Split Stirling linear cryogenic cooler for a new generation of high

temperature infrared imagers

A Veprik, S Zehtzer and N Pundak

Ricor, Cryogenic and Vacuum Systems, En Harod, 18960, Israel

ABSTRACT

Split linear cryocoolers find use in a variety of infrared equipment installed in airborne, heliborne, marine and vehicular

platforms along with hand held and ground fixed applications. An upcoming generation of portable, high-definition night

vision imagers will rely on the high-temperature infrared detectors, operating at elevated temperatures, ranging from 95K

to 200K, while being able to show the performance indices comparable with these of their traditional 77K competitors.

Recent technological advances in industrial development of such high-temperature detectors initialized attempts for

developing compact split Stirling linear cryogenic coolers. Their known advantages, as compared to the rotary integral

coolers, are superior flexibility in the system packaging, constant and relatively high driving frequency, lower wideband

vibration export, unsurpassed reliability and aural stealth. Unfortunately, such off-the-shelf available linear cryogenic

coolers still cannot compete with rotary integral rivals in terms of size, weight and power consumption.

Ricor developed the smallest in the range, 1W@95K, linear split Stirling cryogenic cooler for demanding infrared

applications, where power consumption, compactness, vibration, aural noise and ownership costs are of concern.

Keywords: high temperature infrared detector, compact infrared imager, split Stirling linear cooler, vibration, aural

noise, reliability.

1. INTRODUCTION

Infrared (IR) imagers play a vital role in the modern tactics of carrying out surveillance, reconnaissance, targeting and

navigation operations. By converting the thermal battlefield into dynamic visual imagery, such equipment dramatically

enhances the observation and command control capabilities of the leaders of combat infantry and Special Forces.

In spite of the recent advances and widespread use of uncooled infrared detectors, it is still generally accepted that the

“best technology for true IR heat detection is the cooled detector” [1]. They are superior to the uncooled competitors in

terms of working ranges, resolution and ability in detecting/tracking fast moving objects in dynamic infrared scenes.

The superior performance of such imagers is achieved by using novel optronic technologies along with maintaining the

IR focal plane arrays (FPA) at cryogenic temperatures (77K, typically) using microminiature Stirling cycle cryogenic

coolers.

Over the past few years industrial progress has led to the development of a new InAlSb diode technology relying on

Antimonide Based Compound Semiconductors (ABCS), offering lower dark currents or higher operating temperatures

(being in the 100K region) [2]. The SWIR MCT technology [3] offers the possibility of operating a FPA at even higher

temperatures, in excess of 200 K. The authors of [4] report on a 320×256 middle-wavelength infrared focal plane array

based on InAs quantum-dot/InGaAs quantum-well/InAlAs barrier detector operating at temperatures of up to 200K.

The direct benefits of using such high temperature IR FPAs are the lowering of the cooling constraints resulting in a

simplified system design, using smaller and more cost effective, long life cryocoolers which typically consume less

electrical power and show faster cool-down times.

Traditionally, integral rotary cryogenic coolers were used for maintaining the, such as above, IR FPA at optimal

cryogenic temperatures. As compared to their military off-the-shelf linear competitors they were lighter, more compact

and normally had better electromechanical performance. However, their inherent drawbacks, such as high wideband

vibration export and limited lifespan spurred the development of the microminiature linear Stirling cryogenic coolers.

These novel, long-life, acoustically and dynamically quiet cryogenic coolers, while being superior in many respects,

need to be comparable to the above rotary cryocoolers in terms of bulk, power consumption and ownership costs. An

additional advantage of the linear cryocoolers, as compared with their rotary rivals, should be that, at least in theory, they

might operate at elevated driving frequencies without affecting their overall life span. Unfortunately, such coolers are

still not off-the-shelf available; the existing linear tactical coolers are too oversized/overweight and power-thirsty for the

above purposes.

The authors report on the successful development of a novel Ricor

model K527 microminiature 1W@95K split Stirling linear

cryogenic cooler [5], which was designed to provide cryogenic

cooling to a wide range of forthcoming infrared imagers, where

power consumption, compactness, vibration, aural noise and

ownership costs are of concern. The cooler, as shown in Figure 1,

comprises "moving magnet" resonant piston compressor and

pneumatically driven expander connected by a flexible transfer

line .

2. K527 EXPANDER UNIT

The expander unit comprises a "mass-spring", pneumatically

driven displacer-regenerator, as is wide accepted across the

industry. As is known, it is favorable to design such a displacer-

regenerator so that to operate it at or near its resonant frequency.

The first benefit of such a resonant tuning is that the displacer

stroke reaches its maximum at given pneumatic force delivered by the compressor, or alternatively, the desired displacer

stroke may be obtained at minimum pneumatic force delivery. The second benefit is that tuning the displacer to resonate

delivers almost automatically the desired 90º phase lag between the pressure pulses arriving through the transfer line

from the compression space and the displacer motion, as needed for producing maximum cooling effect [6]. The above

two benefits contributes essentially to increasing the displacer's expansion work at given acoustic work delivered by a

compressor. Resulting from this is improved heat pumping effect from the expansion space to which the heat payload is

thermally linked. Since the spring effect of equivalent gaso-dynamic visco-elastic loading in expander is negligibly

small, the inertia of the moving assembly and the equivalent elasticity of the above-mentioned spring define the

closeness to the resonance condition. There are two different approaches to forming the above "spring" component. In

the first approach, helical wire spring or, alternatively, two counter-facing preloaded helical wire springs are arranged as

to link the stationary housing with the movable displacer-regenerator assembly. Using mechanical spring yields compact

design and convenient resonant tuning which appears to be almost independent on environmental temperature. The

inherent disadvantages of such a design are unreliable mechanical spring connection, radiation of acoustic noise, tiny

metal debris generation and, most important, exertion of large side forces on dynamic seals leading to excessive friction,

parasitic losses, tear and wear.

In another implementation, a "pneumatic pillow" [6] may be formed by the plunger tail cyclically protruding the sealed

rear space. This approach is widely accepted across the industry because of the compactness and almost frictionless

operation. Unfortunately, the rate of such a pneumatic spring appears to be a function not only of the rear space volume

and plunger area, but also of average pressure and environmental temperature having strong influence on the above

resonant frequency. The additional disadvantage is excessive overheating of the cold finger housing because of the

irreversible compression losses associated with cyclic compression of the particular portion of gas sealed inside the rear

space, forming the above "gas pillow".

In the present design, the authors capitalize on using the highly accurate double starting helix machined spring with

integral retainers [7]. The high mechanical accuracy has been achieved through the fine machining of the spring blanks

followed by a CNC assisted electro-discharge machining the flexure slot.

(b)

Figure 2. Schematic and external layout of a pneumatically driven expander

Expansion space

Cold finger

Stepped driving plunger

Bushing Hot space Machined spring

Transfer line Rear space Housing

Figure 1. K527 split Stirling linear cryogenic cooler

(a)

Displacer-regenerator

The self-explanatory Figure 2 shows the schematics (a) and the external layout (b) of the pneumatically driven sprung

expander. In Figure 2, the displacer-regenerator comprising a stack of stainless steel mesh disks of optimized geometry

and porosity is attached to the stepped plunger and arranged so as to slide freely inside the accurately aligned bushing

and thin-walled cold finger made of stainless steel.

The above explained machine double-helix spring is placed inside the rear space. One of the spring ends is attached to

the stationary bushing and the other to the movable plunger tail, thus forming almost friction-free design.

3. K527 LINEAR COMPRESSOR (SINGLE-PISTON VS DUAL PISTON)

As is known, the single-piston compressor is a powerful source of vibration export at driving frequency and higher-order

harmonics [8]. It is widely accepted, therefore, that dual-piston, should-be dynamically counterbalanced, compressor is

the only solution of choice for the inherently vibration sensitive cooled IR applications. The major cryogenic coolers

vendors like Thales, AIM, Carleton, DRS, Raytheon, etc. are manufacturing microminiature (below 1.5W@77K) split

Stirling cryogenic coolers comprising pneumatically driven expanders actuated by dual-piston, "moving magnet",

flexural bearing compressors, see [9,10] for example. This goes, however, on expenses of size, weight, manufacturing

costs, spoiled electromechanical performances and reliability. What is even worse, since the back-to-back compressors

forming the dual-piston one are not entirely similar (different piston-cylinder gaps and friction forces, magnet force,

driving coils resistances, etc) the residual vibration export at a driving frequency and higher-order harmonics is not really

encouraging; reaching low level of vibration export may require additional vibration isolation or, sometimes, active

vibration cancellation. With this in mind, the authors rejected the dual-piston approach and adapted the single-piston

compressor design where vibration control relies on a field proven combination of high frequency vibration isolator and

tuned dynamic absorber, as explained, for example, in [11].

The single-piston K527 linear compressor is driven by a resonant “moving magnet” actuator. This allows it to obtain a

high coefficient of performance [5] and also allows the driving coil to be separated from the working agent (Helium,

typically), thus preventing the cooler interior from being contaminated by the outgasing originated from the wire

isolation varnish, eliminating the need for flying leads and potentially leaking feedthroughs. It is also important to note

that in such an approach, sinking the intrinsic copper loses from the externally located driving coil is much improved.

Figure 3 shows the schematics (a) and the external layout (b) of such a compressor. In Figure 3.a, the radially

magnetized ring is placed movably between the stationary internal and external armature yokes which are made of a

magnetically soft material. The above yokes are shaped in such a fashion so as to accommodate the driving coil carrying

the AC current and to provide for the quasi uniform distribution of the magnetic flux within the working air gap without

over-saturating the yoke's material [5].

In such an actuator, the interaction of the permanent magnetic field produced by the abovementioned radially magnetized

ring with the alternating electromagnetic field delivered by the driving coil yields axial Lorentz forces applied to the

above magnet ring and to the actuator stator formed by the static armature and driving coil.

(a) (b)

Figure 3. Schematics and external layout of a linear compressor

The above magnetic ring is bonded upon the magnet form, which is, in turn, rigidly attached to the compression piston,

arranged so as to slide freely inside the tightly matched cylinder liner being placed inside the internal yoke.

For the sake of compactness, a double starting helix machined spring, connecting the movable piston and stationary

housing, as is needed primarily for centering of the piston-magnet assembly, is placed inside the piston. Further, for the

sake of performance, compactness, manufacturing costs, ease of assembly and maintenance we have abandoned the most

fashionable “contactless” approache relying on flexural bearings and accurate (sometimes robotic) alignment and

assembly [10]. The piston and cylinder liners are tightly matched to 4µm radial clearance and made of M42 steel being

hardened to HRc 65 and machined to N3; no exotic surface treatment or plating were applied. As above, using precise

double starting helix machined spring allows for improved alignment, essential reduction of side and friction forces

along with eliminating metal debris generation. Additional advantage of this approach is that the piston-cylinder sleeves

may be matched more tightly as compared to the above contactless design; this results in a substantial decrease in blow-

by losses. This decision, to abandon the flexure bearing design, is based on the proven technology used currently in the

Ricor model K529N cryogenic cooler, where in the course of the accelerated life test the experimental cooler

accumulated in excess of 27,500 working hours (equivalent to 45,000 hours under standard test profile). When

investigating this experimental cooler’s failure, it appeared that electrical short in the driving coil happened due the

overheating of the inappropriately chosen varnish type [12]. The "post mortem" inspection revealed that the outer

diameter of the piston and inner diameter of the cylinder liners were still within the manufacturing tolerances and that the

working surfaces were free of abrasive scratches and wear.

Based on the theoretical analysis of [5], optimizing "moving magnet" linear actuator requires maintaining the ratio 2

R α ( R and α stand for the driving coil resistance and force/current constant, respectively) at the lowest possible level.

In an ideal case this means increasing the "copper volume" and magnetic flux in the air gap along with using low

resistive wire for making the driving coil. Not less important, in this regard, is to also provide for effective sinking of the

Joule heating originated from the driving coil.

However, in real life, the size and weight of the linear actuator are always subject to strict limitations. For this reason,

allocating more space for the driving coil leaves less space available for the yoke's material. This, in turn, leads to the

potential of spoiling the motor performance due to the local, or even global, yoke's material over-saturating and uneven

distribution of magnetic field in the air gap. Furthermore, using too much powerful permanent magnet and too tight air

gaps may also contribute to the harmful yoke's over-saturation. As a result, the design of such an actuator is always a

compromise between the space available for the driving coil and armature yokes, between the magnet power and the air

gap geometry. The optimal actuator should have uniform current-to-force transformation rate along the entire piston path

and the smallest 2R α ratio, subjected to limitations imposed on the actuator size. It is obvious that for the best

performance a thorough FEA modeling and optimization [5] of the linear actuator is needed. In practice, the rear-earth

NEODYMIUM-FERRUM-BORON materials having highest possible remanence, coercivity and energy product is the

best choice for the motor magnets. Along with these lines, the PERMENDUR alloys having highest permeability values

at very high magnetic flux densities appear to be the best choice for the motor magnets and yokes.

Eddy current control has been achieved by cutting the potential eddy current paths. Another important factor to consider

is keeping the needed working frequency of the compressor close to the resonant frequency; doing this may contribute

significantly to the overall system performance. The above

resonant condition depends primarily on the mass of the movable

assembly and the rate of the equivalent "gas spring" formed by

the portion of the working agent located "above" the piston face.

It goes without saying that varying the said mass is the easiest

way of bringing the system to the desired resonant condition and

improving the overall power consumption figure.

4. DIMENSIONS AND WEIGHTS

Figure 4 shows the outline dimensions of the components of

K527 cryogenic cooler. The compressor unit weight is circa

200grm, by replacing the stainless steel parts by Aluminum, some

extra 50 grm weight reduction may be handy achieved. The

expander weights circa 70 grm, replacing stainless steel rear cover

produces further 20 grm weight reduction. Figure 4. Outline dimensions

0

200

400

600

800

1000

1200

0 20 40 60 80 100 120

Ambient temperature, C

Add

ed

he

at lo

ad

, m

W

Figure 6. Maximum cooling capacity at different ambient temperatures

5. COOLER PERFORMANCES

MAPPING

Initial optimization of the cryocooler

has been performed using the Sage

Stirling cryocooler modeling software

[13]. Further fine tuning and

optimization was performed on the pilot

cryocooler. This involved varying the

rates of the compressor and expander

springs, regenerator screens geometry

and porosity, charge pressure and

driving frequency.

In the best obtained configuration, the

optimal charge pressure and the driving

frequency were found to be 15 bar and

75Hz, respectively. It is important to

note that this frequency equals precisely

the resonant frequency of a single

degree of freedom "mass-spring"

system formed by a movable assembly

"displacer-plunger". The above obtained

optimal driving frequency remained

almost optimal over a wide range of

ambient conditions and heat loadings;

therefore all of the subsequent testing

on attainable performance was carried

out using this particular frequency.

Since the cooler has been designed to

serve a wide range of payloads,

requiring different temperatures and

heat lifts, the authors mapped the cooler

performances at different cold tip

temperatures: 77K, 95K, 115K, 130K,

150K, 200K and 250K with added heat

load ranging up to 1000mW.

It is widely accepted to assess a cooler

performance by comparing the steady

state power consumption in the temperature stabilization mode at different added heat loading and environmental

temperatures.

In this particular case, the heat load inherent to the evacuated Dewar was 130mW@77K@23C. Figure 5 shows the

typical experimental outcomes.

The K527 cryogenic cooler has been optimized to work at 95K. Figure 6 shows the typical dependence of the maximum

added heat load on environmental temperature at 45W AC max. From Figure 6, this cooler might be classified as

1W@95K.

6. VIBRATION CONTROL

6.1. GENERAL CONSIDERATIONS

As is known, the reciprocation of a piston assembly inside a single piston linear compressor and displacer-regenerator

assembly inside the cold head typically produces a significant vibration export at the driving frequency and higher order

harmonics [13]. This will be demonstrated below using experimental setup comprising a 4-component Kistler type 9272

0.1

1

10

100

0

100

200

300

400

500

600

700

800

900

1000

Added heat load, mW

Pow

er

consu

mptio

n, W

AC

77K 95K 105K

115K 130K 150K

170K 200K 250K

Figure 5. Performance mapping at 23C

Figure 7 . K527 cooler under vibration export test

dynamometer, multichannel Bruel@Kjaer Nexus charge

conditioner and Data Physics QUATTRO signal analyzer. Figure

7 shows the K527 cryogenic cooler mounted upon the above

dynamometer, where compressor and cold head axis are directed

as to coincide with dynamometer axis, thus allowing independent

measurement of vibration export in compressor and expander

axis. Figures 8.a,b show the typical spectra of vibration export

produced by the compressor and cold head. From Figure 8, the

vibration export comprises well pronounced driving frequency

component 12N@75Hz (compressor) 2N@75Hz (cold head) and

along with low magnitude higher order harmonics. The radial

vibration export is negligibly small in both cases. From Figure 8,

the vibration export produced by such a compressor in a typical

tactical application has a potential to cause a harmful line-of-

sight-jitter, low frequency noise, high frequency microphonics,

etc. and, therefore, needs to be closely controlled. At the same

time, the vibration export produced by the expander is essentially

lower as compared to this of compressor and is considered to be

adequate to allow rigid mounting of such an expander to an optical bench or imager enclosure, as needed for the optical

stability, heat sinking, etc.

0.001

0.01

0.1

1

10

100

0

10

0

20

0

30

0

40

0

50

0

60

0

70

0

80

0

90

0

100

0

Frequency, Hz

Vib

ration

exp

ort

, N

(a)

0.0001

0.001

0.01

0.1

1

10

0

100

200

300

400

500

600

700

800

900

1000

Frequency, Hz

Vib

ration

exp

ort

, N

(b)

Figure 8. Vibration export produced by K527 cryogenic cooler in a typical temperature control mode.

One of the objectives of the current paper is to show the system designer that attenuation of vibration export produced by

the single-piston compressor of a split Stirling cryogenic cooler is not something of black art and is quite doable within

tight restraints imposed on size, weight and heatsinking performances.

As different from the case of counterbalanced dual-piston compressor, there is no universal recipe to designing vibration

protective and heat conductive arrangement. In each particular application case, this requires individual consideration.

The proper use of the below disclosed principles allows the system designer to do so without affecting cryocooler size,

weight and performances.

Heavy and rigid instrument.

To start with, let us consider the case when such a cryocooler is used inside heavy and rigid instrument, or, alternatively,

the instrument is mounted rigidly to a heavy and rigid platform. On this occasion, because of the large inertia and

stiffness, the vibration export will not be an issue at all. In this case it is recommended to clamp the compressor casing to

the system shield or the optical bench using thin (0.5 - 1mm) layer of thermo-conductive material. This is needed for

tolerance compensation, high frequency noise protection and efficient heat sinking.

Instrumentation operating in harsh environmental conditions.

In applications undergoing frequent exposure to harsh environmental extremes like random and sine vibration and no

vibration isolation is provided, the vibration exerted by the environment will be more intensive than this produced by the

compressor unit. The same mounting principle, as explained above is applicable.

Lightweight instrumentation operating in mild environmental conditions

These are hand-held, ground fixed portable and vibration isolated gyro-stabilized IR imagers. The optical performances

of the above listed equipment will be affected by the above mentioned massive vibration export, the suppression of

which may be achieved, even within a stringent weight and size budget, by combining the principles of vibration

isolation and tuned dynamic absorber, as explained in [14]. In this approach, the compressor is clamped to the imager

casing or the optical bench through a compliant vibration mount for attenuation of high frequency portion of vibration

export and a tuned dynamic absorber should be mounted coaxially on the compressor housing for suppressing the

vibration export at the driving frequency. The use of the tuned dynamic absorber becomes possible since the driving

frequency is essentially constant and may be adjusted and maintained with a very high accuracy, 0.01Hz, typically. The

heatsinking may rely on heat conductivity of compliant heat link/pipe or natural/ forced convection using the finned

radiator and micro-fan.

6.2. DYNAMIC MODEL AND EQUATIONS OF MOTION

In Figure 9, the 3-degree of freedom system schematically represents the relevant dynamic model. In particular,

1 2 3, ,M M M represent the masses of the system platform, compressor housing and tuned dynamic absorber, respectively.

Further, 1 1,K B represents the stiffness and damping of compliant mounting

to the stationary ground, 2 2,K B represents stiffness and damping of

vibration mounting of compressor and 3 3,K B represents the stiffness and

damping of the spring to support the dynamic absorber. The time function

( )r t represents actual vibration export, the spectrum of which is shown in

Figure 8.a. Resulting from the action of this force are dynamic deflections

1 2 3, ,x x x measured from appropriate positions of static equilibrium. This

model describes all possible cases of system and compressor mountings.

The differential equations of motion take the form

( ) ( )( ) ( ) ( ) ( ) ( )( ) ( )

1 1 1 1 1 1 2 1 2 2 1 2

2 2 2 2 1 2 2 1 3 2 3 3 2 3

3 3 3 3 2 3 3 2

0

0

M x K x B x K x x B x x

M x K x x B x x K x x B x x r t

M x K x x B x x

+ + + − + − =

+ − + − + − + − =

+ − + − =

(1)

Using complex Fourier transform, ( ) ( ) j tG j g t e dt

ωω∞

−

−∞

= ∫ , where ω is

angular frequency and 1j = − is complex unity, we make a transition

from time into a frequency domain

( ) ( ) ( ) ( ) ( ) ( ) ( ) ( )1 1 2 2 3 3; ; ; x t X j x t X j x t X j r t R jω ω ω ω⇔ ⇔ ⇔ ⇔

substitute into (1) and obtain instead of the set of differential equations the set of linear equations

1M

2M

3M

2 2,K B

1 1,K B

( )3x t

( )r t

Compressor

Platform

3 3,K B

Tuned dynamic absorber

Figure 9. Dynamic model

( )2x t

( )1x t

Ground

( ) ( )( ) ( ) ( ) ( ) ( )

( ) ( )

1 1 1 1 1 1 2 1 2 2 1 2

2 2 2 2 1 2 2 1 3 2 3 3 2 3

3 3 3 3 2 3 3 2

0

0

M X K X B X K X X B X X

M X K X X B X X K X X B X X R j

M X K X X B X X

ω

+ + + − + − =

+ − + − + − + − =

+ − + − =

(2)

The solution to (2) is trivial using, for example, the Cramer's rule. In particular, we find the complex frequency response

function (receptance) relating the dynamic response of the platform and the force applied to the compressor housing, this

is:

( ) ( )( )

( )( )( ) ( )

( ) ( ) ( )( )

2

2 2 3 3 31

1 2

1 1 2 1 2 2 2

2

2 2 2 3 2 3 2 3 3

2

3 3 3 3 3

0

0

K j B M K j BX jH j

R j M K K j B B K j B

K j B M K K j B B K j B

K j B M K j B

ω ω ωωω

ω ω ω ωω ω ω ω

ω ω ω

+ − + += =

− + + + + − +

− + − + + + + − +

− + − + +

(3)

From (3), setting 2

3 3M Kω = and

30B → , minimizes the magnitude of the above FRF at particular driving frequency

ω , i.e. the dynamic response of the base may approach zero independently on other system parameters. The above

explains the operational principle of tuned dynamic absorber, which needs to be "tuned" such as to have resonant

frequency equal exactly the driving frequency, e.g. 3 3 3

K M ωΩ = = . In practice, it is easier to adjust the driving

frequency to the above resonant frequency of tuned dynamic absorber. The complex frequency response function (3)

covers all the possible cases of systems design.

Study case 1. Portable hand held IR imager with vibration-mounted compressor

For numerical calculations in this case we accepted that the mass of the imager (excluding compressor and dynamic

absorber) is 1

3M kg= . The masses of compressor and dynamic absorber are 2

0.2M kg= and 3

0.1M kg= ,

respectively. This case will be also characterized by the low frequency primary vibration isolator mimicking the

dynamic compliance of the human upper limbs firmly holding the IR imager; these are

1 1 120 (10 )K M rad s HzπΩ = = and

1 1 1 12 0.1B Mζ = Ω = .

1.E-09

1.E-08

1.E-07

1.E-06

1.E-05

1.E-04

20

30

40

50

60

70

80

90

100

110

120

130

140

150

160

170

180

190

200

Frequency, Hz

|H1(j

ω)|

, m

/N

15Hz

50Hz

Reference

20 Hz

Hz25 Hz

Hz30 Hz

Hz35 Hz

Hz

40 Hz

Hz 45 Hz

Hz

0

100

200

300

400

500

600

700

800

15 25 35 45

Resonant frequency of vibration mount, Hz

Supre

ssio

n r

atio

(a) (b)

Figure 10. Performance of tuned dynamic absorber

Figure 11. Exploded view of tuned dynamic absorber

The dynamic absorber is assumed to be tuned to the fixed driving

frequency 75Hz and has typical low damping ratio 0.35%; namely

3 3 3150K M rad sπΩ = = and

3 3 3 32 0.0035B Mζ = Ω = . In all

the calculations we also accepted the case of moderately damped

vibration mount characterized by the damping ratio

2 2 2 22 0.15B Mζ = Ω = .

Figure 10.a shows the spectra of the modules of the system

receptances at different resonant frequencies of compressor

vibration mounting over the frequency range 20-200Hz; also

shown is the reference case (rigid compressor mounting and no

dynamic absorber). From Figure 10.a, at driving frequency 75Hz

we observe deep and wide antiresonant notches, the depth and

width of which depend strongly on the above resonant frequency,

namely, the softer compressor mounting yields better suppression

ratio (as compared with the reference case) and the frequency

tuning will be more convenient. In Figure 10.b, solid curve shows

dependence of this suppression ratio on the above resonant frequency. From Figure 10.b, using reasonably soft and

damped vibration mount allows achieving very impressive suppression ratios typically over-performing this attainable by

the actively assisted counterbalancing of dual-piston compressors.

(a) Assembled view

(b) Exploded view

Figure 12. Vibration mounting of a linear compressor.

(c) Finned radiator for forced convection

(d) Heat conductive thermal link

Figure 13. Heat sinking options

In a practical implementation, the exploded view of which is shown in Figure 11, the tuned dynamic absorber may

comprise heavy inertial ring (made of Tungsten for compactness) which is clamped between two sets of flat flexural

bearings separated by spacers for eliminating damping effects associated with dry friction between adjacent

springs. As it was already stated above, the vibration export produced by the single piston compressor is primarily axial

and negligibly small in radial direction, the single degree of freedom vibration mount of compressor to the IR imager

will be adequate.

Figure 12.a shows one of the possible implementations of such a single degree of freedom vibration mount relying on a

set of two sets of flat metal springs supporting linear compressor. Using stacked metal flat springs provides for desired

damping in compressor's vibration mount and eliminates parasitic degrees of freedom, as needed for mechanical

stability. The tuned dynamic absorber is mounted inline with compressor. Figure 12.b shows the exploded view.

Experimentation

Figure 14 shows the experimental set-up, where

vibration mounted compressor carrying 100gr

tuned dynamic absorber is mounted upon the 3kg

baseplate mimicking the inertia or the above

mentioned IR imager which in turn is placed

upon soft sponge to reproduce the case of low

frequency vibration mounting. The vibration of

the base plate was measured using miniature

Bruel@Kjaer Type 4394 accelerometer. The

curvefitting of the experimentally obtained

absolute transmissibility has shown the resonant

frequency 75Hz and a typical 0.35% damping

ratio.

Figure 15.a,b show the superimposed spectra of

the base absolute acceleration and relative

displacement for the reference case and for the case of vibration mount of compressor having resonant frequency 40Hz.

In this experimentation, the expander was mounted separately with purpose of eliminating its interference.

1.E-03

1.E-02

1.E-01

1.E+00

1.E+01

1.E+02

1.E+03

0

100

200

300

400

500

600

700

800

900

1000

Frequency, Hz

Basep

late

vib

ratio

n,

mg

Reference

40Hz

1.E-07

1.E-06

1.E-05

1.E-04

1.E-03

1.E-02

1.E-01

1.E+00

1.E+01

0 200 400 600 800 1000

Frequency, Hz

Base

pla

te v

ibra

tion,

µm

Reference

40Hz

(a) (b)

Figure 15. Platform vibration

Figure 14. Experimental setup

From Figure 15, in the reference case the working cryocooler produces a vibration export mostly at the driving

frequency. The magnitude of the primary harmonic is 367 mg; the secondary harmonic is 46-fold smaller – 8 mg. Low

powered higher frequency content is also seen up to 1000Hz. Making use of the above explained tuned dynamic absorber

produces 175-fold vibration attenuation at driving frequency which correlates well with theoretical prediction in Figure

6. Along with these lines, we observe massive vibration attenuation over the high frequency range; this may be attributed

to the action of low frequency vibration mount.

It is important to note that this significant

vibration attenuation has been obtained using a

very light dynamic absorber, the weight of

which is only 3% relative to the entire system.

Further improvement (if needed) may be easily

achieved by incorporating a slightly heavier

tuned dynamic absorber.

The experimentation was also performed for

different vibration mounting of compressor.

The obtained suppression ratios strictly follow

prediction in Figure 10.b; the experimentally

obtained points are marked as circles.

It is important to notice that after such a

massive suppression of vibration export

produced by the compressor, the cold head

appears to be the major source of vibration

disturbance. However, the vibration export

produced by the cold head results in a vibration

of the entire imager occurring primarily in a

direction which is normal to FPA. Because of

the typically essential focus depth, the imager

may tolerate this sort of mechanical disturbance without developing excessive imagery blur. In case of super sensitive IR

equipment, the authors strongly recommend using additional (smaller) tuned dynamic absorber mounted inline with the

cold head, as shown in Figure 16, where from experiment the vibration of platform plate was below 2mg@75Hz in all

directions.

Study case 2. Portable hand held IR imager with rigidly mounted compressor Sometimes restrictions imposed on the imager internal packaging, weight and heat sinking do not allow using

compressor vibration mounting. This case is typical for the lightweight systems (ultra portable handheld and

gyrostabilised imagers as used in UAVs). On this occasion, because of the small imager weight, the compressor might be

mounted rigidly and tuned dynamic absorber will suppress the base-plate vibration directly. The expression for the

magnitude ( )3H Ω at working frequency in case of rigidly mounted compressor may be easily reduced from (3), this is

( )1

2 2 2 2

3 3 1 3 34H M M ζ

− Ω ≈ Ω +

(4)

From (4) follows that for the light platforms, 2 2 2

1 3 34M M ζ<< , its vibration at antiresonance is defined primarily by the

mass of dynamic absorber and damping ratio, this is: ( )3 3 32H MζΩ ≈ ; the suppression ratio at antiresonance is,

therefore, 3 3 1

2M Mη ζ= . From the above, for the relatively light hand held imagers, the vibration suppression level is

primarily defined by the mass ratio 3 1

M M and the damping ratio 3 3 3 3

2B Mζ = Ω . This gives clear indication how

heavy the dynamic absorber should be to provide the required vibration suppression at given damping ratio. It is

important to notice that, from practice, heavier tuned dynamic absorbers normally show lower damping ratios.

7. COMPARISON WITH COMPETITORS

The self-explanatory Table 1 compares dimensions, weights, vibration export and MTTFs of the typical dual-piston

compressors of microminiature split Stirling cryogenic coolers offered by world leading vendors.

Figure 16. Using two dynamic absorbers is super-sensitive equipment

0

10

20

30

40

50

60

100

300

500

700

900

11

00

Total heat lift, mW

Pow

er

con

sum

ption, W

DC

SF070B (AIM), DC SF100A (AIM), DCSF100B (AIM), DC SL100(AIM) DCLC1040 (Carleton), DC LC1047 (Carleton), DCLC1056 (Carleton), DC LC1057 (Carleton), DCLC1062 (Carleton), DC LSF 9587 (Thales), DCLSF 9599 (Thales), DC 1W LINEAR (DRS), DC7062&196S (Raytheon), DC B602 (BEI), DCK527 (Ricor), DC UP 8497 (Thales), DC

Figure 17. Comparison of microminiature linear split Stirling

cryogenic coolers

Table 1. Comparison of microminiature linear split Stirling cryogenic coolers

In Table 1, the last three rows present, for comparison,

Ricor model K527 single-piston compressor in different

modifications. From Table 1, all the listed the dual-

piston compressors are essentially bulkier, heavier and

have shorter MTTF as compared with Ricor model

K527. The conservatively predicted MTTF in excess of

30,000 hours relies on experimental data obtained

during life testing of similar Ricor model K529

cryogenic cooler [12] which acquired 45,000

operational hours. It is important to remind the reader,

that the linear compressor of K529 cooler is driven by

the "moving coil" actuator and therefore comprises the

driving coil inside the Helium charged interior, flying

leads, feed-through and problematic heat sinking. The

K527 compressor is free of the above disadvantages and

has, therefore, a great potential for even longer MTTF.

The actual life testing is in the very beginning, the

authors will report on the outcomes.

Further, from Table 1, the vibration export produced by

vibration mounted K527 compressor is comparable with

this of its dual-piston rivals; adding 100gr tuned

dynamic absorber reduces the vibration export to typical

0.03N, which is of order of magnitude lower than this of

the listed rivals.

Figure 17 compares the power consumption of the

above coolers operating in the temperature control

(77K@23C) mode at different heat lifts, where the

marked points represent the performances of

"competitors" and bold solid line represents this of

Ricor model K527 cooler. It is important to notice that

Ricor model K527 cryogenic cooler has been designed

VENDOR / MODEL

LXOD, mm

WEIGHT, gr

VIBRATION EXPORT, N MTTF, HOURS

AIM / SF070B 115X45 1000 >20,000

AIM / SL070A 129X45 940 >8,000

AIM / SL100 122X60.5 2000 >4,000

AIM / SF100A 119X60.5 1600 >20,000

AIM / SF100B 122X57 1600 >20,000

CARLETON / LC1040 142X45 1000 1.35 >8,000

CARLETON / LC1056 120X45 1000 1.35 >8,000

CARLETON / LC1062 120X45 1000 1.35 >10,000

CARLETON / LC1055 120X46 1000 1.35 >8,000

CARLETON / LC1047 120X60 1700 2.25 >8,000

RAYTHEON / 7050 120X51 1300 2.25 >4,000

RAYTHEON / 7049/7051 203X45 1250 2.25 >2,500

RAYTHEON / 7052 120X52 1300 2.25 >5,000

RAYTHEON / 7060 120X40 450 1.35 >4,000

RAYTHEON / 7062&196S 114X33.5 400 1.35 >4,000

RAYTHEON / 7070 120X45 600 1.35 >4,000

THALES / LSF95-80/87/88/89/97/99 122X60 1500 2.5 >20,000

THALES / UP70-80/87/88/89 122X55 1500 2.25 >4,000

THALES / UP70-86/97 120X44 1150 2.25 >4,000

THALES/UP8497 125X35 450 2.25 >15,000

L-3 / B500C 400 0.9 >4,000

L- 3 / B600 1000 0.9 >4,000

L-3 / B1000 1600 2.25 >4,000

BEI / B602 102X45 800 2.25

DRS / 1W LINEAR COOLER 116X60.5 1600 2.25

RICOR / K527 63X33.5 200 11 >30,000

RICOR / K527 + 30HZ VIBRATION MOUNT 63X33.5 220 2 >30,000

RICOR / K527 + 20HZ VIBRATION MOUNT 63X33.5 221 0.9 >30,000

RICOR / K527 + 30HZ VIBRATION MOUNT +TUNED DYNAMIC ABSORBER 93X33.5 300 0.03 >30,000

to show the best performances at 95K and beyond [5], meaning that it is not really optimized to work at 77K. However,

since the above-mentioned vendors are not reporting on their coolers performances at such temperatures, the authors

were forced to make comparison at 77K. Anyway, from Figure 17, the K527 cooler over-performs the others over the

range of total heat loads of 750mW.

CONCLUSIONS

The new Ricor model K527 1W@95K cryogenic cooler comprises a single-piston compressor, driven by a resonant

"moving coil" actuator, and pneumatically driven resonant expander. The decision to abandon a fashionable dual-piston

and contact-less compressor approach allowed for drastic simplifying the mechanical design. This resulted in smaller,

lighter, more power efficient and reliable cryogenic cooler, as compared with COTS available linear rivals.

The problem of vibration control was efficiently resolved by a combined use of heat conductive vibration isolator and

tuned dynamic absorber. The attainable performance of such a combination over-performs this typical for dual-piston

compressor even operated under actively assisted vibration control.

This cryogenic cooler is ideally suited for a wide range of forthcoming high-temperature electro optical instrumentation

like portable hand-held cameras, thermal weapon sights, ground fixed and vehicle mounted surveillance cameras,

gyrostablised imagers, etc.

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