Cryogenic cooler

An improved cryogenic cooler 100 includes a flange 106 with an elongated pressure vessel 120 extending therefrom. The pressure vessel 120 is connected to the flange 106 at a proximal end thereof. The pressure vessel 120 is adapted to cool a surface in the proximity of the distal end thereof. Vibration isolation is provided at both proximal and distal ends of the elongated pressure vessel. A coupler 126 serves to maintain a gap between the distal end of the pressure vessel 120 and the surface at cryogenic temperatures. In a specific embodiment, the coupler 234 has a coefficient of thermal expansion which is less than the coefficient of thermal expansion of an end cap 224 on the pressure vessel. The coefficients of expansion are chosen to provide a tight slip fit between the cooler and the coupler at ambient temperatures and a very small continuous air gap at cryogenic temperatures. Another novel feature is the provision of an energy-absorbing ring 114 within the flange to dissipate vibration therein.

FIELD OF THE INVENTION 
The present invention relates to cooling systems for infrared devices. More 
specifically, the present invention relates to methods and apparatus for 
minimizing mechanical vibrations of cryogenic expansion engines. 
While the present invention is described herein with reference to 
illustrative embodiments for particular applications, it should be 
understood that the invention is not limited thereto. Those having 
ordinary skill in the art and access to the teachings provided herein will 
recognize additional modifications, applications, and embodiments within 
the scope thereof and additional fields in which the present invention 
would be of significant utility. 
DESCRIPTION OF THE RELATED ART 
Infrared detectors, high performance cameras and other devices typically 
require cryogenic cooling for optimal performance. Typically, a cryogenic 
cooler is inserted into a dewar (or housing) into which one or more 
detector elements are mounted. On receipt of light or heat energy, the 
detector elements provide a low level output signal which is carried out 
of the dewar to detector circuitry over very fine wires to reduce the 
conducted thermal load between the detector element and the warm connector 
surfaces on the exterior of the dewar. Any relative motion of the wires 
creates a change in the relative capacitance of the signal leads. The 
change in capacitance in the presence of current generates a false signal 
known in the art as "microphonics". 
Another source of microphonic signals is due to stress and strain between 
the cool surface of the dewar and the detector. The strain in the detector 
induces a piezoelectric effect which generates a false signal in the form 
a voltage in the detector element at the frequency of motion of the cool 
surface. 
Thus, in general, infrared devices are extremely sensitive to vibration. 
Unfortunately, current designs for a very efficient, effective and 
frequently used cooling system, the Stirling cycle cooler, tend to induce 
a considerable amount of vibration into the dewar. 
A Stirling cycle cooler is an efficient and compact cryogenic cooling 
device. The original Stirling cycle engine consisted of a compressor 
piston with a cylinder, an expansion piston with a cylinder, and a drive 
mechanism. The drive mechanism converted the rotary motion of a motor and 
crankshaft to a reciprocating motion of the two pistons. The two pistons 
were arranged to be ninety degrees out of phase. A regenerative heat 
exchanger (regenerator) was included in the expansion piston to thermally 
isolate gas at the compressor piston head space from gas at the expansion 
piston head space. The original Stirling cycle engine in which the 
compressor piston and expander piston are mechanically linked is known as 
the Integral Stirling Engine. 
When operated between two temperature sinks, the Stirling cycle mechanism 
can produce shaft power when operated in one direction (Stirling engine) 
or pump heat from a low temperature to a high temperature, and thereby 
provide refrigeration, when driven in the reverse direction (Stirling 
cooler). 
The original thermodynamic cycle developed early in the last century 
consisted of two isothermal processes and two constant volume processes. 
Early implementations of the Stirling cycle did not duplicate the 
distinctly separately phased motions of prior art designs, but were a 
combination of the various motions occurring simultaneously instead. That 
is, in the Stirling engine, the compression and the expansion processes 
occurred simultaneously in a sinusoidal manner in which the expander 
motion lagged the compressor motion by 90 degrees. 
Later developments produced the Split-Stirling cycle cooler. This device 
included all of the components of the integral Stirling cycle cooler, 
without mechanical linkage to the expander piston. This permitted the 
expander to be located remote from the compressor. The expander piston in 
this device was no longer driven by a connecting rod and crankshaft, but 
rather by means of an additional drive piston. The drive piston was 
attached to the warm end of the expander piston and protruded into a small 
cavity at the extreme end of the expander housing. This created a "spring 
volume" as the gas acted as a spring on the drive piston. 
The piston was sealed so that gas could not readily enter the spring volume 
from the expander side. The drive piston was pneumatically reciprocated by 
cyclic gas pressure changes produced by the compressor piston driven by 
the compressor crankshaft. The gas thus supplied to and withdrawn from the 
expander engine traveled through a supply tube commonly referred to as a 
transfer line. The two subassemblies were thus often interconnected with a 
sufficiently small diameter gas transfer line to effectively decouple 
vibration and motion of the expander subassembly from vibration and motion 
of the compressor subassembly. This was particularly of interest when 
detectors in dewars mounted on the expander subassembly were isolated by a 
gimballed mechanism from the compressor subassembly. This configuration 
provided a gimballing of the detector without introducing large, 
detrimental spring torques to the gimbal torque motors. This design 
permitted the compressor, which was large compared to the expander 
assembly, to be remotely located where available volume and heat rejection 
capability existed. 
A natural evolution of this design was the relocation of the relatively 
heavy regenerator from the reciprocating displacer piston to the 
(stationary) cold cylinder. Thus, the displacer piston could be hollowed 
out making it much lighter than prior designs. This provided a reduction 
in the vibration output of the expander. Notwithstanding these 
refinements, Stirling cycle coolers continue to vibrate the associated 
mounting structure and, in particular, to the detector assembly. 
Conventional solutions to this problem include: 
1) the use of a bellows mechanism between the cold tip of the dewar with 
the detector mounted thereto, and the cold tip of the expander, and 
2) the provision of a small air gap between the cold tip of the expander 
and the dewar cold well in which the detector is mounted. 
Unfortunately, the effect of the bellows has been limited inasmuch as the 
mechanism remains mechanically coupled. Accordingly, the low fundamental 
frequency components are coupled into the detector substrate area. 
The noncontacting air gap approach suffers from the fact that a large 
temperature gradient is developed between the warm end of the dewar and 
expander and the associated cold ends. Differences in the materials of 
construction of the dewar cold well and the expander cold cylinder result 
in a variation in the air gap due to differences in coefficient of thermal 
expansion of the materials as the temperature gradient is established. 
Thus, the air gap tends to increase from typically less than 1 mil at room 
temperature to greater than 2 mils at cryogenic temperatures depending on 
the length of the cold cylinder and the difference in the coefficient of 
thermal expansion between the materials of the expander cold cylinder and 
those of the dewar cold well. This complicates the design, making it 
difficult to maintain a consistent gap. To address this problem, some 
designers have added capacity to the cooler in order to maintain the 
performance of the detector. 
Thus, there is a need in the art for a technique for mitigating microphonic 
effects in a cryogenic cooler with minimal thermal loss. 
SUMMARY OF THE INVENTION 
The need in the art is addressed by the present invention which provides an 
improved cryogenic cooler. The invention includes a flange with an 
elongated pressure vessel extending therefrom. The pressure vessel is 
connected to the flange at a proximal end thereof. The pressure vessel is 
adapted to cool a surface in the proximity of the distal end thereof. A 
novel coupler is mounted between the distal end of the pressure vessel and 
the surface. In a specific embodiment, an end cap is mounted at the distal 
end of the pressure vessel and the coupler has a coefficient of thermal 
expansion which is less than the coefficient of thermal expansion of the 
end cap. The coefficients of expansion and relative dimensions are chosen 
to provide a tight fit between the cooler and the cooled surface at 
ambient temperatures and a consistent air gap at cryogenic temperatures. 
Another novel feature is the provision of an energy absorbing ring within 
the flange to dissipate vibration therein so that the elongated pressure 
vessel is vibrating isolated at both distal and proximal ends.

DESCRIPTION OF THE INVENTION 
The invention is described in an illustrative Split Stirling application to 
disclose the advantageous teachings thereof. FIG. 1 shows a typical 
Split-Stirling cycle cryogenic refrigerator 10. The refrigerator 10 
generally includes a compressor 12 having a cylinder 14. A compressor 
piston 16 is mounted in the compressor cylinder 14. A gas medium 
connection line 18 communicates with a cryogenic expander/cooler 22 and 
passes through a cryogenic heat exchanger 20. The cryogenic cooler 22 is 
in physical communication with a detector (not shown) at a cooled terminal 
cap 40. A prime mover (also not shown) drives the piston 16 and charges 
the refrigerator 10. 
Helium is compressed and the heat therein is rejected by the piston 16, the 
compressor cylinder 14 and the heat exchanger 20. Thereafter, the gas is 
forced through a regenerator 30 in the cooler 22 by the motion of a piston 
24 within the cooler 22 causing it to be cooled at constant volume. The 
cooler piston 11 is pneumatically driven. The cooler 22 incorporates a 
pneumatic spring volume 42 to provide reciprocating action of the cooler 
piston 11. 
The heat removed from the gas is stored in the regenerator. Cold gas from 
the regenerator is next expanded isothermally while absorbing heat from a 
surface to be cooled in contact with end cap 40. The gas is then forced 
back through the regenerator where it is heated at constant volume. The 
energy stored in the regenerator 30 is then transferred back to the gas. 
Heat rejection to ambient is provided by the heat exchanger 20. 
A more detailed view of the cooler 22 used with the Split Stirling 
refrigerator of the prior art is shown in FIG. 2. The cooler is surrounded 
by an outer thin-walled pressure vessel 44 and a cylindrical flange 46 
which comprises a housing structure. The outer pressure vessel 44 is a 
long, thin-walled, tubular structure extending between the flange 46 and 
the terminal cap 40. The pressure vessel 44, unlike the flange 46, is 
comprised of a poor thermal conductor such as stainless steel of 
approximately 0.005-0.008 inch thick. The function of the pressure vessel 
44 is to house the internal elements of the cooler 22. The flange 46 
facilitates the mounting of the cooler 22 as well as heat dissipation. The 
pressure vessel 44 is securely attached to the flange 46, typically by 
brazing. 
Mounted on the rear side (e.g., ambient side) of the flange 46, as shown in 
FIG. 2, is a cooler cap 48 which is a structural cover bolted in place by 
a plurality of fasteners 50. The cooler cap 48 houses each of the 
components mounted behind the flange 46. Penetrating the cooler cap 48 is 
a passageway 52 which interconnects with the gas medium connection line 18 
shown in FIG. 1. The passageway 52 provides a means for delivering the gas 
medium from the compressor 12 to the regenerator 30 and to the various 
other volumes within the cooler 22. 
Mounted immediately within a cylindrical displacer structure 24 is the 
regenerator 30 which is a cylindrical structure fashioned to fit within 
the displacer structure. The regenerator 30 is comprised of a porous 
matrix of screens 54 which permit the gas medium to flow therethrough. The 
gas flowing through the porous matrix of screens 54 either absorbs the 
latent heat from the regenerator matrix or deposits latent heat into the 
high thermal enthalpy material comprising the porous matrix. Therefore, 
the exiting gas is either precooled or preheated, depending upon the 
direction of the gas flow. The screens 54 are typically comprised of a 
fine mesh material such as stainless steel, for example. In the assembly, 
the screens 54 are stacked on top of each other so that the layers are 
arranged perpendicular to the flow directions of the gas medium. 
Generally, the gas medium is pumped in from the compressor and enters the 
ambient (warm) end of the cooler 22. From there, the gas medium enters the 
regenerator 30 from the passageway 52. The gas is precooled by 
progressively cooler screens 54 that are stacked in the regenerator so 
that when the gas exits into an expansion volume 58 at the cold end of the 
cooler (e.g., terminal cap 40 as shown in FIG. 2), the gas is nearly at 
the expansion temperature. The terminal cap 40 is the coldest part of the 
cooler 22 and is that portion that is in mechanical communication with the 
detector device that is to be cooled (not shown in FIG. 2). The terminal 
cap 40 is comprised of a metal having a high thermal conductivity and may 
be fashioned from, for example, pure nickel or copper. 
The displacer 24 is comprised of a thin-walled fiber-glass shell 56, an 
enclosed regenerator 30, a cold end closure 64A and a warm end closure 64 
which couples the displacer 24 to a drive piston 72 and a displacer seal 
68 described hereinafter as shown in FIG. 2. The displacer fiberglass 
shell 56 acts as an insulating structural body which prevents heat flow 
from the ambient end to the cold end of the displacer 24 while the gas 
medium flows between the expansion volume 58 and the warm passageway 52. 
It is this fiberglass shell that reciprocates within the pressure vessel 
and which is closed at the cold end by an end cap 64A. In general, the gas 
medium is moved from the ambient end to the cold end of the cooler 22 
during a first stroking motion, and the gas medium is moved from the cold 
end to the ambient end during a second stroking motion. During the 
stroking motions, the gas medium is forced to flow through the displacer 
24 and the regenerator 30. 
Mounted within the interior of the flange 46 is a seal sleeve 66. Just 
inboard of the seal sleeve 66 is the displacer seal 68. The displacer seal 
68 and the warm end closure 64, when assembled within the seal sleeve 66, 
function to seal the sliding displacer 24 so that the gas medium cannot 
flow through the annular space between the displacer 24 and the 
thin-walled pressure vessel 44. Thus, the gas is forced to flow through 
the displacer 24 and through the porous screens 54 of the regenerator 30. 
Connected to the warm end closure 64 by a hinge pin 70 is the small drive 
piston 72. The hinge pin 70 is a small metal pin that passes through and 
retains the drive piston 72 to the warm end closure 64, providing good, 
flexible alignment between the piston 72 and the displacer 24. The drive 
piston 72, also known as a plunger, provides the area differential of the 
two displacer ends necessary to provide the motive forces to the displacer 
24. Thus, under the appropriate conditions, the displacer 24 strokes onto 
or away from the drive piston 72. This is accomplished by virtue of a 
pressure differential that exists across the displacer walls. The 
clearance space between the drive piston 72 and the interior of the cooler 
cap 48 is sealed by a piston sleeve 74. The piston sleeve 74 acts to guide 
the drive piston 72 and to prevent substantial gas leakage into or out of 
the spring volume 42 within the cooler cap 48. 
A displaced (swept) volume 76 exists between the piston sleeve 74 and the 
displacer warm end closure 64. The swept volume is a clearance which 
permits the displacer 24 to stroke to the ambient end of the cooler, the 
displacer 24 being shown at the mid-position in FIG. 2. A sealed clearance 
78 in the form of a small annular space is located between the drive 
piston 72 and the piston sleeve 74. The clearance is utilized to seal the 
displacer 24 from the oscillations of the spring volume 42. Mounted at the 
end of the small drive piston 72 is a bumper 80 that extends into the 
pneumatic spring volume 42. The bumper 80 is comprised of a steel core 
with a rubberlike material affixed thereon. The bumper 80 functions to 
strike the piston sleeve 74 and to stop the displacer 24 from impacting 
the tip of the cold terminal cap 40 when the piston 72 strokes from the 
ambient end to the cold end. Such impact would otherwise generate 
mechanical vibrations that would be transmitted to the detector device. 
When the piston strokes from the cold end to the ambient end, the bumper 
80 serves to cushion the piston 72 from impact with the inside of the 
cooler cap 48. 
Under steady state conditions, the forces within the cooler 22 are balanced 
and reverse quickly enough so that the displacer 24 never strokes to the 
displacer limits or impacts the bumper 80. A centering spring 82, as shown 
in FIG. 2, serves to prevent the displacer 24 from drifting too close to 
either end of its stroke. However, during the cool-down periods, while the 
working fluid (helium gas) is still warm, stroking of the displacer 24 is 
more severe due to lower fluid pressure drop through the regenerator 30, 
and the bumper 80 is typically impacted by the displacer 24. The pressure 
wave produced by the compressor 12 is sinusoidal in nature so that the 
pressure in the various cooler volumes varies sinusoidally. However, the 
gas pressure within the spring volume 42 is roughly constant and is at 
approximately the mean pressure point of the oscillating pressure wave. 
In practice, a small volume of gas medium leaks past the drive piston 72 
through the sealed clearance 78. As the pressure wave varies sinusoidally, 
a state of equilibrium is established in the spring volume 42. Such a 
condition is characterized by equal leakage in both directions of the 
sealed clearance such that the pressure in the spring volume equals the 
mean pressure of the oscillating pressure wave. This mean pressure is with 
respect to the pressure of the swept volume 76 and the expansion volume 
58, each of which experiences the cyclic pressure fluctuations. The 
centering spring 82 connected between the piston sleeve 74 and the bumper 
80, although not essential, is utilized for aligning the displacer 24 at 
the midpoint of the strike. Such a design is useful for preventing the 
displacer 24 from impacting the extreme ends of the stroke cycle. It is 
noted that the relative force generated by the spring volume 42 is much 
greater than the alignment force created by the centering spring 82. 
In operation, the cooler 22 is pressurized with the sinusoidal pressure 
wave so that the pressure rises from some minimum to some maximum 
pressure. The expansion volume 58 and the warm swept volume 76 are then 
pressurized, and a pressure force is established. As gas flows from the 
warm swept volume 76 to the cold expansion volume 58 through the 
regenerator 30, a small pressure difference is established between these 
two volumes. This pressure difference is in opposition to the flow of gas 
through the regenerator 30. Hence, as working gas pressure increases, gas 
flows from the warm swept volume 76 to the cold expansion volume 58. The 
gas pressure, P.sub.W, in the warm swept volume 76 will be somewhat higher 
than the gas pressure, P.sub.c, in the cold expansion volume 58. 
The force balance equation for the displacer neglecting the effect of the 
spring 82 is: 
EQU F=(P.sub.C -P.sub.W)A.sub.D +(P.sub.W -P.sub.S)A.sub.DP [ 1] 
where P.sub.C, P.sub.W and P.sub.S are the respective pressures within the 
cold expansion volume 58, the warm swept volume 76 and the spring volume 
42 and A.sub.D and A.sub.DP are the respective cross-sectional areas of 
the displacer 24 and the drive piston 72. In practice, the pressure drop 
P.sub.C -P.sub.W across the regenerator 30 is small compared with the 
pressure drop P.sub.W -P.sub.S across the drive piston 72. 
Hence, when P.sub.W is greater than P.sub.S, then F is positive, implying a 
net force towards the warm end. It can be seen that the forces reverse as 
the exterior pressure fluctuates during the cycle and that the inertia of 
the displacer is the only opposition to the pressure forces. 
Therefore, when the magnitude of the sinusoidal pressure wave is high, the 
displacer 24 strokes from the cold end to the ambient end. The displacer 
24 continues to stroke from the cold end to the ambient end until the 
bumper 80 impacts the cooler cap 48, or until the sinusoidal pressure has 
dropped sufficiently to reverse the force balance as the compressor 12 
begins to withdraw gas from the cooler 22 during the suction stroke. Thus, 
the gas medium is initially pumped into and then withdrawn from the cooler 
22. The varying gas pressure within the expansion volume 58 begins to drop 
and when the varying pressure drops below the mean point constant pressure 
of the spring volume 42, the forces reverse. While the pressure force 
summation may have reversed direction, the kinetic energy may still cause 
the cooler to continue to move in opposition to that force briefly during 
the cycle. During the steady state under cooled-down operation, the forces 
and the stroke are so designed as to permit the cooler 22 to stroke nearly 
to the limits, but not enough to impact the bumper 80 at either end of 
travel. 
Thus, when the magnitude of the sinusoidal pressure wave is low (e.g., gas 
pressure in the spring volume 42 exceeds the mean value of the oscillating 
pressure wave), the reciprocating drive piston 72 causes the displacer 24 
to stroke from the ambient end to the cold end of the cooler 22. Then, the 
entire cycle repeats with the net effect being that the displacer 24 
strokes from the cold end to the ambient end when the pressure in the 
cooler is high, and from the ambient end to the cold end when the pressure 
in the cooler is low. This constitutes net work performed by the gas in 
the expansion volume on the displacer for providing an equivalent 
refrigeration rating. By performing work on the displacer 24, the gas 
transmits energy to the displacer and a portion of this energy, in turn, 
is simultaneously deposited back into the gas at the opposite (ambient) 
end of the displacer 24. This work expenditure simultaneously lowers the 
temperature of the terminal cap 40 for cooling a detector. 
With the design and operation of a conventional cryogenic cooler in mind, 
the advantages of the present invention may be illustrated with respect to 
FIG. 3 which is a simplified representation of the conventional cooler 22 
of FIG. 2. For infrared detector cooling applications, the cooler 22 is 
mounted so that the cooler stem rests within the well of the dewar housing 
84 on a felt pad 86. The felt pad is loaded with a conducting (silver 
loaded) grease to facilitate the flow of heat energy thereacross. Heat is 
drawn from an infrared detector 90 through the dewar wall and the felt pad 
86. The cooler 22 has a solid annular one-piece flange 46 which is secured 
to the dewar flange by screws 89. The cooler 22 rests on an O-ring 92. The 
cooler flange 46 is integral with the cooler cap 48. 
Because the cooler 22 is connected to the detector 90 by the screws 88 and 
89 and is in contact with the vibrating dewar at the terminal cap 40, the 
mechanical vibrations created by the reciprocating mass are directly 
transmitted to the detector 90 resulting in optical deflection and 
distortion of the images received thereby. Vibrations also couple into low 
signal level electrical leads 91 resulting in relative motion of these 
leads and false signals associated with capacitance changes of the leads 
in the presence of a signal current. The problem becomes particularly 
acute in higher wattage coolers which provide higher net refrigeration 
capacities but which necessarily require a larger displacer piston 24. 
FIG. 4 is a simplified cross-sectional view of the improved expander/cooler 
100 of one embodiment of the present invention. As shown in FIG. 4, the 
cooler is essentially the same as the conventional cooler with the 
exception that the cooler 100 provides vibration decoupling through 
implementation of two-part flange and a coupled end cap having components 
of different coefficients of thermal expansion and specific configurations 
that take advantage of the different coefficients of thermal expansion. 
The cooler 100 has an elongated body 120 and is adapted for use with a 
conventional elongated dewar 102 having a dewar flange 104 and a well 122. 
The flange 106 of cooler 100 has two parts, an annular inner portion 108 
integral with the cap 110 of the cooler 100, and an annular outer portion 
112. The inner portion and the outer portions of the cooler flange 106 are 
concentric and adapted to provide an annular channel therebetween within 
which a molded energy-absorbent elastomer ring 114 is disposed. The 
channel is illustrated in the top view of the cooler provided in FIG. 5. 
The ring 114 is bonded to the inner and outer portions of the flange by a 
suitable bonding agent. The ring 114 may be a filled silicon material or 
other suitable material. The ring is disposed between an annular extension 
109 from the inner portion 108 of the flange 106 and an annular extension 
113 from the outer portion 112 of the flange 106 with a gap provided for 
the displacement thereof. The ring 114 is configured to provide equivalent 
spring compliance of K such that: 
EQU f=.omega./2.pi.=1/2.pi.(K/M).sup.1/2 &lt;&lt;1000 Hz [2] 
where f is the resonant frequency of the vibration isolated cooler, K is 
the equivalent spring compliance and M is the cooler mass excluding the 
outer portion 112 of the flange 106. In this case, frequencies above 1000 
Hertz are effectively decoupled from the detector. In general, the spring 
constant should be sufficient to support the housing and still damp 
vibration. This assures that vibration energy will not be absorbed by the 
ring and is not transmitted to the detector through the dewar flange 104. 
Thus, the novel two-part design of the flange 106 decouples vibration from 
the flange of the cooler to the flange of the dewar. 
Returning to FIG. 4, vibration is also decoupled at the end cap 124 of the 
regenerator stem. In accordance with the present teachings, the end cap 
124 has a protrusion 125 adapted to retain a screw 128. The retaining 
screw 128 is not in contact with the pad 86. The screw is noncontacting 
when inserted into the dewar and serves to allow removal of the stem and 
coupler 126 as an integral assembly. The screw 128 retains a coupling or 
coupler 126 having a coefficient of thermal expansion which is slightly 
smaller than that of the end cap 124. The isolating coupler 126 and the 
end cap may be made of metal alloys known in the art. 
At ambient temperatures, a tight fit exists between the coupling 126 and 
the end cap 124. However, as the temperature of the stem drops to 
cryogenic temperatures, the end cap contracts at a faster rate, thereby 
creating a small air gap, as shown in FIG. 4. A small air gap is created 
between the end of the end cap 124 and the coupler 126 which may be 
retained in contact with the felt pad by a thermal grease. The air gap 
should appear at a temperature between the ambient temperature and the 
cryogenic operating temperature to ensure that vibration is decoupled 
prior to the initiation of operation. 
A presently preferred embodiment of the present invention is illustrated in 
FIGS. 6 and 7 and comprises a modification of the distal end of the 
embodiment shown in FIG. 4. Again, the arrangement uses a coupler having a 
smaller coefficient of thermal expansion and a unique interfitting 
configuration of end cap and coupler that uniquely takes advantage of this 
configuration to provide a small vibration-isolating air gap at low 
temperature. 
In the embodiment of FIGS. 6 and 7, the dewar flange 104 is isolated from 
the proximal end of the cooler in the same manner as shown in FIG. 4. 
However, the vibration isolation provided at the distal end of the cooler 
is arranged to provide a very small vibration isolating gap that 
completely separates the distal end of the cooler (the cooler end cap) 
from the distal end of the dewar and therefore from the detector to be 
cooled. As shown in FIG. 6, cooler body 220 is fitted with an end cap 224 
having a distally projecting hub 225. The end cap hub 225 threadedly 
receives a retaining screw 226 having a shank 228 and a head 230. Screw 
head 230 is axially spaced from the distally facing end surface 232 of the 
end cap hub. 
The embodiment of FIGS. 6 and 7 is illustrated in FIG. 6 in a high 
temperature condition and in FIG. 7 in a low temperature condition. A 
coupler 234 is formed of a substantially right circular cylindrical wall 
236 having a continuous circumferential, radially inwardly projecting 
flange 238. Flange 238 has a radially inwardly facing surface 240 and 
opposite proximal and distal surfaces 242, 244. In both warm and cool 
temperature configurations, radial surface 240 is spaced from the shank 
228 of screw 226, proximal flange surface 242 is axially spaced from the 
end surface 232 of end cap hub 225 and distal surface 244 is axially 
spaced from the head 230 of the screw 226. Hub 225 of the end cap has a 
radially outwardly facing cylindrical surface 246 which, in the warm 
temperature condition of FIG. 6, is a tight slip fit against the inner 
cylindrical surface 248 of the cylindrical wall 236 of the coupler. An 
axially inwardly (proximally) facing surface 250 of the proximal end of 
coupler wall 236 is in contact, in the warm condition, with an axially 
outwardly (distally) facing shoulder or surface 252 of the end cap that 
surrounds the hub 225. The distal end of cylindrical wall 236 of the 
coupler is in firm contact with grease pad 256 which, in turn, is in 
contact with the distal end 260 of the dewar end. As previously described, 
the dewar end is adapted to contact the device to be cooled, which is not 
shown in FIGS. 6 and 7. 
In this embodiment, as in the embodiment of FIG. 4, the material of the end 
cap 224 and the material of the coupler 234 are selected to have different 
coefficients of thermal expansion. The coefficient of thermal expansion of 
the end cap material is greater than the coefficient of thermal expansion 
of the coupler material so that when the apparatus cools to its operating 
temperature, the end cap contracts by a greater amount than does the 
coupler. The end cap contracts both radially and axially and thereby 
creates a small but complete and continuous air gap between the end cap 
and all portions of the coupler that are in contact with the end cap at 
the warmer temperatures. The condition of the parts at the cooler 
temperatures is shown in FIG. 7, illustrating the greater thermal 
contraction of the end cap that creates the air gap. As can be seen in 
FIG. 7, at the cooler temperature, flange 238 of the coupler is still 
physically displaced radially and axially from the end cap and the screw 
230. Importantly, a small, annular radial gap 260 completely encircles the 
end cap protrusion and spaces the radially inner surface of the coupler 
from the radially outwardly facing surface of the end cap protrusion or 
hub. There also occurs a differential axial contraction that creates an 
annular axial gap 262 between the distally facing surface 252 of the end 
cap shoulder that surrounds the protrusion 225 and the proximally facing 
end surface 250 of the proximal end of the cylindrical wall 236 of the 
coupler. The two gaps 260, 262 are continuous and interconnected. Each of 
the gaps 260, 262 is very small, even at the lowest operating temperature. 
Preferably, these gaps are between 0.1 and 0.5 mils. In a preferred 
arrangement, these gaps at the coolest operating temperature are about 0.2 
to 0.3 mils. By insuring that the gaps are no greater than about 0.5 mil, 
there is still provided a good thermal path through the air confined 
within the dewar body 122 and filling the very narrow air gaps. This air 
path, which is not larger than 0.5 mil, is sufficient to provide good 
transfer of heat from the coupler to the cooler end cap at the cool 
temperatures. Note that at both warm and cool temperatures, the distal end 
of the coupler remains in contact with the grease pad 256 which itself 
remains in contact with the distal end of the dewar. The grease pad 
freezes at the low operating temperature to maintain this contact. 
The air gaps 260, 262, although small enough to retain a good heat transfer 
path between the coupler and the end gap, are still sufficient to 
eliminate transmission of vibration of the end cap and other portions of 
the cooler to the coupler and thus to the device being cooled. At neither 
high nor low temperature does the screw or any part thereof contact any 
part of the coupler. The screw is always retained, as shown, in its 
threaded engagement within the end of the protrusion on the end cap. As 
previously stated, the coupler in warm temperature condition is only a 
tight slip fit on the end cap. When the cooler is removed from the dewar 
for servicing or other reasons, it is possible that the coupler, being 
only a tight slip fit on the end cap, might be dislodged and misplaced. To 
avoid this, the screw 226 is provided. Although the screw contacts solely 
the end cap protrusion 225, at no time does any part of the screw contact 
either the coupler or the grease pad, as can be seen in FIGS. 6 and 7. 
At warm temperatures, the coupler is a tight slip fit on the end cap and 
has its radial inner surface and its axial proximal end surface in good 
tight contact with the corresponding surfaces of the end cap and its hub 
and shoulder. Thus, during this warm temperature and for a portion of the 
time during which the apparatus temperature is being lowered from the warm 
condition to its operating condition, there is good direct thermal contact 
(actual physical contact) between the end cap and the coupler. At some 
temperature before the cooler reaches its full operating temperature, the 
disclosed configurations of end caps and cooler and their different 
coefficients of thermal expansion cause the air gaps 260, 262 to appear 
and to remain during the cooling operation. 
Thus, it will be seen that the arrangement of FIGS. 6 and 7 provides a 
cooler which is optimally vibration-isolated from the dewar and the device 
to be cooled. The cooler is isolated at its proximal end by employment of 
a two-part flange, as illustrated in FIG. 4, and is vibration-isolated at 
its distal end by the creation of an air gap having both radial and axial 
portions that completely isolates the coupler from the end cap. 
Use of the described configuration, in which a protrusion or hub and 
surrounding annular shoulder are formed on the end cap and provide both 
axially facing and radially facing surfaces for cooperation with axially 
facing and radially facing surfaces on the coupler, increases the total 
area of contact at warm temperatures and also increases the total area of 
air gap at low temperatures. As stated above, the contact between the end 
cap and coupler, which exists until the lowest operating temperatures are 
approached, provides a good physical and thermal contact. It also helps to 
hold the coupler on the end cap. On the other hand, at the cold 
temperature, the total length of the air gap is increased by the described 
configuration of hub and shoulder, having both radially-directed and 
axially-directed portions. Therefore, the total area of the narrow air gap 
heat transmission path is increased. Accordingly, the described 
arrangement maximizes heat transfer both at warm and low temperatures and 
yet provides adequate distal end vibration isolation. 
Thus, the present invention has been described herein with reference to a 
particular embodiment for a particular application. Those having ordinary 
skill in the art and access to the present teachings will recognize 
additional modifications, applications and embodiments within the scope 
thereof.