Cryogenic refrigeration apparatus

A technique for producing a cold environment in a refrigerant system in which input fluid from a compressor at a first temperature is introduced into an input channel of the system and is pre-cooled to a second temperature for supply to one of at least two stages of the system, and to a third temperature for supply to another stage thereof. The temperatures at such stages are reduced to fourth and fifth temperatures below the second and third temperatures, respectively. Fluid at the fourth temperature from the one stage is returned through the input channel to the compressor and fluid at the fifth temperature from the other stage is returned to the compressor through an output channel so that pre-cooling of the input fluid to the one stage occurs by regenerative cooling and counterflow cooling and pre-cooling of the input fluid to the other stage occurs primarily by counterflow cooling.

INTRODUCTION 
This invention relates generally to cryogenic refrigerant apparatus for 
providing a fluid at extremely low temperatures and, more particularly, to 
such an apparatus which uses a technique for permitting such low 
temperatures to be reached in an efficient manner at reasonable cost in an 
apparatus the size of which can be relatively small and compact. 
BACKGROUND OF THE INVENTION 
A common type of small cryogenic refrigerator in use today is one which 
makes use of the Gifford-McMahon (G-M) operating cycle. This cycle is used 
in both single and multiple-stage configurations. A basic description of 
the G-M operation is set forth in U.S. Pat. No. 3,045,436, issued on July 
24, 1962 to W.E. Gifford and H.O. McMahon. Other apparatus configurations 
using G-M principles of operation are also described, for example, in U.S. 
Pat. Nos. 3,119,237 and 3,421,331, issued on Jan. 28, 1964 and Jan. 14, 
1969 to W.E. Gifford and to J.E. Webb, respectively. 
In such systems, no heat energy is transferred from the expanding fluid 
through the performance of mechanical work external to the refrigerator. 
Thus, while a moveable displacer element is periodically moved within the 
appartus to provide for an expansion chamber, this element is not arranged 
so as to produce an external mechanical energy exchange. Rather, as would 
be well known to those in the art, the displacer moves mass and mechanical 
energy between confined fluid volumes. 
In such an approach, the confined fluid volumes on either end of the 
displacer are connected by a heat exchange passage, often called a thermal 
regenerator. The thermal regenerator undergoes the same pressure cycling 
as the confined fluid volumes. In such a configuration, the heat energy is 
normally fully stored for a half cycle in the regenerator matrix, which 
requires the regenerator matrix to have a relatively large heat capacity. 
In totally regenerative cycles, such as in the G-M approach, the pressure 
ratio is effectively limited by the gas volume in the regenerator, which 
volume must be large enough so that the low-pressure-flow pressure drop 
through the regenerator matrix is not excessive. 
Another type of refrigerator well-known to the art and similar in 
appearance to the Gifford-McMahon type, but different in operation, is one 
which uses a Solvay cycle of operation. Both the G-M and Solvay techniques 
use valved, regenerative operating cycles, but the Solvay cycle performs 
mechanical work extraction from the refrigerant fluid. Thus, the expanding 
gas at the cold end of a piston performs work on a drive mechanism 
attached to the other end of the piston. Because of this operation, a 
Solvay refrigerator requires a high pressure gradient over the piston 
seal, while the G-M approach, with no work interaction, incorporates only 
a low pressure gradient over the displacer seal. While the high pressure 
gradient seal is a significant reliability drawback, the Solvay cycle is 
normally more efficient than the G-M cycle. 
Common regenerator materials have a heat capacity that diminishes at very 
low temperatures. For this reason, the Gifford-McMahon or Solvay cycles 
are not capable of producing effective cooling at, for example, liquid 
helium temperatures, even when multiple stages are used. To reach liquid 
helium temperatures, a second thermodynamic operating cycle, such as a 
well-known Joule-Thomson operating cycle, must be used in combination with 
a Gifford-McMahon cycle, for example. The Joule-Thomson cycle of operation 
utilizes a pre-cooling counterflow heat exchanger and an expansion valve 
(commonly referred to as a Joule-Thomson valve). Since neither the G-M, 
the Solvay, nor the Joule-Thomson cycle is capable of reaching liquid 
helium temperatures independently, in order to reach liquid helium 
temperatures, it has been suggested that various appropriate combinations 
of such techniques be used. Thus, a number of G-M stages can be used to 
provide for a pre-cooling of the helium gas before it is supplied to the 
counterflow heat exchanger of the Joule-Thomson operating cycle in 
preparation for the expansion of the gas during the Joule-Thomson 
operation. Such a combined cycle configuration could be capable of 
producing cooling down to liquid helium temperatures. While such a system 
has been commercially available, it has some severe drawbacks. For 
example, mechanically combining the two configurations results in a 
relatively complex physical configuration which is difficult to 
manufacture, resulting in a system which is often prohibitively expensive 
for many, if not most, applications. Further, such systems have poor 
reliability due to clogging of the Joule-Thomson valve and to the 
difficulty in controlling the operation of such valve. Moreover, the 
optimal mean cycle pressures and pressure ratios for the two cycles are 
not compatible, so that the combination requires a specially designed 
compressor configuration, thereby further increasing the cost and 
difficulty of manufacture. 
A further refrigeration method has been described in U.S. Pat. No. 
4,862,694 issued on Sept. 3, 1989 to J.A. Crunkleton and J.L. Smith, Jr. 
The patent discloses a method for attaining refrigeration at liquid helium 
temperatures in a relatively simple and compact configuration. One 
embodiment of the technique discussed therein incorporates a counterflow 
heat exchange operation which in a preferred embodiment thereof is 
integral with the piston-cylinder structure thereof. Mechanical work is 
extracted from the refrigerant gas during the expansion process. One 
exemplary cycle of operation for a single-stage configuration can be 
described as follows. 
When the piston is in its minimum volume position, an intake valve at room 
temperature opens to allow high-pressure gas at room temperature to enter 
the gap between the piston and cylinder. While the gap is charged to full 
pressure, the intake valve remains open and the piston begins to move, 
thereby drawing more high pressure gas into the expansion space created 
below the piston. The constant high-pressure intake continues until the 
inlet valve is closed. At this time, the expansion portion of the cycle 
begins. When the piston is at the maximum expanded volume position, a cold 
exhaust valve opens and the blow-down portion of the exhaust occurs. 
Movement of the piston then decreases the expansion volume in order to 
exhaust gas at constant pressure. At the appropriate piston position, the 
exhaust valve closes and recompression begins. When the piston reaches a 
position near minimum volume, the intake valve opens and the cycle is 
repeated. 
The gas, which has been exhausted through the cold exhaust valve, enters a 
surge volume. This volume, coupled with the flow restriction in the 
low-pressure return flow path between the cylinder and outer shell, 
results in an effective resistive-capacitive circuit flow arrangement. 
Accordingly, the mass flow rate in the return flow path is more nearly 
constant during the cycle period. The gas exits the surge volume and 
enters the low-pressure return flow passage between the cylinder and outer 
shell. As the low pressure gas is travelling at a nearly constant rate 
between the cylinder and the outer shell, it is exchanging heat with gas 
flowing between the piston and cylinder. Highly efficient counterflow heat 
transfer occurs to cool the high pressure gas entering the expansion space 
in preparation for the next expansion stroke. 
Such a method of refrigeration is also described as one which can be 
performed in multiple stages. Typically, high pressure gas enters at room 
temperature and is pre-cooled as it flows through one or more upper 
expansion volume stages on its way to the coldest expansion volume stage. 
The piston is arranged to have a stepped configuration so that, as it 
moves during the intake and expansion portions of the cycle, such movement 
would create a number of expansion volumes of varying temperature. During 
the exhaust phase, gas would flow through the exhaust valves at each of 
the stages of expansion. 
While the system described in the aforesaid Crunkleton and Smith patent 
operates satisfactorily, it requires a number of "cold" valves, i.e., 
valves which operate at low temperatures, one at each operating stage. 
Such valves not only are costly, but also have lower reliability than 
valves designed for use at warmer temperatures, e.g., at or near room 
temperature. It is desirable to provide an improved technique which 
produces effective and reliable operation at extremely low temperatures 
and which has relatively low manufacturing and operating costs. 
The present invention recognizes that, while counterflow heat exchange is 
essential for attaining liquid helium temperatures at the coldest 
expansion stage, it is not required for the warmer stages. At temperatures 
above about 20.degree. K., for example, the heat capacity of the heat 
exchanger materials is large compared to the net enthalpy flux of the 
helium through the heat exchanger over a half cycle so that the 
regenerative heat exchange operation can be efficient above about 
20.degree. K. but is much less efficient below such temperature. 
The refrigeration method of this invention combines the simplicity and 
efficiency of regenerative heat exchange for the warmer stages of a 
multi-stage cooling device with highly efficient counterflow heat exchange 
at the colder stage or stages. In addition, the warmer expansion stages no 
longer require individual cold exhaust valves at each expansion stage, 
thereby increasing reliability of the system and lowering its cost. 
BRIEF SUMMARY OF THE INVENTION 
The invention is a multi-stage refigeration device, having at least two 
and, preferably, more than two operating stages. The coldest stage 
operates at temperatures where the heat capacity of the heat exchanger 
materials of the device is small compared with the enthalpy flux of the 
helium. 
In accordance with an exemplary two-stage embodiment of the invention, for 
example, displacement or expansion volumes at each stage are periodically 
recompressed to a high pressure by reducing the displacement volume in 
each stage to substantially zero or near zero volume. By opening an inlet 
valve at the warm (e.g., at or near room temperature) end of an input 
channel, and by increasing the displacement volumes, further fluid under 
pressure, as supplied from an external compressor, is caused to flow into 
the input channel at a first relatively warm temperature (e.g., at or near 
room temperature). The fluid that has been introduced into the input 
channel is pre-cooled by regenerative and counterflow cooling as it flows 
through the input channel to the first stage displacement or expansion 
volume at which region it has been pre-cooled to a second temperature 
below the first temperature. A further portion of the incoming fluid and 
residual fluid from the previous cycle continues to flow past the first 
expansion volume and continues to flow in the input channel to the second 
stage displacement or expansion volume at the cold end of the channel. 
This latter fluid portion is further pre-cooled primarily by counterflow 
cooling as well as by some regenerative cooling as it flows in the input 
channel to the second expansion volume at a third temperature below the 
second temperature. 
The displacement volume at the first stage, i.e., a "warm" stage, is 
increased, i.e., expanded, so that the compressed fluid therein is 
expanded from the high pressure at which it had been pressurized to a 
substantially lower pressure so as to reduce the temperature of the fluid 
in or near the "warm" displacement volume to a fourth temperature which is 
substantially lower than the second temperature, but generally higher than 
the third temperature. 
The displacement volume at the second stage, i.e., the "cold" stage, is 
increased simultaneously with that of the first stage to form an expanded 
volume at the second stage so that the compressed fluid therein is 
expanded from the high pressure at which it had been pressurized to a 
substantially lower pressure so as to reduce the temperature of the fluid 
in or near the "cold" displacement volume to a fifth temperature which is 
substantially lower than the third temperature. 
At the end of the expansion stroke (maximum volume), the warm exhaust valve 
and/or the cold exhaust valve open(s), which will result in blowdown if a 
pressure difference exists over the valve(s) before opening. Although both 
exhaust valves are opened during some period of blowdown and 
constant-pressure exhaust, the valves are not necessarily opened or closed 
at the same timing. 
The displacement volume at the warm stage is decreased and the low pressure 
expanded fluid therein is caused to flow back into the input channel from 
the first stage displacement volume, toward the inlet end of the input 
channel and thence outwardly therefrom through a "warm" output valve 
thereat, a portion thereof also flowing to the cold stage. 
Further, the very low temperature, low pressure, expanded fluid which is 
used to produce the cold environment at the second stage is caused to flow 
from the "cold" displacement volume, as a result of the decrease in such 
displacement volume, into an output channel via a "cold" valve and a surge 
volume thereat, a portion thereof also flowing through the input channel 
to the warm stage. The very low temperature expanded fluid, which may be 
two phase, for example, is used to produce a cold environment for a heat 
load applied thereto, heat being transferred from the environmental heat 
load to the expanded fluid thereby boiling the two-phase fluid and/or 
warming the gaseous fluid and cooling the environment. A further heat load 
may be applied to the warm stage for cooling thereof also. 
The fluid, which is caused to flow over a first time duration from the 
"warm" first stage displacement volume at the fourth temperature towards 
the inlet end of the input channel and through the warm output valve 
thereat, is in intimate contact with the warmer surfaces of the piston and 
cylinder used in the device for changing the displacement volumes and 
exchanges heat with these warmer surfaces thereby warming the fluid 
exiting from the warm output valve and cooling the piston and cylinder in 
preparation for the following cycle. This type of heat exchange is 
commonly referred to as regenerative heat exchange. Simultaneously with 
such operation, but over a second longer time duration, the expanded low 
temperature, low pressure fluid from the "cold" displacement volume is 
caused to flow in the output channel at a substantially constant flow rate 
and at a substantially constant pressure to a fluid exhaust exit at the 
warm output end of the output channel. During operation, direct 
counterflow heat exchange is provided between the input and output 
channels to produce a pre-cooling of incoming fluid in the input channel 
and a warming of the fluid in the outlet channel to a temperature at or 
near the first temperature, less allowance of a heat exchange temperature 
difference prior to its exit therefrom. The warm exiting fluid from both 
the input and output channels is compressed, as by being supplied to an 
external compressor system, so as to supply fluid under pressure from the 
compressor system for the next operating cycle. 
Residual portions of the expanded fluid which resulted from the expanded 
operation of a previous cycle remain in the displacement volumes and in 
the input channel. Such remaining fluid may undergo recompression if the 
warm and cold exhaust valves are closed before minimum displacement 
volumes are reached. The device is now ready to execute the next expansion 
cycle. The compressed fluid from the compressor system is next supplied 
via the input channel to the first and second stage displacement volumes. 
The fluid flowing to the first stage displacement volume is pre-cooled by 
regenerative heat exchange with the piston and cylinder structures, and by 
counterflow cooling by the cold fluid flowing in the output channel. The 
fluid flowing to the second stage displacement volume is primarily 
pre-cooled by counterflow heat exchange with the cold fluid flowing in the 
output channel, although there may be some, but much less, pre-cooling due 
to regenerative cooling. 
The overall compression, intake, expansion, and exhaust process is then 
repeated, the fluid in the displacement volumes and in the input channel 
being again periodically compressed and the expansion thereof occurring as 
before. 
Such an approach permits an efficient heat exchange over a relatively wide 
temperature range to be implemented in a relatively compact manner, i.e., 
in a relatively small scale device. As such a device is scaled down in 
size, the amount of surface area available for heat exchange per unit 
volume becomes comparable with the area required for efficient heat 
exchange so that, even for reasonably small and compact scale 
configurations, the overall system readily provides the necessary heat 
transfers to produce efficient operation. There being good thermal 
connections between the input and output channels, the fluid flowing to 
the cold stage enjoys the benefits of efficient counterflow heat exchange. 
The warmer stage, where the heat capacity of the structural materials of 
which the warm stage is constructed is large compared to the convective 
heat flux of the fluid, enjoys the benefits of both regenerative and 
counterflow heat exchange. 
The size of the heat load (i.e., the applied heat load or parasitic heat 
leaks) at either stage has a relatively large impact on the type of heat 
exchange operation at the warm stage. If the heat load at the cold stage 
is much smaller than that at the warm stage, regenerative heat exchange 
dominates at the warm stage. If the heat load at the cold stage is 
relatively larger than that at the warm stage, counterflow cooling may 
account for most of the heat exchange at the warm stage. This is because a 
relatively larger heat load on the cold stage requires more mass flow to 
the cold stage. This larger mass flow rate returns to the compressor 
primarily through the output passage, which results in more counterflow 
heat exchange on the warm stage. 
In a system of the invention which uses more than two stages, in the warmer 
stages, i.e., those generally at about 20.degree. K. and above, heat 
transfer occurs between the fluid and structural material (a regenerative 
heat exchange operation), as well as between fluid flowing in the separate 
input and output cooler channels (counterflow operation). Fluid flowing in 
the output channel originates only from the colder stages having a 
connection (e.g., a valve) between the input and output channels. Thus, 
the technique of the invention is able to achieve the high 
cold-temperature efficiencies of the refrigeration method described in the 
Crunkleton and Smith patent but also benefits further from the inherent 
simplicity of warmer refrigeration techniques of the type used in 
Gifford-McMahon or the Solvay operations.

The system 10, shown in FIG. 1, utilizes a conventional compressor system 
11 and represents a particular embodiment of the invention having a 
three-stage refrigeration configuration requiring only a single cold 
exhaust valve 12 at the coldest operating stage 15. FIG. 1A depicts a 
typical pressure-volume (P-V) plot for explaining the operation of the 
system of FIG. 1. The upper two stages 13 and 14 use both regenerative 
pre-cooling by the piston-to-cylinder gap regenerators, i.e., the walls of 
piston 21 and cylinder 22, and counterflow pre-cooling due to flow of cold 
fluid from the coldest stage 15. A portion of the fluid in the upper two 
stages enters and also leaves the displacement volumes 16 and 17 thereof 
via the same flow passage or input channel 18. A "warm" exhaust valve 19 
is needed at or near room temperature to exhaust low-pressure fluid from 
displacement volumes 16 and 17 via input channel 18. A "warm" inlet valve 
25 at or near room temperature allows high pressure gas to enter input 
channel 18, when open, for the pressurization and intake portions of the 
operation, as discussed below with reference to FIG. 1A. 
Fluid flows to the cold displacement volume 20 in stage 15 which uses 
primarily counterflow heat exchange, as described below, to overcome the 
diminishing specific heat of the heat exchanger walls which provides the 
regenerative cooling in the warmer stages. The fluid to be expanded in the 
coldest stage 15 receives its initial pre-cooling in the upper two stages. 
Fluid flows to displacement volume 20 during intake and expansion. Fluid 
leaves displacement volume 20 primarily through "cold" exhaust valve 12 
when it is opened and also through channel portion 18B of channel 18 
during recompression or when warm exhaust valve 19 is open and cold 
exhaust valve 12 is closed. 
In the two upper stages, following expansion, the low-pressure return fluid 
flowing upwardly to valve 19 via input channel 18 formed between the wall 
of piston 21 and the cylinder wall 22 cools the piston wall and such 
cylinder wall so that when high pressure fluid subsequently enters input 
channel 18, it is then primarily pre-cooled by such structures in a 
regenerative cooling heat exchange operation. Such fluid is also 
pre-cooled by the very cold return fluid counterflowing in output channel 
24 from the coldest stage 15. As discussed in the aforesaid Crunkleton and 
Smith patent, channel 24 may utilize a helical spacer element 24A to 
separate its outer wall 23 and its inner wall 22 (i.e., the outer wall of 
channel 18). Both regenerative and counterflow heat exchange occurs in the 
channel between the piston and cylinder walls at the upper two stages 13 
and 14. Since the specific heat capacity of such heat exchanger walls is 
very small at very low temperatures, e.g., below about 20.degree. K., 
pre-cooling of the fluid flowing in channel 18B to the coldest stage 15 
occurs primarily due to counterflow heat exchange with the very cold 
counterflowing fluid in output channel 24. It should be noted that the 
exhaust valve 19 operates at a relatively warm temperature, e.g., at or 
near room temperature, so that the development and packaging of such a 
room-temperature valve is much less difficult and less costly than for a 
cold valve. Moreover, such warm valve can be located where it is readily 
accessible so that maintenance or service thereof is much easier than it 
would be for a cold valve, i.e. one operating substantially below room 
temperature. 
In the operation of FIG. 1, as explained with reference to the 
pressure/volume plot of FIG. 1A, fluid at high pressure and relatively 
warm temperature, e.g., at or near room temperature, is supplied from 
compressor system 11 via high pressure channel 26 to an inlet valve 25 for 
supply to input channel 18 beginning at point E. The input channel 18, 
including channel portion 18A and 18B, is pressurized to the pressure 
shown at point F by the incoming high-pressure fluid. At point F the 
piston 21 begins to move to increase the volumes of displacement volumes 
16, 17 and 20 from point F to point A. The high pressure fluid, pre-cooled 
in input channel 18, flows to upper displacement volume 16 of stage 13, to 
intermediate displacement volume 17 of stage 14, and thence to lower 
expansion volume 20 of stage 15. 
Inlet valve 25 remains open and piston 21 moves to increase the volumes of 
displacement volumes 16, 17 and 20 and high pressure fluid is supplied by 
compressor system 11 until the inlet valve 25 closes at point A of FIG. 
1A, at which point the expansion portion of the cycle begins. During the 
expansion portion of the cycle, the piston 21 is moved upwardly, and the 
volume increases or expands and the pressure drops (from point A to point 
B in FIG. 1A). 
Either or both exhaust valves 12 and 19 open at point B and an initial 
"blowdown" stage (point B to point C) occurs. Movement of piston 21 to 
reduce the volume during the subsequent exhaust portion of the cycle and 
opening of valve 12 forces low pressure, very cold fluid from displacement 
volume 20 through opened exhaust valve 12 into output channel 24 via surge 
volume 28 for flow to outlet channel 27 via interconnecting channel 27A 
(from point C to point D in FIG. 1A). Low pressure return fluid from 
volumes 16 and -7 is also forced upwardly back through input channel 18 
via channel portion -8A into channel 27 via open exhaust valve 19 and 
interconnecting channel 27B. The return low pressure fluids from channels 
27A and 27B are combined in channel 27 and supplied to a compressor system 
11. 
During the return flow of the cooled fluids from expanded displacement 
volumes 16 and 17 to valve 19, a regenerative heat exchange occurs between 
such fluids in input channel portions 18 and 18A and the warmer walls of 
piston 21 and cylinder 22. The warm exhaust valve 19 closes after a first 
time period (at some time between point B and point D) and the cold 
exhaust valve 12 closes after a second time period which may be shorter or 
longer than the first time period. Both valves 12 and 19 are closed by 
point D. Recompression of the return fluid occurs (point D to point E in 
FIG. 1A) as the piston 21 moves so as to further reduce the displacement 
volumes 16, 17 and 20. The inlet valve 25 opens after the recompression 
portion of the cycle (at point E) to permit the intake of high pressure 
fluid, e.g., at or near room temperature, from compressor system 11 into 
input channel 18, thereby further increasing the pressure (from point E to 
point F), the volume remaining substantially the same. 
As the incoming high pressure fluid flows into and through channel portions 
18 and 18A, the cooled walls of piston 21 and cylinder 22 pre-cool the 
flowing fluid by a regenerative cooling process in stages 13 and 14 so 
that the fluid reaches volumes 16 and 17 at temperatures progressively 
lower than room temperature. The low pressure cold fluid present in output 
channel 24 produces further heat exchange with, i.e., a counterflow 
cooling of, the high pressure fluid which flows through channel 18 and 18A 
to volumes 16 and 17. 
The remaining high pressure fluid which flows through input channel 
portions 18B to volume 20 is further pre-cooled substantially entirely by 
counterflow cooling due to the low pressure, very cold return fluid 
counterflowing in output channel 24. Thus, the high pressure fluid 
temperatures at volumes 16, 17 and 20 are progressively cooler due to the 
regenerative and counterflow pre-cooling in stages 13 and 14 an due 
primarily to the counterflow pre-cooling in stage 15. 
The piston moves to increase the volume (from point F to point A) during 
which time period more high pressure fluid mass is supplied in volumes 16, 
17 and 20. At point A the expansion cycle is ready to be repeated in the 
manner discussed above. 
Another configuration of the invention using a pressure-balanced displacer 
30, rather than a reciprocating work absorbing and drive mechanism as in 
FIG. 1, is shown in FIG. 2. The operation of such a system, as depicted by 
the P-V plot shown in FIG. 2A, is different from that depicted in FIG. 1A. 
Use of the pressure-balanced displacer, as would be well known to those in 
the art, eliminates the need for a work absorbing and drive mechanism and 
results in a simpler drive mechanism. For example, the displacer can be 
driven by allowing the pressure force on the displacer to become 
unbalanced at appropriate points in the cycle by using a balancing chamber 
at the mean operating pressure. In most cases, however, the drive 
mechanism for displacer motion is powered in a reciprocal manner by a 
rotary stepping motor using a suitable scotch yoke mechanism, as would be 
known to the art. The same rotary motor is used to operate the inlet and 
warm exhaust valves 25 and 19, respectively. 
In FIG. 2, the warm exhaust valve 19 and the cold exhaust valve 12 open to 
allow for depressurization of the working volumes while the displacer 
moves to decrease the volume of the working space. The amount of flow from 
the cold expansion stage 15 depends on how long the cold exhaust valve is 
open. The flow resistance from the cold expander volume 20 to the surge 
volume 28 is assumed to be considerably less than that in the 
displacer-to-cylinder gap during low-pressure exhaust. 
As seen in the P-V plot of the system of FIG. 2, as shown in FIG. 2A, a 
constant pressure intake portion of the cycle occurs from point A to point 
B, the inlet valve 25 being open and displacer 21 moving so as to increase 
the volume, the pressure remaining substantially constant. At point B the 
inlet valve 25 closes and at least one of the exhaust valves 12 or 19 
opens. An expansion (effectively a blow down expansion) portion of the 
cycle occurs from point B to point C, the other exhaust valve opening at 
some point therebetween so that by point C both exhaust valves 12 and 19 
are open. The cold fluid flows from stage 15 through output channel 24 via 
valve 12 and surge volume 28, the piston moving so as to reduce the volume 
during the exhaust portion of the cycle from point C to point D. By point 
D, both exhaust valves 12 and 19 are closed and at D the inlet valve 25 
opens. The pressurization portion of the cycle occurs from point D to 
point A as a result of the operation of compressor system 11 and the 
intake of high pressure fluid therefrom into input channel 18, while the 
cold volume remains substantially constant. 
Pre-cooling of fluid flowing in input channel portions 18 to 18A to stages 
13 and 14 occurs via a regenerative cooling process, as in the system of 
FIG. 1, together with pre-cooling occurring due to a counterflow heat 
exchange with the return cold fluid flowing in output channel 24. Further 
pre-cooling of the fluid flowing in input channel portion 18B to stage 15 
also occurs substantially by counterflow heat exchange with the return 
cold fluid, as in the system of FIG. 1, when using a pressure-balanced 
displacer as in FIG. 2. 
Valve losses occurring in the configuration of FIG. 2 can be avoided by use 
of a Stirling-type compression technique, as shown in still another 
embodiment of the invention as depicted in FIGS. 3 and 3A. The compressor 
system 11 is replaced by a compression technique which uses a power piston 
35 to compress the fluid in compressor working volume 32, channel 18 and 
displacement volumes 16, 17 and 20. Return fluid from output channel 24 
flows into volume 32 via surge volume 33 and open flapper valve 34, while 
return fluid in input channel 18 flows directly into volume 32. Power 
piston 35 and displacer 21 operate at the same speed but out of phase with 
each other. 
FIG. 3A effectively depicts the P-V plot of a cycle of operation of the 
system of FIG. 3 with respect to the overall volume represented by the 
compression working volume 32, the volumes 16, 17 and 20 and that of input 
channel 18. As seen therein, at point A, power piston 35 stops and 
displacer 21 moves to reduce the volumes 16, 17 and 20 to their lowest 
levels thereby keeping the overall volume constant and increasing the 
pressure as the fluid warms as it moves from cold to warm locations. 
During this time flapper valve 34 is closed, since the pressure in volume 
32 is greater than that in surge volume 33. Displacer 21 moves so as to 
increase the pressure (from point A to point B), although the overall 
volume remains the same during the pressurization portion of the cycle. 
Next, the power piston 35 moves so as to increase the overall volume and 
reduce the pressure, as shown by the expansion portion of the cycle (from 
point B to point C). At point C, the power piston 35 has reached its 
topmost position and the volume is at its maximum level. From point C to 
point D, the displacer 21 moves and, at the same time, during such time 
interval, the pressure in volume 32 at some displacer position becomes 
lower than that in surge volume 33 so that flapper valve 34 opens. Piston 
35 moves downwardly during the recompression portion of the cycle (from 
point D to point A). 
Operation of cold exhaust valve 12 and flapper-type valve 34 to effect flow 
may be explained as follows. Surge volumes 28 and 33 in conjunction with 
the flow resistance in output channel 24 provide an effective hydraulic 
equivalent of a resistance-capacitance (R-C) circuit arrangement which 
results in substantially constant pressure, constant flow in channel 24. 
Surge volume 28 is at a higher average pressure than surge volume 33. In a 
typical operation, for example, cold exhaust valve 12 opens at point Cl 
and exhausts cold fluid to surge volume 28 (at pressure P28) until the 
pressure in volume 20 and volume 28 are equal, at which time the cold 
exhaust valve 12 closes at point C2. At some later time, the pressure in 
surge volume 33 (pressure P33) is higher than that in volume 32 (at point 
C3), so the flapper-type valve 34 opens and fluid flows from surge volume 
33 to volume 32 until the pressures in the volumes are equal and the valve 
34 closes (at point D1). Beginning at point A, the cycle repeats, starting 
with the pressurization portion of the cycle from point A to point B. 
When power piston 35 reduces the overall volume, the fluid therein 
compresses and the low-pressure channel 24 and surge volume 33 are 
isolated from volume 32 by the closed externally controlled cold exhaust 
valve 12 and by the closed flapper valve 34. 
The configuration of FIG. 3 can be considered to be effectively equivalent 
to a Stirling-type cooler with a counterflow loop superimposed thereon in 
order to reach liquid-helium temperatures. In the configurations, 
discussed above in FIGS. 1 and 2, an aftercooler is generally needed in 
the compression system 11 to cool the compressed gas, which is normally at 
a relatively high temperature, to a temperature at or near room 
temperature, techniques for doing so in compression system 11 being well 
known to those in the art. In the configuration of FIG. 3, however, a heat 
exchanger at the warm end (e.g., a water jacket 36) can be used to remove 
energy from, and to cool, the compressed fluid at input channel 18 to room 
temperature. Although the compressed fluid (which is to be cooled) is 
separated from such water jacket heat exchanger by the low-pressure return 
fluid in the output channel 24, heat transfer from the fluid in channel 18 
via the return fluid in channel 24 to such heat exchanger can be very 
effective so as to cool the high pressure fluid to the desired room 
temperature level. 
While the embodiments discussed represent preferred embodiments of the 
invention, modification thereto and other embodiments thereof may occur to 
those in the art within the spirit and scope of the invention. Hence, the 
invention is not to be construed as limited to the specific embodiments 
disclosed herein, except as defined by the appended claims.