U.S. patent application number 14/895709 was filed with the patent office on 2016-05-12 for pulse tube cooler.
The applicant listed for this patent is ISIS INNOVATION LIMITED. Invention is credited to Michael William Dadd.
Application Number | 20160131399 14/895709 |
Document ID | / |
Family ID | 48875881 |
Filed Date | 2016-05-12 |
United States Patent
Application |
20160131399 |
Kind Code |
A1 |
Dadd; Michael William |
May 12, 2016 |
PULSE TUBE COOLER
Abstract
There is disclosed a cold head for a pulse tube cooler,
comprising: a regenerator having a first end connectable to a
compressor; a pulse tube having a first end and a second end; a
heat exchanger connected between a second end of the regenerator
and the first end of the pulse tube; and a phase control device
connected at the second end of the pulse tube for controlling the
flow dynamics in the pulse tube to provide cooling at the heat
exchanger, thereby maintaining a negative temperature gradient
between the first and second ends of the regenerator, wherein: the
pulse tube comprises a wall having a porous portion for allowing a
working gas to enter or leave the pulse tube directly through the
porous portion, the porous portion being nearer to being parallel
than perpendicular to the temperature gradient between the first
and second ends of the regenerator.
Inventors: |
Dadd; Michael William;
(Oxford, Oxfordshire, GB) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
ISIS INNOVATION LIMITED |
Oxford, Oxfordshire |
|
GB |
|
|
Family ID: |
48875881 |
Appl. No.: |
14/895709 |
Filed: |
June 6, 2014 |
PCT Filed: |
June 6, 2014 |
PCT NO: |
PCT/GB2014/051753 |
371 Date: |
December 3, 2015 |
Current U.S.
Class: |
62/6 |
Current CPC
Class: |
F25B 2309/1412 20130101;
F25B 2309/1406 20130101; F25B 2309/1414 20130101; F25B 9/145
20130101 |
International
Class: |
F25B 9/14 20060101
F25B009/14 |
Foreign Application Data
Date |
Code |
Application Number |
Jun 6, 2013 |
GB |
1310111.8 |
Claims
1. A cold head for a pulse tube cooler, comprising: a regenerator
having a first end connectable to a compressor; a pulse tube having
a first end and a second end; a heat exchanger connected between a
second end of the regenerator and the first end of the pulse tube;
and a phase control device connected at the second end of the pulse
tube for controlling the flow dynamics in the pulse tube to provide
cooling at the heat exchanger, thereby maintaining a negative
temperature gradient between the first and second ends of the
regenerator, wherein: the pulse tube comprises a wall having a
porous portion for allowing a working gas to enter or leave the
pulse tube directly through the porous portion, the porous portion
being nearer to being parallel than perpendicular to the
temperature gradient between the first and second ends of the
regenerator; and the heat exchanger coaxially surrounds the pulse
tube.
2. A cold head according to claim 1, wherein the porous portion is
within 5 degrees of being parallel to the temperature gradient
between the first and second ends of the regenerator.
3. A cold head according to claim 1, wherein the pulse tube is
cylindrical or has a substantially annular cross-section and the
porous portion is within 5 degrees of being parallel to the axial
direction of the pulse tube.
4. A cold head according to claim 1, wherein the porous portion is
a part of a shared wall that is shared between the pulse tube and
one or more of the following: the regenerator, the heat exchanger,
a flow distributer between the pulse tube and the phase control
device.
5. A cold head according to claim 4, wherein the regenerator
coaxially surrounds the pulse tube and the porous portion is part
of a shared wall that is a radially inner wall of the
regenerator.
6. A cold head according to claim 1, wherein the pulse tube is
substantially cylindrical and the regenerator has a substantially
annular cross-section.
7. (canceled)
8. A cold head according to claim 4, wherein the pulse tube
coaxially surrounds the regenerator and the porous portion is part
of a shared wall that is a radially inner wall of the pulse
tube.
9. A cold head according to claim 8, wherein the regenerator is
substantially cylindrical and the pulse tube has a substantially
annular cross-section.
10. A cold head according to claim 1, wherein the porous portion
extends from the second end of the regenerator along a portion of
the regenerator towards the first end of the regenerator, without
reaching the first end of the regenerator.
11. A cold head according to claim 1, wherein the porosity of the
porous portion of the wall decreases as a function of increasing
separation from the second end of the regenerator.
12. A cold head according to claim 1, wherein the pulse tube is
orientated such that the second end of the pulse tube is further
from the first end of the regenerator than the first end of the
pulse tube.
13. A cold head according to claim 1, wherein the phase control
device is configured to control the flow dynamics using one or more
fluid phase control components to control the flow of gas into and
out of the second end of the pulse tube, the fluid phase control
components being configured to operate without using any solid
moving parts.
14. A cold head according to claim 1, wherein the phase control
device provides either or both of damping and inertia.
15. A cold head according to claim 1, wherein the phase control
device comprises a piston and cylinder in fluid communication with
the second end of the pulse tube.
16. A cold head according to claim 1, wherein the porous portion of
the wall is formed from an electroformed sheet.
17. A cold head according to claim 16, wherein the electroformed
sheet is sandwiched between layers of mesh that act to even out
and/or straighten the flow of gas passing through the porous
portion of the wall.
18. A cold head according to claim 1, wherein the cold head is
configured to operate as a multi stage cooler and has a first stage
pulse tube assembly comprising: a first regenerator having a first
end connectable to a compressor; a first pulse tube having a first
end and a second end; and a first heat exchanger connected between
a second end of the first regenerator and the first end of the
first pulse tube, wherein the phase control device is connected at
the second end of the first pulse tube for controlling the flow
dynamics in the first pulse tube to provide cooling at the first
heat exchanger, thereby maintaining a negative temperature gradient
between the first and second ends of the first regenerator; the
first pulse tube comprises the wall having the porous portion for
allowing the working gas to enter or leave the first pulse tube
directly through the porous portion, the porous portion being
nearer to being parallel than perpendicular to the temperature
gradient between the first and second ends of the first
regenerator; and the cold head further comprises a second stage
pulse tube assembly comprising: an additional pulse tube directly
connected to the first pulse tube in the region of the first heat
exchanger and in such a way that there is a continuous fluidic
connection between the first pulse tube and the additional pulse
tube, the additional pulse tube being fluidically coupled to the
phase control device via said continuous fluidic connection; an
additional regenerator configured to receive working gas from the
first regenerator; and an additional heat exchanger configured to
provide cooling at a lower temperature than the cooling provided at
the first heat exchanger.
19. A pulse tube cooler comprising a cold head according to claim 1
and a compressor connected to the first end of the regenerator.
20. (canceled)
Description
[0001] The present invention relates to pulse tube coolers that
have improved efficiency. Pulse tube coolers have evolved as an
alternative to Stirling cycle coolers. Their operation is very
closely related; the main difference is that instead of using a
"solid" piston or displacer to provide an expander component, this
function is provided by a column of gas that acts as a springy,
deformable displacer.
[0002] The motion of the gas column in a pulse tube can be
controlled mechanically by a warm end piston/displacer but more
often the motion is controlled by fluid "phase control" components.
These avoid the need for extra moving components and control the
mass flow in the way that they respond to the applied pressure
variation. Various mechanisms have been employed to provide this
function. These include designs that use orifices or inertance
tubes in combination with reservoir volumes. In addition a second
inlet and orifice combination can be connected between the
compressor input and the warm end of the pulse tube to further
modify the response.
[0003] Pulse tube coolers can be used in the same wide range of
applications as Stirling cycle coolers. Examples include those for
cooling optical components both for space and terrestrial
applications. Their main advantage over Stirling cycle coolers is
the elimination or simplification of the moving expander/displacer
components. This can improve overall reliability and reduce
vibration. Potentially it also offers lower cost as a pulse tube
cold head does not require high precision.
[0004] A good general description of present pulse tube technology
is given in: Proceedings of the Institute of Refrigeration (London)
1999-2000, "Development of the Pulse Tube Refrigerator as an
Efficient and Reliable Cryocooler, Ray Radebaugh, NIST.
[0005] FIG. 1 shows the cold head 103 for a conventional Stirling
cycle cooler. The cold head 103 consists of a regenerator 104 and
an expander cylinder 105 that are connected via a heat exchanger
110 (which also acts as a fluid connection between the regenerator
and the expander cylinder) at the cold end (at temperature Tc). An
expander piston 113 is movably engaged within the expander cylinder
105. The working gas volume that includes the regenerator 104, cold
end heat exchanger 110 and the expansion space 111 above the
expander piston 113 is subjected to a pressure variation provided
by an "AC" compressor (connected via connector 107). Typically, the
compressor, or pressure wave generator, does not have valves and
usually has an ambient heat exchanger (cooler) to reject heat from
the gas before it enters the cold head 103. The compressor and
cooler are not shown in FIG. 1.
[0006] In operation, the movement of the expander piston 113 is
arranged to be out of phase with the pressure variation so that net
work is done by the gas on the piston 113. This loss of energy from
the gas causes the gas temperature to decrease which in turn allows
heat to be absorbed from the cold end heat exchanger 110. Over a
cycle the net result is that heat can be absorbed from an external
load via the heat exchanger 110 and this energy is transported to
the warm end of the cold head (at temperature Th) as work done on
the expander piston 113, which may in turn do work on gas in a
compression space 115.
[0007] In the gamma configuration shown, the expander piston 113
becomes what is termed a displacer. The displacer takes energy from
the gas at the cold end as an expander and directly transfers it
back into a gas volume at the ambient end as a compressor. The
additional gas volume is usually connected to the input from the
compressor so that the expansion work is "recycled" to improve
efficiency. This energy recovery is not essential; for low
temperature coolers the expansion power is relatively small and can
be just dissipated through a damping arrangement.
[0008] In FIG. 1 there is no indication of how the movement of the
displacer 113 is controlled in order to ensure that the work is
done by the gas in the expansion space. In practice a small
diameter shaft is usually attached to the displacer and taken
through a seal in the ambient end to a mechanism that imparts the
required motion. Various means are available for driving the
displacer and these are generally well established in the
literature.
[0009] In discussing Stirling cycle coolers and particularly pulse
tube coolers, it is useful to consider the operation in terms of
the phase relationship between the applied pressure variation and
the mass flow of the gas between the cold end heat exchanger and
the expansion space/volume.
[0010] The work done W by the expander piston is given by [0011]
W=.intg.P.dV over a cycle P is pressure, V is expander volume
[0012] For simplicity, sinusoidal* variations are assumed for
pressure and volume. This integral has a maximum when the pressure
and first derivative of volume are in phase. This is equivalent to
requiring that the pressure variation and mass flow are in phase.
(*Note: Actual pressure and mass flow variations will generally
have additional harmonics but their magnitude will be small and
they will not fundamentally change the operating mechanism).
[0013] It is noted that while this relationship will give the
highest cooling effect it does not generally give the highest
efficiency because losses in the cooler are also phase dependent.
However it is generally the case that if the pressure and mass flow
are 90 degrees out phase then there is no net work done on the gas
and hence no cooling.
[0014] FIG. 2 shows a conventional arrangement for a pulse tube
cold head 102 in which the expansion work is transmitted via a
column of gas contained in a tube 114 (which may be referred to as
a pulse tube). The gas column can be considered to consist of three
different zones. One zone 121 (which may be referred to as a first
tidal volume) behaves like the expansion space 111 of a Stirling
cooler in that gas has a tidal movement into and out of the cold
end heat exchanger 110. At the other end of the pulse tube 114 is a
similar zone of gas 125 (which may be referred to as a second tidal
volume) that has a tidal movement into and out of the ambient
temperature phase control components. This tidal volume 125 behaves
in a similar way to the compression space 115 of the Stirling cycle
cooler. In between there is a third gas zone 123 that separates the
other two. Ideally this gas volume 123 simply acts as a deformable
displacer that transmits work but at the same time provides thermal
isolation.
[0015] Cooling is generated by appropriate phasing of the volume
121 with respect to the pressure variation. Achieving this is a
more complex task than for a Stirling cycle machine and this aspect
will be described more fully below.
[0016] For efficient operation of the pulse tube 114 it is
desirable to minimise any turbulence so that the gas column remains
stratified and convection by mixing is avoided. Flow straighteners
120 may be provided for this purpose, at either or both ends of the
pulse tube 114 (as shown in the example of FIG. 2). These provide
low velocity, uniformly distributed axial flows.
[0017] In the above description of the Stirling cycle cooler it was
shown that a key requirement for cooling is a combination of
pressure and mass flow variations that are in phase. In a pulse
tube the cooling process is more complicated for two reasons:
[0018] Firstly, the deformation of the "gas displacer" (labelled
123 in the example shown in FIG. 2) with pressure requires
additional mass flows into the pulse tube. [0019] Secondly, as one
of the main attractions of a pulse tube is the elimination of
moving components, it is very desirable to control the mass flow
using a system of fluid phase control components rather than using
pistons/displacers.
[0020] In the following, the term "phase control device" is used to
refer to any device which allows the relative phases of the
pressure and mass flow variations to be such as to allow cooling to
take place. The phase control device may operate using
pistons/displacers, a system of fluid phase control components that
may not comprise any moving parts, or a combination of the two.
[0021] The mass flows required into the pulse tube can be divided
into two components: those in phase with the pressure pulse and
those out of phase. If it is assumed that the pressure variation
has a sinusoidal waveform:
P=P.sub.0 Sin(.omega.t)
(Mean pressure level is ignored; pressure is assumed to be AC
component only) then a general expression for the mass flow
variation is:
{dot over (m)}=A. Sin(.omega.t)+B. Cos(.omega.t)
The first component is in phase with the pressure pulse and this
has already been established as the "cooling" component. The second
component is mass flow required by the "deformation" of the "gas
displacer". This component does not give rise to any work input or
output; instead it acts as a spring component that stores energy
during half the cycle and releases it in the other half.
[0022] This combination of two components of mass flow is
illustrated in FIGS. 3A to 3C. The pulse tube in 3A has the total
mass flow required for operation as a cooler. This can be likened
to the combination of a displacer that is both [0023] Reversibly
expanding and contracting--the spring component [0024] Displacing
the tidal gas volumes into and out of the pulse tube The operation
of the pulse tube shown in FIG. 3A can be imagined as the
combination of the separate processes illustrated in FIGS. 3B and
3C.
[0025] FIG. 3B represents the "spring" flows that occur when the
mass flows are completely out of phase with the pressure variation
and no net work is done on the gas. In 3B1 the pressure is
increasing and all the gas flows are inward. In 3B2 the pressure is
decreasing and all the gas flows are outward. It is seen that the
"gas displacer" changes shape in accordance with the distribution
of the mass flows.
[0026] FIG. 3C shows the gas displacer considered as a solid
displacer that is causing tidal mass flows into and out of the
pulse tube by displacement without any pressure variation. In 3C1
the displacement is from the cold end at Tc (top) to the ambient
end at Th (bottom). In 3C2 the displacement is the reverse of 3C1.
The flows into and out of the pulse tube are consistent with these
displacements.
[0027] In FIG. 3B the mass flows associated with the gas spring are
shown equally divided between the two ends of the pulse tube i.e.
the mass flow from the cold end is the same as from the ambient
end. This does not need to be the case; the mass flow could be
entirely from either end or divided in any proportion. The key
characteristic of the "spring" flows is that they are out of phase
with the pressure variation. The direction from which the "spring"
mass flows enter the pulse tube does not affect the amount of
refrigeration generated, however it does affect the losses and
hence the overall efficiency of the cooler. It is generally
preferable for at least some of the "spring" mass flow to enter
from the ambient end. If it comes in from the cold end it has to do
so via the regenerator adding more thermal load to this component
and hence increasing its losses.
Example of the Orifice Pulse Tube
[0028] One of the earliest designs used in pulse tube coolers is
the orifice pulse tube, as shown in FIGS. 4A and 4B, in which gas
is allowed to pass between the ambient end of the pulse tube and a
reservoir 124 via an orifice 122. The orifice 122 is an example of
a fluid phase control component. FIG. 4A shows the pulse tube with
the orifice 122 closed. There is no mass flow at the ambient end
and the mass flow at the cold end consists of only the out of phase
spring component; there is no net cooling:
[0029] For
P=P.sub.0 Sin(.omega.t)
{dot over (m)}.sub.spr=B. Cos(.omega.t)
W=.intg.P.dV=0
[0030] If the orifice 122 is opened (FIG. 4B) then the mass flow at
the ambient end will be determined by the flow characteristics of
the orifice 122. An orifice generally behaves as a resistive flow
component where the mass flow is proportional to the pressure drop
across it. As the reservoir volume 124 remains close to the mean DC
level then the mass flow the through the orifice 122 and hence from
the ambient end of the pulse tube will be given by
{dot over (m)}.sub.orif=const.P. Sin(.omega.t)=A. Sin(.omega.t)
[0031] The mass flow into the pulse tube at the cold end is a
combination of the spring mass flow and the orifice mass flow:
{dot over (m)}.sub.coldend={dot over (m)}.sub.spr+{dot over
(m)}.sub.orifA. Sin(.omega.t)+B. Cos(.omega.t)
[0032] The cold end now has a mass flow component that is in phase
with the pressure variation and there is a net cooling effect at
the cold end.
[0033] (Note: As might be expected, the cooling is equal to the
work lost in pumping gas backwards and forwards through the orifice
122)
Other Types of Pulse Tube
[0034] The orifice pulse tube is probably the simplest design that
has succeeded in producing reasonable levels of cooling at low
temperatures. However it will be seen from FIG. 4A that the spring
mass flow enters the pulse tube entirely from the cold end. This
flow adds a considerable burden to the regenerator and limits the
ultimate performance. Two configurations that have been developed
to overcome this limitation will be briefly described.
Double Inlet Pulse Tube
[0035] In the double inlet pulse tube configuration there is an
additional connection between the ambient end of the pulse tube to
the compressor via a second orifice. The effect of this is to cause
a mass flow into the ambient end of the pulse tube that is in phase
with the pressure drop across the regenerator and cold end heat
exchanger. This can be used to allow some of the "spring" flow to
enter from the ambient end. The reduction in flow through the
regenerator reduces the regenerator loss and allows an overall
improvement in performance.
[0036] Although good performance has been achieved with this
arrangement it has been found that the second orifice is inclined
to be asymmetric in its operation. This attribute tends to generate
net DC flows that circulate around the pulse tube and regenerator.
These flows do not undergo thermal regeneration as intended and
even at low levels can produce unacceptable losses.
Inertance Tube
[0037] An alternative to the Double Inlet configuration that
attracted interest is the inertance tube configuration. In this
configuration the orifice of the Single orifice configuration
described above is replaced by a tube or assembly of tubes that
connect the ambient end of the pulse tube to a reservoir volume.
The tube or assembly of tubes is a further example of a fluid phase
control component(s).
[0038] The tube assembly is arranged to have both damping and
significant inertia. It terms of an electrical analogy the
inertance tube assembly is represented by a series combination of a
resistor and an inductor whereas an orifice is represented by just
a resistor. The inductive component has the effect of allowing some
of the "spring" mass flow to enter the pulse tube from ambient end.
This helps to reduce the thermal load on the regenerator without
setting up the circulating flows that tend to occur with the Double
Inlet design.
Coaxial Designs
[0039] In FIGS. 1 to 4 the cold head geometry is that of a U bend.
FIG. 5 shows a coaxial geometry applied to the general pulse tube
arrangement shown in FIG. 2. The operation is exactly the same
except that the regenerator is now located in an annular volume
surrounding the pulse tube. The wall of the pulse tube no longer
has to withstand the full gas pressure instead it just acts as a
partition. This arrangement is advantageous as it provides a more
compact and convenient envelope for the cold head. Similar
arrangements have also been widely used for Stirling cycle cold
heads.
Pulse Tube Losses
[0040] There are four main loss mechanisms directly associated with
the operation of a pulse tube: [0041] Heat transfer through the
bulk of the gas: The minimum value is set by thermal conduction in
stratified layers but this can be greatly increased if there is
significant mixing/turbulence. [0042] Natural Convection: The
temperature gradient produces a corresponding density gradient. The
latter tends to drive a circulation of gas within the pulse tube.
This effect causes pulse tube performance to be very dependent on
orientation--pulse tubes generally work better with their cold ends
pointing down. [0043] There are two recognised losses associated
with the interaction of the axial oscillation of the gas and the
pulse tube wall. [0044] One loss, referred to as a type of shuttle
loss, is due to the constantly reversing temperature gradient
between the gas and the wall. [0045] A second loss that has been
termed "Streaming convection" does not have any simple explanation
but appears to derive from the velocity boundary conditions imposed
at the interface between the gas and wall. The "Streaming
convection" loss is discussed in: J. R. Olsen, G. W. Swift,
Acoustic Streaming in Pulse Tube Refrigerator: Tapered Pulse Tubes,
Cryogenics, Volume 37, Issue 12, December 1997.
[0046] It is an object of the present invention to provide an
improved pulse tube cooler which at least partially address one or
more of the problems with the prior art discussed above, for
example by reducing losses.
[0047] According to an aspect of the invention, there is provided a
cold head for a pulse tube cooler, comprising: a regenerator having
a first end connectable to a compressor; a pulse tube having a
first end and a second end; a heat exchanger connected between a
second end of the regenerator and the first end of the pulse tube;
and a phase control device connected at the second end of the pulse
tube for controlling the flow dynamics in the pulse tube to provide
cooling at the heat exchanger, thereby maintaining a negative
temperature gradient between the first and second ends of the
regenerator, wherein: the pulse tube comprises a wall having a
porous portion for allowing a working gas to enter or leave the
pulse tube directly through the porous portion, the porous portion
being nearer to being parallel than perpendicular to the
temperature gradient between the first and second ends of the
regenerator.
[0048] Thus, an arrangement is provided in which a porous portion
allows gas to enter or leave the pulse tube substantially laterally
(which may also be referred to as "radially"). This provides a
number of advantages relative to the prior art.
[0049] For example, there is a greater range of possibilities for
the distribution of both the refrigeration process and the "spring"
flows required:
[0050] i) In prior art arrangements such as that shown in FIG. 5
for example, cooling only occurs at the cold end of the pulse tube.
In contrast, the porous wall of embodiments of the present
invention allows cooling to be additionally or alternatively
distributed along the length of the pulse tube. The rate of cooling
depends on the magnitude of the flow through the pores. Therefore,
the spatial distribution of cooling can be set as desired by
appropriate selection of the spatial distribution of porosity of
the walls. Distributing some or all of the cooling along the porous
wall can be advantageous as it allows the pulse tube to act as a
multistage cooler. Cooling part way along the pulse tube can be
used to remove heat at a higher temperature and hence with lower
power requirement. This approach reduces the heat flow to the cold
end and allows more of the low temperature refrigeration to be
available for cooling of the load. Efficiency can therefore be
improved.
[0051] ii) In the prior art the gas "spring" flows can only enter
from the cold end or warm end (or both). However, the actual flow
requirement is distributed along the length of the pulse tube. The
provision of a porous wall allows the flow requirement to be
fulfilled with more favourably distributed gas flows. The resulting
pulse tube can be likened to a "multiple" inlet pulse tube.
[0052] A further advantage is related to the ability to reduce gas
velocities and gas movement with respect to the thermal
gradients:
[0053] i) In a conventional pulse tube all the gas flow has to be
distributed across the end faces of the pulse tube. The flow
straighteners are designed to provide low velocity, evenly
distributed flows but there is clearly a lower limit set by the
cross sectional area of the pulse tube. The provision of porous
lateral walls allows the gas flows to be distributed over much
larger areas allowing gas velocities to be considerably reduced.
This helps to maintain streamlined flow with minimal mixing and
turbulence
[0054] ii) In a conventional pulse tube the gas flows are all axial
and this requires the gas to have significant velocities in the
direction of the thermal gradient. The porous walls of the present
invention allow much of the flow to be radial where the gas
velocity is perpendicular to the thermal gradient. This reduces the
axial gas velocities and allows the establishment of a more
favourable temperature distribution.
[0055] A further advantage is that, where there is radial flow, the
gas flow will be closer to being perpendicular to the wall than
parallel to it. The axial velocity component that interacts with
the tube wall will be much smaller than for a conventional axial
flow design. This will result in lower losses that are dependent on
the gas flow/wall interface. For example the "shuttle" loss will be
reduced as this loss is proportional to the square of the axial
displacement. The other loss related to the interaction between the
pulse tube wall and the axial gas velocity is the "Streaming
Convection" and it is also believed that the large reduction in
axial gas velocity adjacent to the tube wall will also result in
lower values for this loss.
[0056] Furthermore, there is the effect the radial gas velocity
components have on natural convection losses. Natural convection
occurs when buoyancy forces drive a circulation of gas. Typically
this occurs when a hot surface is positioned below a cold
surface--in a simple model the heated less dense fluid rises and
the cooled denser fluid descends. The magnitude of the natural
convection heat transfer is dependent on the Grashof Number--a
measure of the balance between the buoyancy forces that drive the
circulation and resistive processes that reduce it. In a
conventional axial flow pulse tube, the tube's walls are the
principal source of damping--the smaller the diameter the more the
circulation is suppressed. The axial gas flows do not generally
have much influence as they will add to the circulation on one side
and subtract on the other. In the radial flow pulse tube it will be
seen that there are two effects that will tend to reduce natural
convection: firstly, the radial flows close to the wall will
disturb the axial flow and will tend to suppress it; and, secondly,
the core of the gas displacer where circulation could become
established has a significantly reduced diameter which will also
tend to suppress circulation.
[0057] In an embodiment, a phase control device is provided that
comprises a piston configured to move within a cylinder. In
comparison with prior art Stirling cycle coolers, the piston can be
confined to a greater extent to the warm end of the tube than can
the displacer of the Stirling cycle cooler. This allows the
arrangement to be made lighter and/or easier to manufacture, while
also tending to lower vibration and/or manufacturing cost. In
comparison with prior art pulse tube coolers, this approach allows
for a more compact arrangement because there is no need for a
reservoir volume to be provided to implement the phase control.
[0058] Embodiments of the invention will now be described, by way
of example only, with reference to the accompanying drawings in
which corresponding reference symbols indicate corresponding parts,
and in which:
[0059] FIG. 1 depicts a cold head for a prior art Stirling cycle
cooler;
[0060] FIG. 2 depicts a pulse tube cold head according to the prior
art;
[0061] FIGS. 3A-C depict operation of the pulse tube shown in FIG.
2; FIG. 3A illustrates a combination of spring action and
displacement action, FIG. 3B illustrates the spring action in
isolation, FIG. 3C illustrates the displacement action in
isolation;
[0062] FIGS. 4A and 4B depict an orifice pulse tube according to
the prior art; FIG. 4A illustrates the pulse tube with the orifice
closed, FIG. 4B illustrates the pulse tube with the orifice
open;
[0063] FIG. 5 depicts a coaxial geometry version of the pulse tube
cold head of FIG. 2;
[0064] FIG. 6 depicts a pulse tube cold head according to an
embodiment of the invention, in which a coaxial geometry is
adopted;
[0065] FIGS. 7A-C show how the flow regime of the pulse tube cold
head of FIG. 6 can be seen as a combination of spring and
displacement actions in an analogous manner to the flow regime
depicted in FIGS. 3A-C; FIG. 7A depicts the combination of spring
action and displacement action, FIG. 7B depicts the spring action
in isolation, FIG. 7C depicts the displacement action in
isolation;
[0066] FIG. 8 depicts a variation of the embodiment shown in FIG. 6
in which the radial flows have been extended to the cold end heat
exchanger and the connection to the phase control device;
[0067] FIG. 9 depicts an embodiment in which a warm end displacer
piston is used to control the mass flows in place of fluid
components;
[0068] FIGS. 10A and 10B depict a heat exchanger configured to pass
gas into the pulse tube in a radial direction; FIG. 10A is a side
sectional view, FIG. 10B is an end view;
[0069] FIG. 11 depicts an embodiment in which the pulse tube wall
contains an electroformed screen that is solid at the warm end and
then becomes more permeable towards the cold end;
[0070] FIG. 12 depicts an embodiment in which the pulse tube and
the regenerator are cylindrical and axially displaced relative to
each other;
[0071] FIG. 13 depicts an embodiment in which the pulse tube has an
annular cross-section and coaxially surrounds a cylindrical
regenerator;
[0072] FIG. 14 depicts a cold head for use in a two stage
cooler.
[0073] In an embodiment, a cold head for a pulse tube cooler is
provided in which the flow of at least a portion of the working gas
between the pulse tube and other components is diverted through a
porous portion of a wall that is nearer to being parallel with the
temperature gradient of the regenerator than being perpendicular to
it. Preferably, the porous portion is within 5 degrees of being
parallel, preferably within 1 degree of being parallel. The result
of this arrangement is that the working gas enters or leaves the
pulse tube in a predominantly lateral direction (perpendicular to
the temperature gradient). As discussed above, this allows
operational losses to be reduced. An example of such an arrangement
is shown in FIG. 6.
[0074] In this particular embodiment, the cold head 2 comprises a
regenerator 4 which has a first end 6 and a second end 8. In use a
temperature gradient will be maintained along the regenerator 4
such that the second end 8 will be colder (at Tc) than the first
end 6 (at Th). The second end 8 may therefore be referred to as the
cold end and the first end 6 may be referred to as the warm end. In
an embodiment, the warm end 6 is at ambient temperature in which
case it may be referred to as the ambient end.
[0075] The regenerator 4 is connectable via passageway 7 to a
compressor (not shown in FIG. 6) at the warm end 6. As described
above with reference to FIGS. 1 and 2 the compressor may be
configured for example to provide an AC pressure wave input, for
example in the form of a sine wave (or a form in which the dominant
component is a sine wave). The cold end 8 is connected to a heat
exchanger 10 which in turn is connected to a first end 12 of a
pulse tube 14, optionally via flow straightener 20.
[0076] The second end 16 of the pulse tube 14 is connected to a
phase control device via connection 26, optionally via flow
straightener 20. As described above with reference to FIG. 2, a
phase control device in this context refers to any device which
allows the flow dynamics (e.g. the relative phases of the pressure
and mass flow variations) to be such as to allow cooling to take
place. The phase control device may operate using a
piston/displacer, a system of one or more fluid phase control
components without solid moving parts, or a combination of the
two.
[0077] In the embodiment shown, the pulse tube 14 has a cylindrical
form and comprises a wall 15 defining a pulse tube volume. In other
embodiments, the pulse tube may take other forms. In the case where
the pulse tube is cylindrical it will be understood that references
to the axial direction refer to the axis of the cylinder. In the
case where the pulse tube has an annular cross-section, the axial
direction will refer to the axis of the cylinder formed by the
inner or outer surfaces of the annular cross-section. In
embodiments where the pulse tube is not cylindrical and does not
have an annular cross-section it will be understood that references
to the axial direction refer to an axis of elongation or "long
axis" of the pulse tube. Typically, the axis of the pulse tube 14
will be substantially parallel to the direction of the temperature
gradient in the regenerator (i.e. the direction of steepest
temperature gradient--the vertical direction in the orientation of
the figures).
[0078] In the particular embodiment shown the regenerator 4
coaxially surrounds the pulse tube 14 (as in the embodiment of FIG.
5). However, this is not essential. In other embodiments, the
regenerator may surround the pulse tube but not have a common axis,
may only partly surround the pulse, coaxially or not, or simply be
positioned adjacent to the pulse tube.
[0079] In an embodiment, the pulse tube shares a wall with one or
both of the regenerator 4 and the heat exchanger 10 and/or, where
provided, a flow distributer 11 between the second end of the pulse
tube 14 and the phase control device (see FIG. 8 for example). Any
combination of these shared walls may comprise porous portions for
allowing lateral passage of working gas into or out of the pulse
tube 14 directly through the porous portion. In the embodiment
shown in FIG. 6, the pulse tube 14 shares a wall 16 with the
regenerator 4 only. In the example shown the shared wall is a
radially inner wall of the regenerator 4.
[0080] FIG. 8 shows an example embodiment in which the pulse tube
shares a wall with the regenerator 4, the heat exchanger 10 and a
flow distributer 11.
[0081] In an embodiment, at least a portion (the uppermost region
in the marked range 28 in the example shown) of the shared wall 15
is porous to a working gas such that a portion of the working gas
can enter or leave the pulse tube 14 (indicated by arrows 30)
through the porous portion of the wall 15. Allowing some of the gas
to enter into the pulse tube 14 directly from the regenerator in
this way does not change the distinction between "spring" type mass
flows that are out of phase with the pressure variation and
"displacement" flows that are in phase, it just alters the way in
which the gas displacer deforms and the distribution of the cooling
effect. In the example shown, the additional radial inward flows in
the range 28 effectively change the shape of the first tidal volume
(the equivalent of the volume 121 of FIG. 5 that does work on the
gas displacer 123) so that it comprises both a component 121A in
the form of a solid cylinder (as in FIG. 5) and a component 121B
having an annular form located adjacent to the wall 15.
[0082] FIGS. 7A to 7C show how the new flow regime can be
considered as a combination of "spring" flows and "displacement"
flows analogous to the purely axial flow arrangement illustrated in
FIG. 3.
[0083] FIG. 7B represents the "spring" flows that occur when the
mass flows are completely out of phase with the pressure variation
and no net work is done on the gas. In 7B1 the pressure is
increasing and all the gas flows are inward. In 7B2 the pressure is
decreasing and all the gas flows are outward. It is seen that the
"gas displacer" 23 changes shape in accordance with the
distribution of the mass flows.
[0084] FIG. 7C represents the mass flows when the gas displacer 23
is just displacing gas without any pressure variation. In 7C1 the
displacement is from the cold end (at Tc) and the radial inlets
(the pores in the wall 15), which are at intermediate temperatures,
to the ambient end at Th. In 7C2 the displacement is the reverse of
7C1. The flows into and out of the pulse tube 14 are consistent
with these displacements. Although the gas displacer 23 does change
shape it is important to note that it does not change volume.
[0085] FIG. 7A is intended to represent the combination of 7B and
7C when the pulse tube 14 is operating as a cooler with both in
phase and out of phase flows occurring with respect to the pressure
variation.
[0086] In an embodiment, the flows through the porous portion of
the wall are such as to cause minimal mixing or turbulence. The
ideal is that the gas flows follow streamlines so that over a cycle
they tend to return to the point where they entered the pulse tube
14. This (or close to it) is achieved by having low gas velocities
and well distributed apertures/pores.
[0087] In an embodiment, the porous portion of the wall 15 extends
from the second end 8 (the cold end) of the regenerator 4 along a
portion, but not all, of the regenerator 4 towards the first end 6
(the warm or ambient end) of the regenerator 4. This is the case
for example in the arrangement of FIGS. 6 and 7A-C. In this and
other embodiments the porosity of the porous portion of the wall 15
may be configured to decrease as a function of increasing
separation from the second end of the regenerator, or vary in some
other manner.
[0088] In the arrangement of FIGS. 6 and 7A-C the working gas flows
into the pulse tube 14 through a combination of radial flows
(through the porous wall) from the regenerator and axial flows from
the heat exchanger 10. However, other combinations of flow,
optionally of radial and axial flows or of flows directed in other
ways, are possible. FIG. 8 depicts an example in which the heat
exchanger 10 is arranged to coaxially surround the pulse tube 14 at
the cold end. A wall having a porous portion is provided (e.g. a
mesh--see FIGS. 10A and 10B discussed below) to guide the flow from
the heat exchanger 10 into the pulse tube in a radial
direction.
[0089] In the example of FIG. 8, a flow distributer 11 is provided
for distributing flow from the second end of the pulse tube 14 to
the phase control device. The flow distributer 11 may have a heat
transfer function in addition to a flow distributer function. The
flow distributer 11 may therefore optionally have the same or a
similar configuration to the heat exchanger 10. The arrangement
described below with reference to FIGS. 10A and 10B may be applied,
for example, to the flow distributer 11.
[0090] In the embodiment shown, the flow distributer 11 comprises a
wall having a porous portion that allows for a radial flow between
the pulse tube 14 and the phase control device (via the flow
distributer 11). In the arrangement of FIG. 8, therefore, flows
into and out of the pulse tube 14 are predominantly (or entirely)
radial and the expansion and compression spaces 121,125 (first and
second tidal volumes) are radially outside of the gas displacer
123.
[0091] All the pulse tube arrangements described above have assumed
the use of a phase control device comprising a fluid phase control
component, which avoids additional moving components. For example,
a combination of inertance tube and gas reservoir may be used.
[0092] FIG. 9 shows an example of a complete pulse tube cooler,
including a compressor 45 and a phase control device 47, in which
the phase control device comprises a piston 32 (which may be
referred to as a warm end displacer) configured to move within a
cylinder 34. A warm end tidal volume region is no longer required.
The pressure variation within the cold head causes a net force to
act on the warm end displacer 32 that is equal to the product of
the pressure and piston face area. The dynamics of the piston
assembly are arranged so that there is both inertia and damping.
The resulting motion gives the desired mass flows at the boundary
between the cold end heat exchanger 10 and the pulse tube 14. The
damping is provided in this example by a damper piston 40, mounted
via a shaft located by suspension spring 46, and a flow restrictor
42. The damping can be adjusted by varying the flow restrictor
42.
[0093] This use of a warm end displacer 32 can be regarded as a
gamma configuration Stirling cycle cooler in which the thermal
insulation function of a conventional displacer is transferred to a
deformable "gas displacer"--the "gas displacer" effectively becomes
an insulating piston crown. One advantage of this arrangement over
a conventional Stirling cycle cooler is that the displacer 32 is
confined to the warm end. This allows it to be both lighter and
easier to make with lower vibration and lower manufacturing
costs.
[0094] At the cold end the operation of the pulse tube 14 shown in
FIG. 9 is the same as the embodiment illustrated in FIG. 8. Some of
the gas flow enters the pulse tube 14 via a radial connection with
the cold end heat exchanger 10. The remaining gas flow is
distributed along a section of the pulse tube wall 16 and enters
via connecting apertures between the regenerator 4 and the pulse
tube 14.
[0095] The warm end displacer arrangement will not be the first
choice for all pulse tube applications because of the additional
moving components. However it possibly more compact as there is no
longer a requirement for a reservoir gas volume which can occupy a
significant volume.
[0096] In the example shown, the compressor 45 (which may also be
referred to as a linear pressure wave generator) is configured to
impart a modulating pressure on the cold head 2 via connector 7 (in
this case a length of connecting tube). The compressor 45 comprises
a piston cylinder assembly driven by a linear motor 50. The moving
components are mounted on flexures (suspension springs) 46 so as to
allow a close but non-touching fit between the compression piston
52 and cylinder 53. This arrangement is generally referred to as a
"clearance seal"--there is leakage but it is small enough to be
acceptable. The compressor 45 does not have valves and the
modulating pressure can be regarded as analogous to a voltage that
has both AC and DC components. The basic operating frequency may
typically be in the range 50 to 100 Hz.
[0097] The configuration of the phase control device of the example
of FIG. 9 is independent of compressor configuration and could be
replaced by one or more fluid phase control components without any
modification to the compressor.
[0098] Implementation of Radial Flows
[0099] For the radial flows into the pulse tube 14 there are number
of approaches that can be used. FIGS. 10A and 10B show one possible
arrangement for the case where the flows enter from the heat
exchanger 10 (arrows 61). In this embodiment, the axial gas flow 63
from the regenerator 4 enters slots 60 in the heat exchanger 10
that direct the flow radially into the pulse tube. A layer of mesh
64 between the slots 60 and the pulse tube 14 is used to even out
and/or straighten the flow distribution. As mentioned above, a
similar arrangement can be used for flows between a pulse tube and
a flow distributer (e.g. flow distributer 11 of FIG. 8).
[0100] For radial flows between the regenerator 4 and the pulse
tube 14 the porosity of the pulse tube walls 16 needs to be
controlled so as to give the required flow distribution. As the
pressure drop across the pulse tube wall 15 varies along the
regenerator 4 it is expected that porosity will also need to vary
significantly. The gas flows within the regenerator 4 and porous
wall 16 are expected to be laminar and this allows them to be
modelled using simplified methods--e.g. software intended for
mathematically similar processes.
[0101] A finite element thermal conduction model was used to
estimate the range of permeability that would be needed to achieve
the required gas flows. For a pulse tube 14 where the radial flow
was comparable with the axial flow as per the arrangement shown in
FIG. 6, the range of local permeability varied by a factor of
100.
[0102] If the volume flow dV/dt through the regenerator 4 is given
by:
V t = A f .kappa. R .mu. P x ##EQU00001##
[0103] where A.sub.f is flow area, .kappa..sub.r is permeability of
regenerator 4, .mu. is viscosity of gas and dP/dx is pressure
gradient, the permeability of the pulse tube wall 15 would need to
vary in the range 0.01.kappa..sub.r to .kappa..sub.r. The
regenerator mesh used in pulse tube coolers is usually very fine. A
typical mesh specification is:
TABLE-US-00001 Mesh Number (No. Of Wires per inch): #400 Wire
diameter: 0.030 mm Aperture dims: ~0.034 .times. 0.034 Porosity:
0.59.
The permeability of porous material is generally: [0104]
Proportional to flow area and hence porosity [0105] Inversely
proportional to hydraulic diameter/aperture size.
[0106] As it is required to reduce the permeability by a factor of
100 from material that already has only 34 micron pore size it will
be seen that it is necessary to produce a tube wall 16 with
apertures of 1 micron or larger. A variation of 1 to 34 microns
does not give the required permeability range so it is also
necessary to alter the effective porosity.
[0107] Although there are a number of technologies that can be used
to produce apertures in sheet material, many are not suited to the
small dimensions required. One approach that is suited is the
technology of electroforming. For example an electroformed screen
can be made to give a varying permeability by controlling both the
size and density of the apertures (the density effectively
determines the flow area/porosity). The thickness of the screen
needs to be between 5 and 10 microns to allow apertures of .about.1
micron to be defined. The screen is sandwiched between two layers
of fine mesh which can then be formed into a tube and installed
between the regenerator and pulse tube volume. The mesh is used for
two reasons: [0108] To produce a more robust assembly that can be
handled [0109] To give a symmetric flow characteristic so as to
avoid any tendency to produce net circulations between the pulse
tube and the regenerator as these will tend to generate losses. The
range of permeability that can be defined with a single
electroformed screen is large but if necessary more than one screen
can be used in conjunction with additional layers of mesh to
further reduce the permeability.
[0110] FIG. 11 shows an example where the pulse tube wall 15
comprises an electroformed screen that is solid at the warm end
(portion 66) and then becomes more permeable towards the cold end
(portion 64) so as to provide a controlled radial flow into pulse
tube 14.
[0111] The embodiments described above relate to single stage pulse
tubes where the regenerator 4 has an annular form and the pulse
tube 14 is located concentrically within the regenerator 4.
However, these features are not essential. Multi-stage pulse tubes
may be provided. Additionally or alternatively, the regenerator and
pulse tubes may take different forms. Some example configurations
are described below with reference to FIGS. 12-14.
[0112] FIG. 12 shows an embodiment in which the pulse tube 14 is
provided outside of the regenerator 4. In the particular example
shown, both the pulse tube 14 and the regenerator have a
cylindrical form. In other embodiments, either or both may have
different shapes. The pulse tube 14 is axially displaced relative
to the regenerator, extending in an opposite direction relative to
the regenerator than arrangements such as those shown in FIGS. 6
and 8. The second end 16 of the pulse tube 14 is further from the
first end 6 of the regenerator 4 than the first end 12 of the pulse
tube 14 rather than the other way round (as in FIGS. 6 and 8). In
the particular example shown the regenerator 4 is coaxial to the
pulse tube 14 but this is not essential. Providing the regenerator
4 outside and/or axially displaced relative to the pulse tube 14
removes constraints on the relative sizes and shapes of the
regenerator 4 and pulse tube 14. For example, the regenerator 4 is
not constrained to have a hollow (e.g. annular) form. Furthermore,
the lengths of the regenerator 4 and pulse tube 14 are not
constrained to be equal or nearly equal. This removal of
constraints provides greater flexibility for optimising the
properties of the cold head 2.
[0113] The embodiment of FIG. 12 is an example of an embodiment in
which the connection 26 to the phase control device is made at the
opposite end of the cold head 2 (i.e. the warm end) compared with
configurations discussed previously, for example with reference to
FIGS. 6 and 8. Providing the connection 26 at the opposite end
facilitates high performance.
[0114] In an embodiment, the pulse tube 14 comprises one or more
flow shaping features 70. The flow shaping features 70 are
configured to deflect flow towards the axial direction and/or make
unavailable to the flow volumes of the pulse tube 14 in which the
flow rate would otherwise be relatively low. Such volumes in which
the flow rate would be relatively low are sometimes referred to as
"dead volumes". The existence of dead volumes reduces efficiency
and can cause undesirable stagnant and/or swirling flow patterns.
The flow shaping features 70 improve efficiency by reducing dead
volumes and helping to provide a smooth transition (arrows 80)
between radial flow into the pulse tube (for example through meshes
64) and the predominantly axial flow (arrows 82) that exists in the
bulk of the pulse tube. The embodiment of FIG. 12 is an example of
an embodiment comprising two of the flow shaping features 70: one
at either end of the pulse tube 14.
[0115] In an embodiment, the pulse tube 14 is provided radially
outside, optionally coaxially surrounding, the regenerator 4. In an
example of such an embodiment the regenerator 4 has a cylindrical
form and the pulse tube 14 has an annular form surrounding the
regenerator 4. The porous wall through which the working gas can
enter or leave the pulse tube 14 may in this embodiment comprise a
part of a shared wall (e.g. shared between the pulse tube 14 and
the regenerator 4) that is a radially inner wall of the pulse tube
14. An example of such an embodiment is depicted in FIG. 13. In
this embodiment, the working gas flows radially inwards from the
pulse tube 14 into the regenerator 4. In the example shown, the
flow passes through a mesh 64 (an example of a porous wall)
connected between the regenerator 4 and the heat exchanger 10. It
is expected that arrangements with the pulse tube 14 radially
outside of the regenerator 4 will be more efficient than
arrangements in which the pulse tube 14 is radially inside of the
regenerator (for example as shown in FIGS. 6 and 8) for larger
sizes of cold head 2.
[0116] In an embodiment, the cold head 2 is configured to operate
as a multi stage cooler, with cooling at different temperatures
being provided at different heat exchangers. The multi stage cooler
may comprise a two stage cooler. FIG. 14 depicts an example of such
a two stage cooler.
[0117] The multi stage cooler comprises at least one additional
regenerator 72 (one in the case of a two stage cooler), at least
one additional pulse tube 74 (one in the case of a two stage
cooler), and at least one additional heat exchanger 76 (one in the
case of a two stage cooler). The original pulse tube 14,
regenerator 4 and heat exchanger 10 may be referred to as a first
stage pulse tube assembly. Each set of additional elements may be
referred to as a second (or third, fourth etc.) stage pulse tube
assembly. Thus, in the example of FIG. 14, the second stage pulse
tube assembly comprises one additional regenerator 72, one
additional pulse tube 74 and one additional heat exchanger 76.
[0118] The second stage pulse tube assembly is attached to the cold
end of the first stage pulse tube assembly (e.g. to the heat
exchanger 10 of the first stage pulse tube assembly). The cold head
2 is configured so that a portion of the working gas from the first
stage pulse tube assembly is directed, for example via appropriate
passages 86, into the additional regenerator 72 of the second stage
pulse tube assembly. The additional pulse tube 74 of the second
stage pulse tube assembly is connected directly (i.e. has a
continuous fluidic connection) to the pulse tube 14 of the first
stage pulse tube assembly and shares the same connection 26 to the
phase control device (via the pulse tube 14 of the first stage
pulse tube assembly). It is noted that the phasing may not be
simultaneously ideal for both stages. However, this is a workable
arrangement as the performance of the first stage pulse tube
assembly is not overly phase sensitive and the phase can therefore
be adjusted to be close to ideal, or ideal, for the second stage
pulse tube assembly. The portion of the flow that enters the
additional regenerator 72 is cooled and input to the additional
pulse tube 74 at the cold end of the additional pulse tube (the top
of the additional pulse tube 74 in FIG. 14). The first stage pulse
tube assembly works between temperatures Th and Tc1, providing
cooling at the heat exchanger 10. The second stage pulse tube
assembly works between temperatures Tc1 and Tc2 and provides
cooling at the heat exchanger 76. Th is warmer than Tc1 and Tc1 is
warmer than Tc2. The arrangement can provide cooling more
efficiently over the temperature range Th to Tc2 than a single
stage cooler working directly between Th and Tc2. In the embodiment
shown, three flow shaping features are provided: one at the hot end
of the pulse tube 14, one at the cold end of the pulse tube 14, and
one at the cold end of the additional pulse tube 74. The use of
radial flows into and out of the pulse tubes in such multi stage
coolers, according to embodiments of the invention such as that
shown in FIG. 14, allows multi stage coolers to operate
particularly efficiently in comparison with otherwise equivalent
arrangements that do not use radial flows.
* * * * *