isolate the refrigerant in each of the tubes from
the refrigerant supplied by a common room temperature
compressor/expander assembly.
[0036] FIGS. 5A to 5G are diagrams of the typical temperature
profiles of the active regenerator refrigerant/thin tubes of an
individual AGRR stage of the AGRL at various points of the active
gas regenerative cycle. The flow indicated in the diagram is that
of the heat transfer fluid through the regenerator in its periodic
flow that sequentially connects thermally to the process stream
thermal load and to the heat sink heat exchanger.
[0037] FIG. 6 is a schematic showing the flow of the heat transfer
fluid in the refrigeration cycle in the first stage of an AGRL
system with dual active regenerators. The three-way valve ports are
shown as black if closed and white if open. The process stream is
cooled in this AGRR stage from about 300 K to about 220 K The heat
rejection exchanger shows the heat transfer fluid being cooled from
330 K to 300 K. This embodiment could be the first stage of a
natural gas or a hydrogen AGRL.
[0038] While the invention is susceptible to various modifications
and alternative forms, specific embodiments are shown by way of
example in the drawings and are described in detail herein. It
should be understood, however, that the invention is not intended
to be limited to the particular forms disclosed. Rather, the
invention is to cover all modifications, equivalents, and
alternatives falling within the spirit and scope of the invention
as described.
DETAILED DESCRIPTION OF THE ILLUSTRATED EMBODIMENTS
[0039] Referring to FIG. 1A, a preferred embodiment of the active
gas regenerative liquefier or AGRL (10) is schematically shown. A
process stream (11) (e.g. natural gas) enters the AGRL (10) from
the left in the figure. The natural gas process stream (11) in this
embodiment is initially at a temperature of 300 K and a pressure of
about 20 psig (0.24 MPa). The AGRL (10) comprises three AGRR stages
(12). In parallel configuration, each AGRR stage (12) has a heat
rejection exchanger (13), or heat rejection means, at a common
temperature (300 K in this embodiment) and a cold heat exchanger
(14) at a temperature designed to maximize the FOM of the liquefier
system. Each AGRR stage (12) receives the process stream (11) and
cools it to about the temperature of the cold heat exchanger (14)
of that stage. In this embodiment, the first AGRR has a cold heat
exchanger at about 220 K, the second AGRR has a cold heat exchanger
at about 164 K, and the third AGRR has a cold heat exchanger at
about 123 K. The cold heat exchanger (14) of each stage removes
heat from the process stream (11) in a process stream heat
exchanger (15). The heat rejection exchanger (13) of each stage
expels heat to a cooling fluid (16) (e.g., water) in a heat sink
exchanger (17). In the final stage of refrigeration in AGRR 3 in
FIG. 1A, the latent heat of vaporization is removed, liquefying the
process stream (11). Thus, after passing through the third AGRR,
the process stream output (18) is liquid natural gas at 123 K in
this embodiment.
[0040] In one embodiment, the cooling fluid will pass through each
heat sink exchanger to a separate water-to-air heat exchanger,
where the water is cooled and returned to the liquefier system to
cycle through the heat sink exchangers again. In another
embodiment, a common heat sink exchanger may be used that is
capable of handling the heat rejected from each AGRR stage and
connected to a closed loop water chiller. This water chiller will
have a circulation pump for the water as well as a fan to drive air
convection through the water radiator. The pump and fan powers for
the chiller will be small and supplied by small motors. The process
stream heat exchangers (15) and the cryogenic AGRR cold heat
exchangers (14) may be contained within a cold box or otherwise
thermally isolated from the surroundings to reduce parasitic heat
leaks into the cryogenic liquid product (18).
[0041] FIG. 1B schematically shows a preferred embodiment of the
AGRL (10) for production of LH.sub.2. The hydrogen process stream
(11) in this embodiment is initially at a temperature of 300 K and
a pressure of about 1 atmosphere (0.1 MPa). The process stream (11)
enters from the left in the figure and passes through six
successive AGRR stages (12) with cold heat exchangers (14) at
temperatures of about 192 K, 120 K, 76 K, 48 K, 32 K and 20 K.
Between each stage, a continuous ortho-to-para catalytic converter
(19) converts the ortho form of hydrogen in the process stream (11)
to an equilibrium concentration of the para form of hydrogen at
that particular temperature, thereby increasing the efficiency of
the AGRL (10). The remaining elements are as described for FIG. 1A,
resulting in a process stream output (18) of liquid hydrogen at 20
K in this embodiment.
[0042] The designed cooling capacity of the AGRR stages scales with
the rate of production of LNG or LH.sub.2. The AGRL design can be
scaled from several hundred or thousand gallons/day upwards to much
larger liquefaction capacities. This type of AGRL for natural gas
or for hydrogen has been designed to have a FOM within the range of
about 0.52 to 0.69.
[0043] Referring to FIG. 2, a unique and novel AGRR stage (12) is
depicted. This AGRR design can be used for the several AGRR stages
required to make the AGRLs shown in FIGS. 1A and 1B. The AGRR stage
(12) includes dual regenerators (20) comprised of rectangular
arrays of many small tubes (21) and a refrigerant
compressor/expander assembly (24) operating at room temperature,
comprised of a manifold (26), pistons (28), and a drive mechanism
(30). The tubes (21) are filled with a refrigerant (23) (e.g.
helium gas) and a heat transfer gas displacer or circulator (32) is
also filled with a heat transfer fluid (35) (e.g., nitrogen gas for
LNG and helium gas for LH.sub.2). The tubes (21) in each active
regenerator (20) are connected to the refrigerant manifold (26) by
an array of passive micro-regenerators (22) located at the entrance
of each of the tubes (21) to the common refrigerant manifold (26).
Pistons (28) are located within the manifolds (26) to alternately
compress and expand the refrigerant (23) within the regenerators
(20). The pistons (28) are coupled such that one piston will
compress the working refrigerant (23) in a first (left most in FIG.
2) portion of the manifold (26) while another piston simultaneously
expands the refrigerant (23) in a second (right most in FIG. 2)
portion of the manifold (26). The pistons (28) are driven by the
drive mechanism (30).
[0044] After temperature gradients are established in arrays of
tubes (21) during a short startup sequence, the AGRR stage (12)
operates as follows: A piston (28) expands the refrigerant (23) in
all the small tubes (21) of one of the dual regenerators (the right
regenerator in FIG. 2) (20) and compresses the refrigerant (23) in
all the small tubes (21) of the second of the dual regenerators
(the left regenerator in FIG. 2) (20). The pressure decrease of the
refrigerant (23) in the tubes (21) of the right regenerator (20)
causes a polytropic expansion with a corresponding temperature
decrease of the refrigerant (23) in each tube (21) such that the
temperature in the active regenerator spans from near room
temperature at the top to near a cryogenic thermal load temperature
at the bottom. A displacer or circulator (32) drives the heat
transfer fluid (35) from a heat rejection exchanger (13) cooled by
a cooling fluid (16) downward through the array of tubes (21) in
the right regenerator (20), further cooling the heat transfer fluid
(35) to near a cryogenic thermal load temperature. The heat
transfer fluid (35) then flows through a cold heat exchanger (14)
that is thermally coupled to a process stream heat exchanger (15)
(the right most process stream heat exchanger in FIG. 2) where it
picks up a thermal load from the process stream (11) before passing
through a second cold heat exchanger (15) (the left most cold heat
exchanger in FIG. 2) and into the array of tubes (21) in the left
active regenerator (20). There is no process stream flow through
the left most process stream heat exchanger (15) in FIG. 2 so the
heat transfer fluid temperature does not change in the left most
cold heat exchanger (14). The refrigerant (23) in the small tubes
(21) of the left active regenerator (20) has been heated by a
corresponding polytropic compression caused by the motion of the
second piston (28) of the compressor/expander assembly (24). The
heat transfer fluid (35) flowing from the bottom to the top of the
left regenerator (20) is warmed as it picks up heat from the left
regenerator (20) and is directed to a heat rejection exchanger
(13), where the heat transfer fluid (35) is cooled when heat is
transferred to a cooling fluid (16) via a heat sink exchanger (17).
Three way valves (36) direct the flow of the process stream gas
(11) and the heat sink cooling fluid (16) to the appropriate
process stream heat exchanger (15) and heat sink exchanger (17) in
a counterflow direction. The displacer or circulator (32) then
reverses the flow of the heat transfer fluid (35) and the drive
mechanism (30) of the compressor/expander assembly (24) moves the
pistons (28) simultaneously to compress the refrigerant (23) in the
small tubes (21) of the right active regenerator (20) and expand
the refrigerant (23) in the small tubes (21) of the left active
regenerator (20). The heat transfer fluid (35) circulates in a
reverse direction through its flow path, now being cooled in the
left regenerator (20), picking up a thermal load from the left
process stream heat exchanger (15), receiving heat from the right
regenerator (20), and expelling the heat out to the heat rejection
exchanger (13) to the cooling fluid (16).
[0045] In the embodiment depicted in FIG. 2, there are two process
stream heat exchangers (15) and two heat sink exchangers (17).
Three-way valves (36) control the flow direction of the cooling
fluid (16) from the heat sink and the flow direction of the process
stream gas (11), allowing the periodic flow of the dual active
regenerators (20) to be coupled in counterflow with the process
stream heat exchangers (15) and the heat sink exchangers (17), one
set of exchangers being used in the first half of the cycle, and
the second set of exchangers being used in the second half of the
cycle. An alternate embodiment utilizes two additional three-way
valves to control the flow of the heat transfer fluid without the
need for two sets of exchangers, as will be described in detail
below.
[0046] During the cycle, the average temperature of the refrigerant
(23) in the array of tubes (21) is increased in one regenerator
(20) and decreased in the other regenerator (20) as the heat
transfer fluid (35) flows. The direction of flow of the heat
transfer fluid through the active regenerator (20) is reversed by
the displacer or circulator (32) when the temperature of the
bottommost layer of tubes (21) increases by about half the
temperature decrease of the small tubes caused by the polytropic
expansion of the refrigerant in bottommost layer of tubes in the
active regenerator (20). This flow reversal of the heat transfer
fluid is synchronized with the compression/expansion of the
refrigerant in the dual regenerators.
[0047] The work required to pump the heat from the thermal load to
the heat sink is distributed over all the tubes (21) comprising the
dual active regenerators (20) of each AGRR stage. By coupling the
pistons (28) together via a direct linkage, the net work of
refrigerant compression that must be externally supplied by the
drive mechanism (30) is the net work required for the thermodynamic
refrigeration provided by the AGRR stage. The offset of most of the
work of compression of the refrigerant by work recovery from
simultaneous expansion of the refrigerant and the distributed work
input as a function of temperature are two of the fundamental
reasons for a high efficiency in an individual AGRR stage (12) of
an AGRL. The cooling power of the AGRR stage (12) is proportional
to the heat transfer fluid flow rate and the effective temperature
changes of the refrigerant caused by compression/expansion. As
described, the heat transfer fluid (35) is cooled or heated by the
effective temperature changes fluid, helium gas at a modest
pressure, (e.g., 1-2 MPa), in the process stream heat exchanger for
each AGRR stage. The thermal cooling load at each stage is related
to the He and H.sub.2 mass flow rates via the equation
{dot over (m)}.sub.Hec.sub.He.DELTA.T.sub.cold={dot over
(Q)}.sub.load={dot over (m)}.sub.H2(h.sub.f-h.sub.i)
where c.sub.He is the heat transfer fluid heat capacity, h.sub.f
and h.sub.i are enthalpies of the H.sub.2 at the entrance and exit
of the process stream heat exchanger for the respective AGRR stage,
.DELTA.T.sub.cold is the effective temperature change of the He
heat transfer fluid as it passes in counterflow through the process
stream heat exchanger with an average H.sub.2 exit temperature of
about T.sub.cold for that particular process stream heat exchanger.
After passing through the process stream heat exchanger, the warmer
He heat transfer fluid then flows back through the other dual
active regenerator of the AGRR stage where the refrigerant within
all tubes in this array is compressed with corresponding effective
adiabatic temperature increases above the mean temperature at a
particular longitudinal position along the active regenerator. At
the beginning of this "hot blow" period, the temperature spanned by
the dual active regenerator with the compressed refrigerant in each
tube is .about.T.sub.cold to .about.T.sub.hot+.DELTA.T.sub.hot. The
heat transfer fluid picks up heat from each of the tubes in the
active regenerator as it flows and eventually leaves the hot end of
the regenerator at a temperature higher than the mean heat sink
temperature.
[0048] The periodic motion of the heat transfer fluid is
synchronous with the operation of the refrigerant-filled tubes and
only shifted in phase by the ratio of the time for the
compression/expansion step to the time for the blow periods (this
ratio is usually small, i.e., 0.05 to 0.1). To accomplish this
effectively with a single room temperature heat transfer fluid
displacer or circulator, the preferred embodiment uses a valve
arrangement to create a periodic of the refrigerant (23) in the
array of tubes (21) of the active regenerators (20) caused by the
compressor/expander assembly (24).
[0049] The frequency of periodic flow reversal in the displacer or
circulator (32) and the operation of the three-way valves (36) is
properly phased with the temperature changes in the array of tubes
(21) and can typically operate at reasonable frequencies near 1 Hz.
The dual regenerator configuration in this AGRR stage (12) allows
the heat transfer fluid loop to be hermetic and reversible
according to the motion of the displacer or circulator (32) and
pistons (28). The dual regenerators (20) are identical and operate
180.degree. out of phase with each other. In other words, the
compression of the refrigerant (23) by a piston (28) connected via
manifold (26) to the array of tubes (21) in one of the dual
regenerators (20) is synchronous with the expansion of the
refrigerant (23) by a piston (28) connected via a manifold (26) to
the other dual regenerator (20). Similarly, the heat transfer fluid
flow or cold blow (from top to bottom) in one active regenerator
(20) is synchronous with the heat transfer fluid flow or hot blow
(from bottom to top) in the other regenerator (20). The heat
transfer fluid (35) reciprocatively flows through each of the dual
regenerators (20) while flowing semi-continuously in counterflow
through the cold heat exchanger (14) coupled to the process stream
heat exchanger (15) and the heat rejection exchanger (13) coupled
to the heat sink exchanger (17).
[0050] In one embodiment, each passive micro-regenerator (22) for
each individual tube in a given layer of tubes in the
two-dimensional array of tubes (21) comprising the dual active
regenerators (20) has a thermal mass of .about.30-50 times the
thermal mass of the refrigerant (helium gas) that flows in and out
of each tube (21) during compression or expansion. A typical
passive micro-regenerator material is stainless steel spheres with
a diameter of .about.200-300 microns (0.2-0.3 mm). The temperature
span across each layer of passive micro-regenerators (22) feeding
each layer of tubes (21) in the active regenerator will be from the
average cold temperature of the refrigerant plus tube combination
in that layer of tubes (21) and the near room temperature
refrigerant in the manifold (26) connecting the compressor/expander
to the active regenerator. The diameter of each cylindrical passive
micro-regenerator (22) is the same as the diameter of the small
tube it is connected to in the dual active regenerators. The length
of these passive micro-regenerators (22) is a design variable and
was chosen as a few centimeters in one tested embodiment. The
pressure drop for the helium gas refrigerant as it flows through
the passive micro-regenerators (22) in and out of the individual
tubes (21) was designed to be very low at an operational cycle time
of .about.1 Hz. The pressure changes of the helium refrigerant (23)
within the tubes (21) can be from .about.215 psia to .about.430
psia in a typical operation. The mean pressure of the He heat
transfer fluid that flows through the dual active regenerators (20)
and the heat transfer fluid in the reversible displacer or
circulator (32) is an operating variable and was about 200 psia in
a test embodiment. It may be higher or lower if desired. The
pressure drop of the heat transfer fluid (helium gas in the
preferred embodiment for LH.sub.2) through optimally designed dual
regenerators (20) is typically very small, i.e., 10's to 100's of
Pa.
[0051] The pressure versus temperature diagram of FIG. 3
illustrates the thermodynamic cycle of a typical tube filled with
refrigerant gas to a mean pressure of about 1800 psia. This figure
shows how the temperature of the refrigerant changes throughout the
active gas regenerative cycle. The pressure increase from the
compression of the refrigerant in the tubes in the same layer
within an array of tubes comprising an active regenerator increases
the pressure from about 1800 psia to about 2600 psia. The
corresponding temperature of the refrigerant gas plus the thin tube
shell increases from about 270 K to about 285 K, i.e. to point 2 on
P-T coordinates of FIG. 3. This point corresponds to the
temperature T.sub.hot+.DELTA.T.sub.hot. The heat transfer fluid is
circulated by and around the tubes in this particular layer of
tubes to partially remove the heat of compression and cool the
refrigerant and tube to point 3 (at temperature T.sub.hot) on the
P-T coordinates in FIG. 3. The expansion of the refrigerant gas by
the compressor/expander assembly from about 2500 psia to about 1000
psia cools the refrigerant to about 235 K as shown by point 4 on
the coordinates of FIG. 3, at temperature
T.sub.cold-.DELTA.T.sub.cold. The heat transfer fluid again flows
by and around the tubes in this layer of the active regenerator in
the opposite direction while heat is added to the cold refrigerant
and tube until the temperature reaches about 250 K as indicated by
point 1 (at temperature T.sub.cold) on the coordinates in FIG. 3.
The compression of the refrigerant by the compressor/expander
assembly is repeated to increase the pressure from about 1000 psia
to about 2600 psia with a corresponding temperature increase to
about 285 K as shown in point 2. The cycle is now complete and
repeats at the frequency of operation of about 1 Hz.
[0052] FIG. 4A shows a cross side view of a simplified active
regenerator (20). The layers of tubes (21) are shown attached to
the passive micro-regenerators (22) of each tube in each layer of
the active regenerator (20). These passive micro-regenerators (22)
allow the common compressed and expanded refrigerant to always be
near room temperature in the manifolds (26) by highly effective
heat transfer to or from the refrigerant as it flows in and out of
individual thin-walled tubes (21) of the active regenerators (20).
The thermal mass of the passive micro-regenerators is chosen to be
much larger than the refrigerant as required for high performance
designs. From the top of the active regenerator, each layer of
passive micro-regenerators (22) will have a slightly lower
temperature on the tube end than the layer above it, thereby
creating a temperature span across the active regenerator in the
AGRR. For example, in the first stage AGRR of a three-stage AGRL
for LNG, the upper layer of micro-regenerators (22) will have its
average tube-end temperature close to 290 K and the bottommost
layer of passive micro-regenerators (22) will have an average
tube-end temperature close to 215 K. The passive micro-regenerators
(22) at the coldest layer of this AGRR stage will have a
temperature gradient along their longitudinal axis (from the
tube-end towards the manifold (26)) from about 215 K to about 300
K.
[0053] FIG. 4B is a side view of the active regenerator to
illustrate the staggered array of thin-walled tubes (21) filled
with refrigerant (23). In the preferred embodiment, these tubes are
typically 0.25'' in diameter with a wall thickness of
.about.0.001''. FIG. 4C illustrates the two helium flows
(refrigerant (23) and heat transfer fluid (35)) required for the
operation of the AGRR stage. One transverse flow in the active
regenerator is the refrigerant (23) that is compressed and expanded
by the combined compressor/expander assembly. The other vertical
flow is the heat transfer fluid (35) through the array of tubes to
periodically transfer heat to and from the refrigerant-filled tubes
(21) and coupling these heat flows to the heat rejection exchanger
and cold exchanger of each AGRR stage within the AGRL. The active
regenerator (20) is hermetically sealed by a suitably chosen
assembly container (38).
[0054] FIG. 5 illustrates the combined operation of one of the dual
active regenerators of an AGRR stage. The temperatures span from
about 300 K on the hot end to about 20 K on the coldest end of the
sixth AGRR stage of an AGRL for hydrogen or to about 123 K on the
coldest end of the third AGRR stage of an ARGL for natural gas.
[0055] FIG. 5A shows the temperature changes of the individual
tubes along the longitudinal axis of the active regenerator after
the AGRR cycle as described earlier in FIG. 3. In other words, FIG.
5A depicts the temperature span from a cold temperature to a hot
temperature of the individual refrigerant-filled tubes along the
longitudinal axis of each AGRR active regenerator. FIG. 5B shows a
side view of the layers of thin-walled tubes as described in FIG.
4. FIG. 5C shows the steady state of temperature of the array of
layered tubes along the active regenerator after a significant
period of operation. FIG. 5D shows the temperature profile of the
refrigerant-filled tubes after the compression step of the cycle,
the temperature in each refrigerant-filled tube having increased.
FIG. 5E shows the temperature profile after the hot blow period of
the active regenerator with heat transfer fluid going from left to
right in the figure. FIG. 5F shows the temperature reduction of the
refrigerant-filled tubes in the regenerator after the expansion
step. FIG. 5G shows the temperature profile after the cold blow of
the cycle where the heat transfer fluid flows from right to left in
the figure.
[0056] In the hot blow, the heat transfer gas comes out of the
right end of the active regenerator at a temperature
T.sub.hot+.phi..DELTA.T.sub.hot where .phi. ranges from 1 to 0
during the blow period (usually .phi. averages about 0.5) and
.DELTA.T.sub.hot is the temperature change of the hottest layer of
tubes. This hot gas can reject heat to the cooling fluid as the
heat transfer fluid cools back toward T.sub.hot. Similarly, during
the cold blow, the heat transfer gas comes out of the left end of
active regenerator at a temperature of
T.sub.cold-.phi..DELTA.T.sub.cold and it can absorb heat from the
thermal load from the process stream as it warms toward T.sub.cold.
The operation of this active regenerator is similar to that of high
performance regenerators in other regenerative refrigerators with
the added feature that each refrigerant-filled tube in the
regenerator has the ability to actively change its temperature and
thus enable distributed refrigeration within the regenerator rather
than just be a passive heat sink/source as is traditionally the
case. Hence, there is a large surface area for heat transfer
between the tubes and the heat transfer fluid in the preferred
embodiment. Also, the pressure drop to flow, the longitudinal
(axial) conduction, and porosity of the regenerator are as small as
possible in the preferred embodiment. However, other active
regenerator configurations may be used in other embodiments. The
temperature increases of the refrigerant-filled tubes along the
regenerator longitudinal axis must satisfy the second law of
thermodynamics and be in the ideal ratio of absolute temperatures
along the regenerator as has been schematically illustrated in FIG.
5; frames D and F.
[0057] The cooling power of the AGRR stage at the cold temperature,
the hot sink temperature, the thermal load temperature, the heat
transfer fluid pressure, the effective tube wall temperature change
of the cold heat transfer fluid leaving the cold end of the dual
active regenerators, and the .phi. factor that specifies the
fraction of the effective tube wall temperature change that can be
used for a given heat transfer fluid flow period are all design
variables that can be set. The flow rate of the He heat transfer
fluid, {dot over (m)}.sub.He, is calculated from the equation:
{dot over (Q)}.sub.C={dot over
(m)}.sub.Hec.sub.p.DELTA.T.sub.C.phi.
where {dot over (Q)}.sub.C is the cooling power at the cold
temperature, T.sub.C, c.sub.p is the heat capacity of He at
constant pressure, and .DELTA.T.sub.C is the effective tube wall
temperature change at the coldest row of refrigerant-filled tubes
in AGRR active regenerator. The typical variables for an active
regenerator with 50 W of cooling power at 240 K are presented in
Table 1 below.
TABLE-US-00001 TABLE 1 Typical AGRR stage specifications for
efficiency design calculations He heat Q.sub.Cdot T.sub.H T.sub.C
.DELTA.T.sub.C .PHI. He heat transfer gas transfer flow (W) (K) (K)
(K) (dim) pressure (psia) rate (g/sec) 50 290 240 10 0.5 200
1.93
[0058] In the preferred embodiment, a rectangular tube array is
used with a staggered tube arrangement in successive layers from
the cold end to the hot end of the dual active regenerators. The
length in the x direction is the active length of the
refrigerant-filled tubes, or the longitudinal axis of the
regenerator. The length in the y direction is the length of each
row of refrigerant-filled tubes where the number of tubes depends
upon the tube diameter and separation between each tube in a row.
The z direction is the heat transfer fluid flow direction and this
length is determined by the tube diameter and the separation
between the layers of rows of tubes. The total number of tubes in
the dual active regenerators on each side of the AGRR stage is the
product of the number of tubes in each row and the number of layers
of rows. In one embodiment, the length of x and y have the same
value and the z length has a value three times the y length,
although other dimensions are possible. The primary independent
variable becomes the tube diameter once the rectangular dimensions
of the dual regenerators are chosen with the constraints above.
[0059] The average He heat transfer gas properties at 200 psia can
be calculated at the average temperature of the dual active
regenerators. The density, heat capacity, viscosity, and thermal
conductivity can be obtained and used to calculate the Prandtl
number. The Reynolds number of the heat transfer fluid flow can be
determined using the accepted equation from the literature for this
geometry. For the various configurations, the heat transfer
coefficient, the friction factor, and the effective thermal
conductivity can be calculated. These values can then be used to
calculate the entropy generated from the three mechanisms described
above and the total entropy generated can then be used to obtain
the FOM of the AGRL.
[0060] Of immediate note is that the AGRR stage efficiencies ranges
from a relatively small value in the case with 10 total tubes of
0.635 cm (0.25'') outer diameter, 2.5 cm length, and a reasonable
.DELTA.T of 10 K to a very impressive value in the case where
0.15875 cm ( 1/16'') outer diameter, 7.5 cm length, and a higher
but achievable .DELTA.T of 15 K. These dimensions along with Table
2 are provided as examples of one embodiment only and should not be
construed to limit the scope of the disclosed invention.
TABLE-US-00002 TABLE 2 Calculated efficiencies for various AGRR
stage geometries with staggered tubes. AGRR dimensions are shown.
.DELTA.T.sub.C L.sub.x L.sub.y L.sub.z D.sub.t S.sub.y (K) (cm)
(cm) (cm) (cm) (cm) S.sub.z (cm) N.sub.tubes/row N.sub.layers/AGRR
N.sub.total Efficiency 15 7.5 7.5 22.5 0.3175 0.158 0.3175 15 35
525 0.77 15 7.5 7.5 22.5 0.1588 0.079 0.1588 31 70 2170 0.82
[0061] In the preferred embodiment, each of the dual regenerators
in each AGRR stage has the following characteristics:
[0062] high specific area (.about.10,000 m.sup.2/m.sup.3)
[0063] very thin wall tubes
[0064] appropriate .DELTA.T vs. T characteristics
[0065] mechanically strong enough for modest pressures
[0066] leak tight and able to withstand cyclic mechanical pressure
loads
[0067] low pressure drop
[0068] high transverse and low longitudinal (axial) conductivity
and
[0069] mass producible at low/modest cost.
[0070] As mentioned, one variable that can be manipulated to affect
the efficiency of the AGRR stage is the effective adiabatic
temperature change of the refrigerant-filled small diameter
thin-walled tubes. For example, a 5/32'' outer
diameter.times.0.003'' wall stainless tubing with a compression
pressure ratio of 3 has a possible .DELTA.T of about 19 K with a
mean operating temperature of 300 K and an initial He refrigerant
pressure of 215 psia.
[0071] In the preferred embodiment of a hydrogen AGRL, the thermal
load, {dot over (Q)}.sub.load, from the hydrogen process stream is
transferred via convective heat transfer to the heat transfer flow
stream for dual active regenerators from a continuous stream in the
circulator and heat rejection exchanger. The pressure drop of the
active regenerator and heat exchanger combinations can be kept
reasonably low by design choices, so this circulator needs high
volumetric efficiency at relatively low head. The extra work for
this component of each AGRR stage is modest so its effect on
overall efficiency of each stage and FOM of the AGRL is relatively
small. The operation of the active regenerator in this AGRR stage
is similar to that in any regenerative refrigerator with the added
feature that each refrigerant-filled tube in the active regenerator
has the ability to provide refrigeration rather than acting only as
a passive heat sink/source. The selection of many small diameter
tubes in the preferred embodiment creates a large surface area for
heat transfer between the refrigerant-filled tubes and the heat
transfer fluid, thereby reducing the dominant irreversible entropy
mechanism in the active regenerator, in each AGRR stage, and in the
AGRL as a whole.
[0072] The preferred embodiment of a hydrogen AGRL includes the
following features: [0073] Active regenerator tube arrays,
refrigerant, and passive micro-regenerator arrays for each AGRR
stage [0074] Heat transfer fluid periodic flow assemblies for each
AGRR stage [0075] mechanical drive for displacer or circulators
[0076] fluid transfer system pipes, heat exchangers, flow control
valves, seals, insulation, circulators and/or blowers [0077]
Compressor/expander assemblies for each AGRR stage [0078]
mechanical drive for the compressor/expander for each AGRR stage
[0079] System integration, heat exchanger assemblies, skid for
mounting all vessels, etc. [0080] cold box for He/H.sub.2 process
stream heat exchangers [0081] He/H.sub.2O heat sink heat exchangers
at .about.300 K [0082] He/H.sub.2 heat exchangers with
ortho-to-para hydrogen catalysts [0083] H.sub.2 supply and LH.sub.2
storage vessels [0084] Auxiliary systems [0085] electrical power
for the AGRL [0086] vacuum station for start-up and cryopump for
long-term operation [0087] Instrumentation and control [0088]
temperature, pressure, velocity, loads, power, and flow rate
transducers [0089] flammable gas detectors, shut-down valves,
process stream pressure and flow rate controls, level and pump
controls [0090] control panel, DAQ racks, PC and PC
interface/software.
Note that although the preferred embodiment would contain all of
the foregoing elements, one could manufacture an AGRL that
incorporates only some of the above.
[0091] A hydrogen AGRL also preferably has the characteristics
described in Table 3 for each AGRR stage:
TABLE-US-00003 TABLE 3 Design calculations of the six hydrogen AGRR
stages. Key parameters for each stage are: hi = 0.75, i = 1 6,
.DELTA.T.sub.ad = 15 K, .PHI. = 1/2, f = 1 Hz. He temperatures Heat
transfer fluid Total length of at the Cold HEX (He) flow rate
Regenerator tubes Stage T.sub.cold[K] .DELTA.T.sub.cold[K] [g/sec]
per stage [m] 1 200 7.85 0.39 41.0 2 134 4.95 0.40 81.6 3 90 3.19
0.43 113 4 60 2.07 0.53 141 5 40 1.36 0.66 150 6 27 0.91 1.62
281
[0092] With these total regenerator tube lengths, the geometries of
each AGRR stage can be obtained with choices of 1.sub.x, 1.sub.y,
and 1.sub.z, the respective lengths of the rectangular regenerator
as described above. Table 4 summarizes the selected geometries of
each AGGR stage of one embodiment of the AGRL with tubes of
diameter 0.3175 cm. Again, these geometries could be varied.
TABLE-US-00004 TABLE 4 Selected/calculated dimensions and the total
number of regenerator tubes of the six AGRR stages of a 1 kg/day
LH2 AGRL. AGRR Stage L.sub.x (cm) L.sub.y (cm) L.sub.z (cm)
Tubes/layer Total tubes 1 7.5 10 13 21 547 2 7.5 10 25 21 1088 3
7.5 10 34 21 1518 4 7.5 15 29 32 1884 5 7.5 15 30 32 1998 6 7.5 15
57 32 3753
[0093] FIG. 6 schematically illustrates the operation of the first
AGRR stage (12) of one embodiment of an AGRL configured as a
hydrogen liquefier. As described above, the He heat transfer fluid
system connects the dual active gas regenerators (20) of each AGRR
stage to a heat sink exchanger (17) and a process stream heat
exchanger (15). The operation of the circulator (32) is
synchronized with the reciprocating motion of the pistons (28) of
the compressor/expander assembly (24) to pump heat from the process
stream (11) at the cold heat exchanger (14) via the process stream
heat exchangers (15) to the heat rejection exchanger (13) via the
heat sink exchangers (17). A three-way rotary valve (36)
facilitates this operation, as will be described here in detail.
The three-way valves are white if open for flow and black if closed
to flow. Note that to simplify depiction of the flow paths in FIG.
6, the cold ends of the dual active regenerators (20) are at the
ends nearest each other in the figure in contrast to that
orientation shown in FIG. 2.
[0094] For purposes of this description, the refrigerant (23) in
the upper dual active regenerator (20) of FIG. 6 has been
compressed, resulting in a temperature increase in the
refrigerant-filled tubes (21). The refrigerant (23) in the lower
dual active regenerator (20) of FIG. 6 has been expanded, resulting
in a temperature increase in the refrigerant-filled tubes (21). The
heat transfer fluid (35) enters through the lower regenerator (20)
and flows toward the top of the figure, executing a cold blow.
During the cold blow, the lower regenerator (20) accepts heat from
the He heat transfer fluid (35), thereby lowering the temperature
of the heat transfer fluid (35) to about
T.sub.cold-.DELTA.T.sub.cold. This cold heat transfer fluid (35)
then circulates through the cold heat exchanger (14) thermally
coupled to the process stream heat exchanger (15), thus accepting
heat from the thermal load of the hydrogen process stream (11) as
explained above. After accepting heat from the process stream (11),
the He heat transfer fluid (35) returns through the upper active
regenerator (20) of this AGRR stage and flows toward the top of the
figure, executing a hot blow. During the hot blow, the heat
transfer fluid accepts heat from the upper regenerator (20),
thereby elevating the temperature of the heat transfer fluid (27)
to about T.sub.hot+.DELTA.T.sub.hot at the exit of the upper dual
active regenerator. The heat transfer fluid (35) then circulates
through the heat rejection exchanger (13), where the cooling fluid
(16) accepts heat from the AGRR stage.
[0095] Each of the three-way valves (36) can be switched, and the
pistons (28) of the compressor/expander expand the refrigerant (23)
in the upper active regenerator (20) and compress the refrigerant
(23) in the lower regenerator (20). The use of four three-way
valves (36) eliminate the necessity of reversing the flow of the
heat transfer fluid (35) through the system, insure counterflow
fluid flows in both the cold heat exchanger and heat rejection
exchanger at all times, and eliminate two heat exchangers as
depicted in the embodiment of FIG. 2. These valves (36) are
positioned on either side of the active regenerators (20).
[0096] The second half of the cycle is analogous to the first half
of the cycle. In this instance the heat transfer fluid (35) flows
through the upper regenerator (20) toward the bottom of the figure
and is cooled by the refrigerant-filled tubes before passing
through the cold heat exchanger (14) thermally coupled to the
process stream heat exchanger (15) to accept heat from the process
stream (11). The heat transfer fluid (35) then circulates through
the lower regenerator (20), flowing toward the bottom of the
figure, where it accepts heat from the refrigerant-filled tubes.
This heat is expelled when the heat transfer fluid (35) passes
through the heat rejection exchanger (13) that is thermally coupled
to the heat sink exchanger (17). The valves (36) are then switched
back to the configuration in FIG. 6. This sequence of events is
continuously repeated at the frequency of about 1 Hz. The layout of
each stage of an AGRL will be approximately the same except the
temperature spans across the subsequent AGRR stages will increase
as the process stream is further cooled toward liquefaction.
[0097] Various heat exchangers may be used for the AGRL. In one
embodiment, small brazed-plate heat exchangers are used for the
process stream heat exchangers. The heat rejection and heat sink
exchangers may be compact plate-fin liquid-to-gas exchangers that
have been used successfully in previous LNG liquefier designs and
are available in the prior art. The duty of these exchangers is
significantly more than the process stream heat exchangers. In the
preferred embodiment, each AGRR stage rejects heat to a common
water stream (the cooling fluid) that subsequently rejects heat to
the environment through a water to air heat exchanger. The heat
exchanger for removing excess heat from the cooling fluid can take
the form of a finned tube with air cross flow as used in many air
conditioners or similar vapor compression cycle refrigerators.
These two operations are well defined and understood by those
skilled in the art.
[0098] As mentioned, the preferred embodiment includes a cold box.
Preferably, this cold box is vacuum insulated and of a simple
design with a top plate for mounting the six stages of the AGRL.
The assembly can be done easily with a crane to raise or lower the
AGRL into the cold box. In this embodiment, the cold box is
evacuated with a high vacuum turbo pump prior to operation. The
cold box vacuum can be maintained by Cryopumping with zeolite
containers attached to the cold end of one or the stages (e.g., an
about 40 K stage because zeolite will adsorb helium at this
temperature). Superinsulation is wrapped on all the cryogenic
sections of the AGRL to reduce radiative heat leaks. All
instrumentation can be mounted to the individual AGRR stages and
the process stream may enter through vacuum feed throughs in the
top plate of the cold box. The controls for the flows, the
compressor/expanders, all valve drive motor controllers, and other
operational components can be located outside the cold box.
[0099] In the preferred embodiment, the cryogenic liquid produced
(i.e., LNG or LH.sub.2) is stored in a small vessel within the cold
box with a double-walled vacuum jacketed or suitably insulated
transfer line out of the vessel through the top plate for storage
in an external cryogenic storage vessel. A level detector and or a
mass flow meter may be placed in the vessel to directly measure the
rate of liquefaction.
[0100] While the invention is susceptible to various modifications
and alternative forms, specific embodiments thereof have been shown
by way of example in the drawings and herein described in detail.
For example, although the refrigerant has been described throughout
as helium gas, any fluid that can be compressed or expanded to
cause a temperature change in the desired range may be used. Other
mixed refrigerants or combined compressor/expander assemblies are
also possible. One skilled in the art will recognize suitable
substitutions. Similarly, the dimensions of the various components
may be varied. The AGRL may be configured to liquefy other fluids.
It should be understood, however, that it is not intended to limit
the invention to the particular forms disclosed, but on the
contrary, the intention is to cover all modifications, equivalents,
and alternatives falling within the spirit and scope of the
invention as described.
* * * * *