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Description  |
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DESCRIPTION
1. Field of the Invention
This invention relates to cryogenic refrigeration systems such as split
Stirling cryogenic refrigerators. In particular it relates to such systems
having displacers and/or compressors driven by linear motors.
2. Background
A conventional split Stirling refrigeration system is shown in FIGS. 1-4.
This system includes a reciprocating compressor 14 and a cold finger 16.
The piston 17 of the compressor provides a nearly sinusoidal pressure
variation in a pressurized refrigeration gas such as helium. The pressure
variation in a head space 18 is transmitted through a supply line 20 to
the cold finger 16.
The usual split Stirling system includes an electric motor driven
compressor. A modification of that system is the split Vuilleumier. In
that system a thermal compressor is used. This invention is applicable to
both of those refrigerators as well as to others such as Gifford-McMahon
refrigerators.
Within the housing of the cold finger 16 a cylindrical displacer 26 is free
to move in a reciprocating motion to change the volumes of a warm space 22
and a cold space 24 within the cold finger. The displacer 26 contains a
regenerative heat exchanger 28 comprised of several hundred fine-mesh
metal screen discs stacked to form a cylindrical matrix. Other
regenerators, such as those with stacked balls, are also known. Helium is
free to flow through the regenerator between the warm space 22 and the
cold space 24. As will be discussed below, a piston element 30 extends
upwardly from the main body of the displacer 26 into a gas spring volume
32 at the warm end of the cold finger.
The refrigeration system of FIGS. 1-4 can be seen as including three
isolated volumes of pressurized gas. A working volume of gas comprises the
gas in the space 18 at the end of the compressor, the gas in the supply
line 20, and the gas in the spaces 22 and 24 and in the regenerator 28 of
the cold finger 16. A second volume is a relatively large dead space in
the compressor including the space 25 behind the compressor piston 17. It
is sealed from the working volume by the piston seal 27. The third volume
of gas is the gas spring volume 32 which is sealed from the working volume
by a piston seal 34 surrounding the drive piston 30.
Operation of the conventional split Stirling refrigeration system will now
be described. At the point in the cycle shown in FIG. 1, the displacer 26
is at the cold end of the cold finger 16 and the compressor is compressing
the gas in the working volume. This compressing movement of the compressor
piston 17 causes the pressure in the working volume to rise from a minimum
pressure to a maximum pressure and this warms the working volume of gas.
The pressure in the gas spring volume 32 is stabilized at a level between
the minimum and maximum pressure levels of the working volume. Thus, at
some point the increasing pressure in the working volume creates a
sufficient pressure difference across the drive piston 30 to overcome
retarding forces, including a pressure differential across the displacer
and the friction of displacer seal 36 and drive seal 34. The displacer
then moves rapidly upward to the position of FIG. 2. With this movement of
the displacer, high-pressure working gas at about ambient temperature is
forced through the regenerator 28 into the cold space 24. The regenerator
absorbs heat from the flowing pressurized gas and thereby reduces the
temperature of the gas.
With the sinusoidal drive from a crank shaft mechanism, the compressor
piston 17 now begins to expand the working volume as shown in FIG. 3. With
expansion, the high pressure helium in the cold space 24 is cooled even
further. It is this cooling in the cold space 24 which provides the
refrigeration for maintaining a temperature gradient of over 200 degrees
Kelvin over the length of the regenerator.
At some point in the expanding movement of the piston 17, the pressure in
the working volume drops sufficiently below that in the gas spring volume
32 for the gas pressure differential across the piston portion 30 to
overcome retarding forces such as seal friction. The displacer 26 is then
driven downward to the position of FIG. 4, which is also the starting
position of FIG. 1. The cooled gas in the cold space 24 is thus driven
through the regenerator to extract heat from the regenerator.
It has been understood that the phase relationship between the working
volume pressure and the displacer movement is dependent upon the braking
force of the seals on the displacer. If those seals provided very low
friction, it had been understood that the displacer would move from the
lower position of FIG. 1 to the upper position of FIG. 2 soon after the
working volume pressure increased past the pressure in the spring volume
32. Because the spring volume is at a pressure about midway between the
minimum and the maximum values of the working volume pressure, movement of
the displacer would take place during the midstroke of the compressor
piston 17. This would result in compression of a substantial amount of gas
in the cold end 24 of the cold finger, and because compression of gas
warms that gas this would be an undesirable result.
To increase the efficiency of the system, upward movement of the displacer
is retarded until the compressor piston 17 is near the end of a stroke as
shown in FIGS. 1 and 2. In that way, substantially all of the gas is
compressed and thus warmed in the warm end 22 of the cold finger, and that
warmed gas is then merely displaced through the regenerator 28 as the
displacer moves upward. Thus, the gas then contained in the large volume
24 at the cold end is as cold as possible before expansion for further
cooling of that gas. Similarly, it is preferred that as much gas as
possible be expanded in the cold end of the cold finger prior to being
displaced by the displacer 26 to the warm end. Again, the movement of the
displacer must be retarded relative to the pressure changes in the working
volume.
In prior systems, the seals 34 and 36 are designed and fabricated to
provide an amount of loading to the displacer to retard the displacer
movement by an optimum amount. A major problem of split Stirling systems
is that with wear of the seals the braking action of those seals varies.
As the braking action becomes less the displacer movement is advanced in
phase and the efficiency of the refrigerator is decreased. Also, braking
action can be dependent on the direction of the pressure differential
across the seal.
In addition to the problem of wear of the seals, the refrigerator is often
subjected to different environments. For example, a refrigerator may be
stored at extremely high temperature and be called on to provide efficient
cyogenic refrigeration. On the other hand, the refrigerator may be subject
to very cold environments. The sealing action and friction of the seals is
generally very dependent on temperature.
Due to the problems associated with synchronizing the regenerator movement
with the pressure waves from the compressor, efforts have been made to
utilize linear drive motors rather than the pneumatic drive discussed
above. An example of such a system can be found in U.S. Pat. No. 3,991,586
to Acord. That system also utilizes clearance seals and thus avoids the
problems associated with wear of conventional seals. The problem
associated with such a linear motor system is that the linear drive motor
is bulky and heavy and generates heat at the cold finger portion of the
refrigerator. In a split Stirling refrigerator, it is often critical that
the cold finger portion of the refrigerator be minimized in size and
weight and that little heat be generated in that portion of the system. It
is for those reasons that the pneumatic displacer drive has been so widely
used in split Stirling systems.
The Acord patent shows a rotary compressor which is typically used in split
Stirling refrigerators. Recent efforts have been made to use linear
compressors in such refrigeration systems because the low number of parts
in such systems make linear motors very cost effective.
DISCLOSURE OF THE INVENTION
This invention relates to several improvements in linear drive motors which
can be utilized in either a compressor or a displacer drive. A primary
improvement is that the motor includes a permanent magnet mounted to the
moving armature of the motor which in turn drives a piston element. The
permanent magnet is surrounded by a stationary housing which supports one
or more drive coils. This structure allows for a compact, clean assembly
in which the gas refrigerant is not exposed to organic materials on the
drive coil. Further, the moving magnet arrangement allows for minimization
of radial forces on the armature without unduly reducing the efficiency of
the motor. Minimization of radial forces is particularly important where
gas lubricated clearance seals provide the only mechanical support for the
armature as in the present design.
In further improvements of the linear motors of the present design, the
magnetic flux paths may be adjusted to assure that the magnetic axis and
mechanical center of the armature and the magnetic axis of the surrounding
coils are the same. In one adjustment, a flux return is generally
cylindrical and is split along its length into two elements. The relative
angular positions of the two elements can be adjusted about the center of
the flux return path in order to adjust the magnetic axis of the path. In
the embodiments disclosed, that split flux return path surrounds the drive
coils of the stator. The magnetic axis of the armature can be adjusted by
adjusting the angular positions of a number of permanent magnets on the
armature. By adjusting the armature magnets and the stator flux return
paths together, the two magnetic axes of each can be brought into
collinearity with the mechanical axis of the armature.
In order to avoid contamination of the gas refrigerant, an object of the
present design has been to avoid the use of organic materials, such as
epoxy, in any of the working volume, compressor dead space, and cold end
spring volume. To that end, in one motor, a permanent magnet assembly is
joined to the armature by an expansion element. Specifically, a number of
radial permanent magnetics are positioned about an expansible sleeve and
retained on that sleeve by an outer retainer ring. The sleeve is expanded
by a tapered collet which also presses inward against a sleeve which grips
the driven piston element. In another drive motor, an axial magnet is
retained on the driven piston element by a screw threaded connector,
specifically a threaded flux return path.
BRIEF DESCRIPTION OF THE DRAWINGS
The foregoing and other objects, features and advantages of the invention
will be apparent from the following more particular description of
preferred embodiments of the invention, as illustrated in the accompanying
drawings in which like reference characters refer to the same parts
throughout the different views. The drawings are not necessarily to scale,
emphasis instead being placed upon illustrating the principles of the
invention.
FIGS. 1-4 illustrate operation of a conventional pnuematically driven split
Stirling refrigerator;
FIG. 5 is a side view of a compressor in a split Stirling refrigerator
embodying this invention partially in section to show the linear motor
assembly;
FIG. 6 is an exploded view of the armature assembly in the compressor of
FIG. 5;
FIG. 7 is a side view, primarily in section, of the tapered sleeve in FIG.
6;
FIG. 8 is an exploded view of the primary stationary parts, other than the
drive coil assembly, in the compressor of FIG. 5;
FIG. 9 is a perspective view of an alternative embodiment of the outer,
flux return casing of the compressor of FIG. 5;
FIG. 10 is an elevational, cross sectional view of the displacer drive
motor of a split Stirling refrigerator embodying this invention.
DESCRIPTION OF PREFERRED EMBODIMENTS
The preferred compressor design is illustrated in FIG. 5. This compressor
comprises dual reciprocating piston elements 21 and 23 which, when driven
toward each other, compress helium gas in the head space 39. The gas can
pass through a side port 29 in a compression chamber cylinder 31 to an
outer annulus 33 in that cylinder. The gas from the annulus 33 can pass
through an outer housing 35 to a tube fitting hole 37. A tube (not shown)
joined at the fitting hole 37 delivers the gas to a cold end section of
the split Stirling refrigerator. Preferably, the pistons 21 and 23 and the
compression chamber cylinder 31 are of cermet, ceramic or other hard, low
friction material. The elements are close fitting to provide clearance
seal therebetween.
The pistons 21 and 23 serve as the sole mechanical support for respective
armatures of linear drive motors. Identical motors drive the two pistons.
The right hand motor is shown in FIG. 5 and is described below with
reference to FIGS. 5 through 8.
A sleeve 38 (FIGS. 6 and 7) is joined to the piston 21 at its far end from
the compressor head space 39. The sleeve 38 has an inner clearance 41 such
that it is free to shuttle back and forth along the compression chamber
cylinder 31 without contacting that cylinder. The sleeve 38 has a tapered
flange 40 at its left end. An expanding sleeve 42, placed on the sleeve 38
from the right, abuts the flange 40. The expanding sleeve 42 is also an
inner flux return and should have a high magnetic permeability. It
supports two sets of radial permanent magnets 44, 46 spaced by a spacer
48. The six magnets in each set of permanent magnets are retained by
magnet retainer rings 50 and 52.
Although the magnets 44 and 46 are shown closely packed in FIG. 6, they are
preferably dimensioned such that, when placed about the expanding sleeve
42, spaces remain between the magnets. With that arrangement, helium gas
in the dead space 54 of the compressor is free to flow between the magnets
as the drive motor armature and compressor piston assembly shuttles back
and forth. Further, dissimilarities in the magnet elements might cause the
magnetic axis of the grouped magnets to be offset from the mechanical
center of the piston 21. The magnetic axis can be made the same as the
mechanical center by adjusting the relative angular positions of the
magnets about the expanding sleeve 42. Such an offset of the magnetic axis
from the mechanical center would result in radial forces on the piston 21
which would tend to bind the piston within the cylinder 31. The
elimination of radial forces is particularly important in this compressor
design where the sole mechanical support for the armature is the piston 21
within the cylinder 30.
As shown in FIG. 6, the expanding sleeve 42 has slots 60 formed therein to
allow for expansion. To fix the permanent magnets 44 and 46 in position on
the armature, a tapered collet 56 is wedged between the expanding sleeve
42 and the tapered sleeve 38 by a nut 58. As the nut 58 is tightened on
the sleeve 38, the expanding sleeve is pressed outward by the tapered
flange 40 and the collet 56. The expanding sleeve in turn presses the
permanent magnets 44 and 46 against the magnet retainer rings 50 and 52.
The tapered sleeve 38 has two slots 59 and 61 formed in the end thereof so
that, as the collet 56 presses outward against the expanding sleeve 42, it
also presses inward and compresses the sleeve 38 to form a tight joint
between the sleeve and the piston 21. The use of the expansion and
compression joints in the armature avoids the need for any epoxy or other
adhesive which might contaminate the helium gas.
The armature assembly just described is enclosed and hermetically sealed
within the housing 35 best shown in the exploded view of FIG. 8. The
central raised portion 62 of the compression chamber cylinder 31 is
retained within an annulus 64 of the housing 35 when that housing is
joined end to end with a left hand housing 66. Each of the housings is
provided with a number of bolt holes 68 through which bolts extend to join
the housings. The housings also have a number of gas port holes 70 through
which helium gas flows in the dead space behind the two pistons elements
21 and 23. Before the housings 35 and 66 are assembled, coil supporting
bobbins 72 and 74 (FIG. 5) are fitted over the housings. Bobbin 74, for
example, supports two coils 76 and 78 and Hall effect position sensors
(not shown) between the two coils. The bobbins are spaced apart by a
spacer 77. The spacer 77 is a ring which surrounds the two housings 35, 66
and is split to permit the tube fitting to be connected to the fitting
hole 37.
The compressor assembly is completed by positioning end covers 80 and 82 at
the ends of the housings 66 and 35 to close the helium gas space, and by
positioning half sleeves 84 (FIG. 8) and 86 (FIG. 5) over those end covers
and the coil assemblies. Set screws through the holes 88 in the cover
sleeves 84 and 86 are seated in grooves 87 and press the end caps 80 and
82 tightly against indium seals 90 and 92 (FIG. 5) and the housings 66 and
34.
The end cap 82 includes an assembly which allows for charging of the
housing with helium gas. Specifically, a ball 94 closes a port 96 in the
end cover 82. The ball is retained against the port by a retainer screw 98
and is protected from contamination by a plug 44.
The outer cover sleeves 84 and 86 serve to both enclose the coil assemblies
and hold the end covers in place. Further, they are of ferromagnetic
material and serve as flux return paths for the coils.
Because the half sleeves 84, 86 do not completely surround the compressor
assembly, they may be shifted angularly relative to each other about the
coils to adjust the flux paths and thus adjust the magnetic axis of the
coil assembly. Such adjustment of the flux return paths can be used to
particular advantage in the embodiment of FIG. 9 wherein a separate set of
outer covers is positioned over each linear motor of the compressor.
Specifically, outer flux return covers 102 and 104 surround the right hand
assembly, and covers 106 and 108 surround the left hand assembly. For each
motor assembly, by adjusting the armature magnets and the flux return
sleeves, the magnetic axes of the armature and coil assemblies can both be
collinear with the mechanical axis of the armature assembly. The outer
flux returns of FIG. 9 load the end covers 80 and 82 in a similar fashion
as in FIG. 5, and the tension would be absorbed by a modification to
spacer 77.
It can be noted that the coils are positioned entirely outside of the
helium gas environment. With that arrangement, there is no danger of
contamination of the helium gas by the potting material along the coils.
Further, no electrical feedthroughs into the helium environment are
required. With all coils outside of the helium enclosing housing, a
compact, easily constructed assembly is provided and heat generated in the
coils is most easily removed from the system.
The use of permanent magnets in the linear motors reduces the power
requirements of the motors. By positioning those permanent magnets on the
armature, rather than on the stator with the coils, a larger gap between
the ferromagnetic material of the armature and the ferromagnetic material
of the stator can be provided while retaining reasonable efficiency. Such
a wide gap reduces the radial forces on the armature which might tend to
bind the piston assembly within the compression chamber cylinder or cause
excessive wear of those parts.
The cold end displacer drive of the split Stirling refrigerator is shown in
FIG. 10. The cold end includes an outer cylindrical casing 150 fixed to
and suspended from a cold finger head 152. The opposite, cold end of the
cylinder 150 is not shown. A displacer 154, mounted for reciprocating
movement within the cylinder 150, includes a fiberglass epoxy cylinder
155. The cylinder 155 may be packed with nickel balls 156 sandwiched
between short stacks of screen 158 at each end of the regenerator. The
screen is held in place by a porous plug 160. The porous plug 160 is
positioned at the end of a bore 166 in a cermet clearance seal element
162.
The cermet clearance seal element 162 is fixed to the cylinder 155. It is
seated within a second cermet clearance seal element 168 to provide a
clearance seal 170. The clearance seal 170 is preferably a 0.00015 inch
(0.0038 millimeter) radial gap between the two cermet clearance seal
elements. The gap is half the diametrical clearance between the clearance
seal elements. That clearance seal allows for virtually dragless movement
of the element 162 within the element 168 while providing sealing between
the warm end 174 of the cold finger working volume and an annulus 176
between the cold finger cylinder 150 and the displacer cylinder 155. The
sealing action of the clearance seal is due to the small radial gap
extending axially along the approximately 0.25 inch (6 millimeter) length
of the seal.
Channels 180 are formed in the top of the clearance seal element 168 to
provide fluid communication between the warm end 174 of the displacer and
an annulus 182. The annulus 182 is connected to the compressor of FIG. 5
through a port 186 and fitting 187.
Another outer clearance seal element 188 is positioned within the cold
finger head 152. This element is also formed of cermet. The clearance seal
element 188 has a smaller inner diameter than the lower clearance seal
element 168. The element 188 forms a clearance seal 190 with a cermet
drive piston 192. The cermet piston 192 and the cermet clearance seal
element 188 are of nonmagnetic material, preferably cermet. Grooves 193
are provided in the cermet piston element 192 to minimize diametral
pressure force differentials which might otherwise cause side forces which
tend to bind the displacer.
The clearance seal element 188 is clamped against the seal element 168,
which in turn is pressed against a seal 189 and the cold finger head 152,
by a clamping nut 200. The nut 200 is screw threaded with the head 152 and
acts on the element 188 through a tapered washer 191 and seal ring 193.
The piston 192 reciprocates with the main body of the displacer, and in
fact the pressure differential across the drive piston serves to drive the
entire displacer. The piston 192 is joined to the cermet element 162 by
means of a pin 196 extending through a transverse slot 198 at the lower
end of the piston 192. This joint allows radial accomodation no greater
than 0.005 inch to occur between piston 192 and clearance seal element
162. By this means, the concentricity and alignment tolerances between
clearance seals 190 and 170 may be relaxed.
A samarium cobalt magnet 202, sandwiched between flux return plates 204 and
206, is mounted to the drive piston 192. The magnet 202 is an axially
polarized magnet which by the action of the plates 204 and 206 yields a
radial magnetic flux. Because the outer diameter of the motor assembly is
about ten times the stroke of this motor there is sufficient support space
for this construction. But in the compressor design of FIGS. 5-9 the outer
diameter is only about three times the stroke, and radially megnets were
required. The flux return plate 206 serves as the magnet connector by
means of its threaded connection to the end of the piston 192. An
expanding connector as used in the compressor could be used in this motor
as well, but the annular axial magnet allows for this simpler connection.
As before, however, no adhesive is required.
A gas spring volume 208, in which the armature assembly reciprocates, is
formed by a housing 210 which is screw threaded onto the cold finger head
152 against a metal spring seal 212. A bobbin 214, which supports two
coils 216 and 218 and a Hall effect position sensor 220, is positioned
over the housing 210. Flux returns 222 and 224, in the form of half
sleeves as in the compressor motor, are positioned around the coils 216
and 218. During assembly, these flux return half sleeves can be positioned
against the coils and their angular positions relative to each other can
be adjusted to assure that the magnetic axis of the flux returns is the
same as that of the armature. This is done by adjusting the flux return
relative angles while observing the drive current through the coils on an
oscilloscope. A minimum current indicates a minimum of friction force.
Once the flux returns 222 and 224 are properly positioned and secured by an
adhesive, they are enclosed by an outer housing 226 which is joined to the
spring volume housing 210 by a screw connection.
The linear motor of FIG. 10 is for the purpose of trimming the motion of
the displacer to assure that the displacer makes full strokes without
rapping the ends of the cold finger and that it moves in proper phase with
the pressure wave. A patent application filed Dec. 6, 1982 in the names of
Niels O. Young, Robert Henderson and Peter J. Kerney discloses feedback
and control circuitry designed for that purpose.
While the invention has been particularly shown and described with
reference to preferred embodiments thereof, it will be understood by those
skilled in the art that various changes in form and details may be made
therein without departing from the spirit and scope of the invention as
defined by the appended claims.
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