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= 18 bar. The maximum attainable peak to peak pressure variation at the cooler inlet is Ap = 7.8 bar for a flexible tube length of L = 1 m between rotary valve and cold head. Under these operating conditions, with radiation shielding of the cold head and without the gradiometer installed, the minimum no-load temperature of the cooler is Tmin = 41 K and the net cooling power is 0 = 2.8 W at 80 K. The gradiometer was found to introduce an additional heat load of,~ 0.2 W (see Figure 4 in Reference [1]). Figure 2 displays the variation of Tm~n and 0 ( 8 0 K) with the length L of the flexible tubing. These measurements were performed without gradiometer and with a lower peak to peak pressure variation of Ap = 7 bar for L = 1 m. Ap decreases roughly by 0.18 bar/m between L = 1 and 10 m. As seen from Figure 2, at a tube length of 10 m a no-load temperature of 70 K and a cooling power of 0(80K) = 0.4 W are achieved, which are still sufficient for cooling of YBCO SQUIDs. For the present gradiometer, as shown below, lengths of 2 to 4 m are enough for a suppression of the noise from the rotary valve. YBCO RF-SQUID Gradiometer Test For measurements of the magnetic flux noise the two SQUIDs, mounted at the cold head, were operated in flux-locked loop in an unshielded environment. The typical flux noise spectrum below 25 Hz is depicted in Figure 3. The measurements were performed in electronic gradiometer configuration at three different distances between the rotary valve and the cold head. At a distance of 1 m (Fig. 3a)) the spectrum exhibits
ICEC16/ICMC Proceedings 75
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Figure 2 Net cooling power at 80 K and minimum no-load temperature of the coaxial PTR as a function of the flexible tube length between rotary valve and cooler inlets. 10-2
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Figure 3. Flux noise spectrum of a microwave YBCO RF-SQUID electronic gradiometer cooled by the coaxial PTR. The distance between the rotary valve and the cold head is a) 1 m, b) 1.5 m, and c) 2 m. Cold head temperature: T = 52 K. interference peaks located at 1 and 2 Hz which can be related to the rotation of the valve shall that occurs with a frequency of 1 Hz. With increasing distance, these interference signals decrease rapidly and at 2 m they are comparable in magnitude to the white background noise of the order of 6 x 10.5 (I)0/~/Hz (see Fig. 3c)). With a field to flux transfer function of about 10.7 T/~0 of the used SQUIDs, this white flux noise corresponds to a field gradient noise of a few pT/x/(Hzcm). These results show that the rotary valve is the main source of interference in our system. However, even at a distance of 4 m no further reduction of the interference signals was obtained, indicating a remaining noise source at the cold head. Three possible sources of pressure-wave induced noise signals were considered: a) periodic variations of the cold head temperature, b) periodic variations of the diamagnetic moment of the helium gas, and c) elastic length oscillations of the regenerator tube to which the SQUID holder is attached. Estimates of the effects of sources a) and b) show that they are of minor importance in the present set-up. The elastic length oscillation of the regenerator tube is calculated to be + 6 lam, which
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can give rise to a periodic noise signal in a local magnetic field gradient. As a source of such a field gradient, stainless steel components and welded joints in the vicinity of the SQUIDs could be identified. Further Improvements of the Cooler On the basis of the above results the cooler was reconstructed using only proven non-magnetic materials. In particular, the following changes to the design in Figure 1 were made: pulse tube fabricated from Teflon (~18.3 • 2.4 • 251 mm3), regenerator tube made of titanium alloy (qb31.8 x 1.55 • 205 mm 3) and filled with mesh 247 bronze screens, vacuum vessel completely made of acrylic glass. Figure 4 shows the flux noise spectrum of a single SQUID magnetometer channel of the gradiometer obtained with the improved version of the cooler. In order to detect only magnetic noise from the cold parts of the cooler itself, these measurements were performed in a two-layer mumetal shield. It can be clearly seen from Figure 4 that after the removal of the stainless steel components the residual interference signals from the pressure wave are no longer visible, even with a single SQUID operating as magnetometer. 10-3 N T ""e-O ~) 1 0 -4 . ~
0
z
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i,
10-5
,
0
I
5
,
I
i
10
Frequency
I
15
i
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20
i
25
(Hz)
Figure 4 Flux noise spectrum of a microwave YBCO RF-SQUID magnetometer operated with the improved version of the coaxial PTR. The data were taken in a mumetal shield, f = 1.9 Hz; cold hcad temperature: 60 K; distance to rotary valve: 2m; L = 4 m .
In conclusion, the present results demonstrate that a properly designed PTR can be used for low-noise cooling of sensitive high-T~ SQUIDs.
ACKNOWLEDGEMENTS The authors thank Y. Zhang and A.I. Braginski (KFA Jfilich) for supplying the microwave SQUIDs. Financial support by a BMBF grant (No. 13N6176 0) is gratefully acknowledged.
REFERENCES 1 2 3 4.
Thummes, G., Landgraf, R., Giebeler, F., MOck, M. and Heiden, C., Pulse tube refrigerator for highTr SQUID operation Adv. Cryog. Eng. (1996) 41, to appear Ishizaki, Y. and Ishizaki, E., Prototype of pulse tube refrigerator for practical use Adv.Cryog. Eng. (1994) 39 B 1433 Thummes, G., Giebeler, F. and Heiden, C., Effect of pressure wave form on pulse tube refrigerator performance In: Cryocoolers 8, Ross, R.G. Jr., ed., Plenum Press, New York (1995) 383. MOck, M., Progress in RF-SQUIDs IEEE Trans. Appl. Supercond. (1993) AS-3 2003
Analysis and Investigation of Large Diameter Pulse Tube Refrigerator
Youyi Guo*, Xiaoxin Wang**, Shaoxiong Bian**, Yanzhong Li ~$ *School of Chemical Engh~eering, Xi'an Jiaotong University, Xi'an 710049, China **School of Energy and Power Engineering, Xi'an Jiaotong University, Xi'an 710049, China
In this paper the application possibility of the pulse tube refrigerator under large mass flow rate is studied. A type of cryocooler named large diameter pulse tube refrigerator is suggested.With gas expansion analysis and control volume analysis, theoretical analysis has been carried on and verified by experiments. The results show that some parameters such as working frequency and inlet gas pressure have significant influence on the large diameter pulse tube working conditions.
INTRODUCTION The main purpose of tiffs paper is to solve the problem that small diameter pulse tube (common) generates less refrigeration power. At present, the diameter of pulse tube ranges from 3ram to 9ram , and its principles of refrigeration are fully.discussed. Now we hwent a type of cryocooler named large diameter pulse tube refi,igerator[ 1] whose diameter is 20mm. It can supply more refrigeration power. Similar to the small diameter one, the large diameter pulse tube has the same inlet-outlet types such as shift type (basic type) and reversible type, and the same hot-end improved models as reservoir-orifice structure and double inlet structure. The analysis methods of common pulse tube are manifold. So it is possible to discuss the mechanism of large diameter pulse tube refrigerator on the basis of the theory of small diameter one. Therefore the aim of this research is to deliberate the pulse tube refrigerator's principle all-roundly.
|
electric ~i lnotor []
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||
ilII
l[i]i
pulse tube
I~g e"~at~
I;
:~ rotathlgvalve Ill I[Ioutlet
double-inlet hot-endheat~ valve exchf~
-lllll~
cold-end
(~ reservoir
~ orifice
1
Figure 1 Structure of large diameter pulse tube refrigerator
DESCRIPTION Fig. 1 shows briefly the construction of the large diameter pulse tube, which has a cluster structure[2]. The ill'St part of the retiigerator is the inlet-outlet shift system, which is composed of rotating valve, electric motor and speed changing system. Flanges are used.here to join the cold-end to pulse tube so as lo disassemble the refrigerator easily and dhninish the void volume of cold-end. Double inlet pipe located between gas-divided-part and hot-end is a red copper pipe about 1 meter long. The diameter of this pulse tube is 2 -- 7 times as others in literature oll hand. Speed changing system consists of a worm reduction device with speed ratio !:20 and several pairs of belt pulleys. Therefore, the pulse tube can obtain a few sorts of working fiequencies by using the worm and belt pulleys or using the worm only. Because double inlet valve and orifice valve can be disassembled, selecting different valve plate leads to variation of passing 287
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hole. A valve is composed oI three sections valve body, valve seat and valve plate with a hole. The choice of plate depends on the real need. The working gas is air, which is easily gained. The thermodynamic parameters of air are similar Io that of nitrogen that is a kind of common cryogenic working gas. Parameters to be measured are gas pressure, temperature, mass flow rate and refrigeration power et al. For a specific parameter to be measured, we install a corresponding measuring system on the sample machine. ANALYSIS Gas Expansion Analysis Fig.2 shows that the working gas hi the pulse tube may be divided into three sections: the gas bundle I , which flows hlto the hot-end from the hot-end heat exchanger, discharges at the hot-end after one period; the gas brindle I11, which flows hRo the cold-end from regenerator, discharges at the cold-end after one period; the gas brindle II, which oscillates fi'om the cold-end to the hot-end, is kept in the pulse tube permanently. According to the energy equation of the opening system, we can obtahl the energy equation,s satisfied by the gas bundle I , II and III in one period on the pulse tube's steady working conditions[3] (l/'r)~pdv, + H, = 0
(I)
(1/'~)~pdv~ = 0
(2)
H n +(1/"c )} pdv.~ = 0
(3)
H, in equation (3) is refrigeration power the pulse tube can produce. Under the ideal conditions, the cold-end enthalpy H, is equal to the hot-end enthalpy H.. So, if increasing refrigeration power is desired, increasing the work ~pdvm produced by the gas brindle III ill every period is needed. The second term {}f equation (3) can be rewritten as }pd(V.l./x). The more the gas volume flow rate that flows through the cold-end is, the more the work produced by the gas bundle III is, and the more refrigeration power is. This is one reason why the large diameter pulse tube refiigerator is put forward.
,
3
2
0. orifice
1
0q 41
1. hot-end heat exchanger 2. pulse tube 3. cold-end 4. reservoir
Figure 2 Gas sections h~ pulse tube
dQ Mout
CV
m
hi. j /
hOB|
dW
Figure 3 regard pulse tube as control volume
Control Volume kalalvsis From Fig.3, the pulse tube is regarded as a control volume. The following energy equation is obtained from the analysis on tile opening system[4] dQ dU dW ~+m~,,h., = ~+~+lh dt dt dt
....,h ....,
(4)
since tile wall of pulse tube is rigid, dW = 0. The continuity equation is m
lilin - -
mout
--
d~--n-n= d(pV / RT) = m( d___pp_ dT dt dt pdt Tdt )
(5)
m
where T is average temperature along tile pulse tube. The pulse tube refrigerator usually works at steady workhlg conditions, therefore T is regarded as a constant. So the following equation is obtained m~n - Ih,,,,, = m( dp / pdt)
(6)
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Since working frequency is high, the processes in the pulse tube are assumed adiabatic, dQ / dt = (), thus equation (4) changes into 0
dU + m~,.dt
= ~
(ho,. - h ~,, )
m dp rh.,p dt
ho,,,
(7 )
and dU dt
=
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d(mT) dt
~
=
c,, d(pV) R dt
--
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(8)
Then the following equation is obtahaed by substituting equation (8) into (7) and simplifying hou,-hm=
V ( k T 2 - 1 ) dp k--~l T dm~,----~
(9)
To enlarge the volume of pulse tube is an effective method of producing more refrigeration power. The reasons can be analyzed from two aspects: one is to increase the length of pulse tube, another is lo enlarge diameter of that. It is known that volume of pulse tube V in equation (9) is proportional to the length, but proportional to square of the diameter of pulse tube. In order to obtain more refrigeration power, it is more effective to enlarge diameter of pulse tube. This is another reason why the large diameter pulse tube refrigerator is put forward.
EXPERIMENT Sample lnachhle of large diameter pulse tube is applied here for the sake of matching for the air compressor in our laboratory. It is shift double inlet type. The experimental set is composed of followhlg main parts ...........large discharge ah" compressor, rotating valve, gas-divided-part, regenerator, large diameter pulse tube, double inlet pipe, orifice and reservoir. Some structural parameters of the large diameter pulse tube system are as followings: the diameter and the length of ~ 290 pulse tube are 20mm and 400mm respectively, the diameter and ~ 280 the length of regenerator are 57mm and 200mm respectively, ~ 270 the volume of reservoir is 1.2L. The ch'cumstance conditions 260 are: ambient temperature is 18 ~ atmospheric pressure is I atm. 9~ 250 The surface of mahl cold areas such as cold part of regenerator, cold-end and cold pall of pulse tube is not to be insulated. -~ 240 ~"~ "8 Thermocouple thermometers are used here to measure o 230 . . . . . . . . . 12345678910 temperature. Refrigeration power is measured by electric power 0 frequency (Hz) simulation measurement. The prhlciple is to wrap electric heating wires around surface of cold-end and supply electric Figure 4 Effect of working frequency current, when temperature of cold-end keeps a fixed value, the on cooling performance reading in dynamometer that is joined in circuit is equal to refrigeration power h~ this specific temperature. Experimental measurement is in synchronism with workh~g processes. The diagram of the cold-end temperature of the pulse tube to the working frequency is given in Fig.4. The results shown in the figure indicate that the refrigeration temperatures in very big or very small frequency regions are higher than that in middle frequency region. That is to say, working frequency has an optimum value to refrigeration power. In the experhnental conditions this value is about 4.5Hz. The curves Of cold-end temperature to refrigeration power a r e shown in Fig.5, from which the better effect of higher inlet pressure is illuminated. Ill a cold-end temperature of 260K, the refrigeration power of hflet pressure 0.51MPa is as about 5times big as that ofhdet pressure 0.31MPa. In the conditions of bared cold-end, this refrigeration power is ve~3' considerable. It is proved that large diameter do produce large refrigeration power.
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Fig.0 shows the tendency or cold-end temperature to the diameter of double inlet valve hole. When double inlet valve closed (that means the diameter of double inlet hole is 0mm), the pulse tube refrigerator turns into the basic one, the branch of working gas to be unable to generate cold power will still flow through regenerator, bring about decreashlg of efficiency of regenerator and rising of refrigeration temperature. Double inlet rate has an optimum value to the large diameter pulse tube refrigerator obviously. In experimental conditions, this value is 1.5mm. 280
250
z f
~- 270
~ 260 inlet pressure / E 250 0.3 IMPa 240
/
245 11) t-
= 240
~ 235 inlet pressure 0.51MPa
O
ca 230
,
0
~
,
,
~
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5 10 15 20 25 30 35 refrigeration power (W)
Figure 5 Curves of cold-end temperature to refrigeration power
~ 230 = 225 -6 ~ 220 o 0
0.5 1 1.5 2 2.5 double inlet hole (mm)
Figure 6 Effect of double inlet on cold-end temperature
CONCLUSIONS The authors suggest the large diameter pulse tube refrigerator, which can produce more refrigeration power. Its diameter is as 2 - 7 thnes large as small diameter pulse tube's. Theoretical analysis of the large diameter pulse tube shows that the effect of enlarging diameter on thermodynamics performance is remarkable. The results of analysis h~dicate that there are two important reasons for suggesting large diameter pulse tube refrigerator. The shift type pulse tube with rotation valve (sample machine) works better at the lower rotational speed because inlet process is influenced by shift time. The experimental data show that large diameter pulse tube refrigerator do generate more refrigeration power. The optimutn values of some parameters, such as workhlg fiequency, double inlet rate at specific experimental conditions, are given hi this paper. NOMENCLATURE cv Specific heat at constant volume, J / kg. K H - - Enthalpy, J h - - Specific enthalpy, J / k g m - Mass, kg m -- Mass flow rate, k g / s p -- Gas pressure, Pa Q - - Quantity of heat, J R Gas constant, J / kg. K
T Thermodynamics temperature, K t -- Time, s U - - Internal energy, J V - Volume, 1/13 v Specific volume, m 3 / kg W -- Work, J Time, s
REFERENCES 1 Wang, X.X., Principle analysis and investigation of the large diameter pulse tube, Master thetis, Xi'an Jiaotong University (1993) (In Chhlese) 2 Wang,X.X., Bian S.X.,Xu B., Experiment study of the effect of fi'equency et al on the performance of shift type pulse tube refrigerator Chinese Clyogenic Engineering (1995) 1 30-35 3 Zhu, S.W., Thermodynamic analysis and important improvement of pulse tube ......double inlet pulse tube, Ph.D. thesis, Xi'an Jiaotong University (1990) 4 Zhu, S.W., Wu,P.Y., Chen,Z.Q., A single stage double inlet pulse tube refrigerator capable of reachhlg 42K C1Togenics (1990) 30 257-26 i
Active-Buffer Pulse-Tube Refrigerator
Shao Wei Zhu, Yasuhiro Kakimi, Koji Fujioka, and Yoichi Matsubara* Tsukuba Laboratory, DAIDO HOXAN Inc., 3-16-2, Ninomiya, Tsukuba, Ibaraki 305, Japan * Atomic Energy Research Institute, Nihon University, 7-24-1, Narashinodai, Funabashi, Chiba 274, Japan
This paper introduces a highly efficient GM-type active-buffer pulse-tube refrigerator. Buffers are connected to the hot end of the pulse-tube through on/off valves. Gas flow between the pulse-tube and the buffers is controlled by opening and closing these valves. A prototype single stage active-buffer pulsetube refrigerator has been made and tested. Cooling power of 160 W at 80 K was achieved with input power of 3.65 kW. The percent Carnot was 11% at 80K. The lowest temperature achieved was 28 K.
INTRODUCTION GM-type pulse-tube refrigerators have great potential due to their simplicity and high reliability. There are four types of large loss in pulse-tube refrigerators which decrease their efficiency. In double inlet pulse-tube refrigerators and orifice pulse-tube refrigerators, there is a large pressure difference across the high pressure valve just after it is opened, because the pressure in the pulse-tube is near the low pressure of the compressor. When the low pressure valve is opened, there is also a large pressure difference across the valve, because the pressure in the pulse-tube is near the high pressure of compressor. This problem can not be solved by enlarging the opening area of the valves. It results from the void volumes of the regenerator and the pulse-tube. The loss is called intrinsic loss. Orifice pulsetube refrigerators have a different type of loss: large regenerator loss due to the large mass flow rate which does not give expansion work. Double inlet pulse-tube refrigerators solve this problem. However, they also have a different type of loss: work flow loss through the double inlet. While four-valve pulsetube refrigerators avoid the above three types of loss, large compressor gas loss is generated by the valves at the hot end of the pulse-tube. To solve these four problems and thereby increase efficiency, a GM-type active-buffer pulse-tube refrigerator is introduced. Buffers are connected at the hot end of the pulse-tube through on/off valves. The function of the buffers is different from that of orifice or double inlet pulse-tube refrigerators. The gas flow between the pulse-tube and the buffers is controlled by opening and closing the valves. We call it an active-buffer pulse-tube refrigerator, and it avoids the above four types of loss. To verify its performance, a prototype single stage pulse-tube refrigerator was made and tested. Cooling power of 160 W at 80 K was achieved with an input power of 3.65 kW. The percent Carnot was 11% at 80K. The lowest temperature achieved was 28 K.
STRUCTURE AND MECHANISM Figure 1 shows the schematic structure of an active-buffer pulse-tube refrigerator with three buffers. It 291
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consists of a regenerator, a heater, a pulse-tube, a cooler, valves VH , VL, V~,V2, V~, buffer 1, buffer 2, and buffer 3. Valve VH is a high pressure valve connected to the high pressure output of the compressor. Valve Vc is a low pressure valve connected to the low pressure input of the compressor. Unlike a tube expander [1], the active-buffer pulse-tube refrigerator has a regenerator. The basic working process of gas in the pulse-tube and buffers is similar. Two buffers or more than three buffers are also available as an active-buffer system. Displacement, expansion, and compression of gas in the pulse-tube refrigerator is controlled not only by the compressor, valve VH, and valve VL, but also by buffer 1, buffer 2, buffer 3 and valves /'],V2,//3. The gas is moved between buffer 1, buffer 2, buffer 3, and the hot end of the pulse-tube through valves V~, V2 and V~, respectively, by the opening and closing of the valves. The opening and closing process is as follows: (1) VL and Vl close. V2 opens. (2) V2 closes. V3 opens. (3) Vtt opens. (4) VH and V~ close. V2 opens. (5) V2 closes. Vl opens. (6) VL opens.
V3 Buff
9
_
Heater Pulse tube
Compressor ~ ~ VL Regenerator
-~H I
r 3]
~ Y ~ ~ ' 1 !! Buffer 2] ~2~
Cooler/Jf~ ["~~Buffer V1
1]
Figure 1 Active-buffer pulse-tube refrigerator /]1, Pz, ~ , t~, and Pt,, the pressures in buffer 1, buffer 2, buffer 3, the high pressure output of the compressor, and the low pressure output of the compressor, respectively, have the following relations: is slightly lower than 1~,~, t~ is slightly higher than t~. t~ is moderate. In process (1), gas flows into the hot end of the pulse-tube from buffer 2. Then the pressure in pulse-tube increases to near t~. In process (2), gas flows into the pulse-tube from buffer 3. Then the pressure in the pulse-tube increases to near t' In process (3), the gas in the compressor flows into the pulse-tube through the regenerator. The gas in the pulse-tube from buffer 3 by process (2) flows back to buffer 3. In process (4), the gas in the pulse-tube flows into buffer 2. Then, the pressure in the pulsetube decreases to near t~. In process (5), the gas in the pulse-tube flows into buffer 1. Then the pressure in pulse-tube decreases to near 1]. In process (6), the gas in the pulse-tube flows out to the compressor. The gas in buffer 1 from the pulse-tube by process (5) flows back to the pulse-tube. During the process of opening the high pressure valve or the low pressure valve, the pressure difference across the high pressure valve or the low pressure valve is small. Thus the intrinsic loss caused by the pressure difference is small. When the pressure increases in the pulse-tube only by movement of gas from the buffers, gas flows from the buffers to the hot end of the pulse-tube, no gas flows at the hot end of the regenerator, and gas flows from the pulse-tube to the cold end of the regenerator. When the pressure decreases in the pulse-tube only by movement of gas into the buffers, gas flows from the hot end of the pulse tube to the buffers, no gas flows at the hot end of the regenerator, and gas flows from the cold end of the regenerator to the pulse-tube. Thus the regenerator loss is decreased. The efficiency of the active-buffer pulse-tube refrigerator is high. 3
9
EXPERIMENTAL RESULTS A prototype single stage active-buffer pulse-tube refrigerator with three buffers was made and tested to verify its performance. Its regenerator is a stainless steel tube of inner diameter 54 m m , wall thickness
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0.5mm, and length 94 mm, filled with 200 mesh stainless steel screens of packing factor 0.27. Its pulsetube is a stainless steel tube of inner diameter 49 mm, wall thickness 0.5 mm, and length 202 mm. The hot end of the pulse-tube is cooled by water. The volume of each buffer is 4.4 liters. The refrigeration temperature was measured by a Pt-Co thermometer. An electric heater was installed to measure the cooling power. Every valve is electric and is controlled by a personal computer. The compressor is a water-cooled GM refrigerator compressor with a standard input power of 3.3 kW. Experiments were done at 1.8 MPa of initial charge pressure of helium gas. Figures 2 and 3 show a typical cooling power and percentcarnot, respectively. The operating frequency was 1.9 Hz. The cooling power at 80 K was 160 W at an input power of 3.65 kW. The percentcarnot was 11% at 80 K. The lowest temperature achieved under these conditions was 34 K. The lowest temperature achieved was 28 K at 3.0 Hz with different valve timing.
180 A
160
~
-,
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100
./:
80 0
o
60 40
./
20 0
30
Z
/-
120 o 0
11 10
L F
0
_/
.4-,a
o o o
7
40
50
60
70
Temperature, K
Figure 2 Cooling power
80
90
8 7 6 5 4 3 2 1 0
J
/
/-
/
30
40
50
60
70
80
90
Temperature, K
Figure 3 Percentcarnot
Figure 4 shows the measured pressure waves in the buffers, the pulse-tube, and the compressor under the 80 K conditions of figure 2. The pressure wave t~ in pulse-tube is almost trapezoidal, showing the opening and closing process of the valves. The pressure ~ in buffer 3 is near the high pressure f'n of the compressor. The pressure P2 in buffer 2 is moderate. The pressure t] in buffer 1 is near the low pressure P~, of the compressor. This figure shows that the pressure differences across valve VH and valve VL are greatly decreased. Most of the intrinsic loss is removed. Figure 5 shows the equivalent PV diagrams for the gas in the cold end and the hot end of the pulse-tube as analyzed from a single cycle of pressure waves in the pulse-tube and the buffers of figure 4. The length of the pulse-tube is normalized. The cold end is 0, and the hot end is 1. The PV diagrams show displacement, expansion, and compression of gas in the pulse-tube.
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I
I
2.3
2.4
2.2
~
2.0
~ g
1.8
I
'
I
'
I
'
2.2 2.1 2.0 .9 1.8
1.6
1.7
1.4
1.6
1.5
12
1.4
10 0.8 0
200
400
600
Time, ms
Figure 4 Pressure waves
800
1000
1.3 1.2 0.0
0.2
0.4
0.6
0.8
1.0
Normalized length of pulse-tube
Figure 5 Equivalent PV diagram
CONCLUSION This paper has introduced a highly efficient active-buffer pulse-tube refrigerator. Buffers are connected to the hot end of the pulse-tube through on/off valves. The process of displacement, expansion and compression of gas is controlled not only by the compressor, but also by the buffers and on/off valves. The pressure difference across the high pressure valve or the low pressure valve is decreased. The regenerator loss is also decreased. A prototype single stage pulse-tube refrigerator was made and tested. Cooling power of 160 W was achieved at 80 K with an input power of 3.65 kW. The percentcarnot was 11% at 80 K. The lowest temperature achieved was 28 K.
REFERENCE Zhu, S., and Matsubara, Y., Proposal for a tube expander, Cryogenics(to be published)
AN I N T E R - P H A S I N G PULSE TUBE R E F R I G E R A T O R F O R HIGH R E F R I G E R A T I O N
EFFICIENCY
J. L. Gao I and Y. Matsubara 2
1 Suzuki Shokan, Kawagoe, Saitama 3 50, Japan Nihon University, Funabashi, Chiba 274, Japan
2
Pulse tube refrigerators, in general, require a phase controlling system located at the pulse tube hot end to obtain an optimum phase between the pressure fluctuation and the movement of the working gas in the pulse tube cold end to increase the regenerator performance. In the case of the double inlet and the fourvalve pulse tube refrigerators, the additional compressor work is exhausted to get the optimum phase which reduces the overall performance of the refrigerator. This paper describes an inter-phasing pulse tube refrigerator with a new concept of a phase controller for increase in the overall refrigeration efficiency. In this pulse tube refrigeration system, there are two sets of the coolers one of which consists of a pulse tube and a regenerator. The pulse tube hot ends of two coolers are connected by either an orifice valve or an on-off valve. Two coolers are operated 180 degrees out of phase only using one valved compressor. By controlling either a mass flow rate through the orifice or a timing for the on-off valve, the phase between the pressure and the mass flow controlled by each other of two pulse tubes could be optimized. Test and analysis results of the performance of this interphasing pulse tube refrigerator are presented in comparison to that of other types of pulse tube refrigerators.
INTRODUCTION In principle, the refrigeration effect in the pulse tube is basically obtained by the process of gas adiabatic expansion, which is similar to that in the Stifling and Gifford-McMahon (G-M) cycles. In the case of the Stirling and G-M cycles, the gas expansion work is transferred out of the expansion space by a solid expansion piston or a displacer. The important factor to operate the Stirling and the G-M refrigerators is the time-phasing The pistons in the Stifling refrigerator or displacer and valves in the G-M refrigerator have to move with correct relative timing for desired refrigeration cycle. With the correct time-phasing, the large cooling capacity per unit mass flow rate could be obtained to decrease the regenerator loss. In the pulse tube refrigeration, the oscillating gas column within the pulse tube functions as a compressible displacer (gas piston). The gas expansion work is transported by this oscillating gas column from the cold end to the hot end of the pulse tube. The correct phasing for the gas piston's movement can be achieved by improvement of the phase shifter located at the pulse tube hot end. Several types of the phase shiflers were developed as an alternative method to the simple orifice pulse tube for further improvement in performance, such as moving plug (second piston) [1], double inlet[2], and fourvalve[3]. In the case of the double inlet and the four-valve pulse tube refrigerators, the additional compressor work is exhausted to get the time phasing which reduces the overall performance of the refrigerator[4]. In the present investigation, we developed another modified pulse tube refrigerator named by Inter-Phasing Pulse Tube Refrigerator for increase in both refrigeration efficiency and simplicity. This paper describes the thermodynamic consideration and verification experiment on the inter-phasing pulse tube refrigerator. 295
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THERMODYNAMIC CONSIDERATION
G-M
I-P
S-O
Ph For simple understanding, the equivalent PressureVolume (P-V) diagram at the cold end of the pulse :3 tube was introduced to express the refrigeration t~ t~ process in the pulse tube refrigerator[3]. The S-O Pm' 13.. diagram in the Figure 1 as the equivalent P-V diagram in the pulse tube cold end could be used to describe the refrigeration process of the simple orifice pulse tube refrigerator. Since there is only a mean pressure source in the reservoir to phase the Expansion volume gas inside the pulse tube, the simple orifice pulse Figure 1.Simplified expansion PV diagrams of refrigerators tube produces a large mass flow through the regenerator cold end which increase the Compressor regenerator loss, and the gas adiabatic expansion is ! finished above the mean pressure Pm which i\J provides only small cooling capacity. By using the double inlet and the four-valve phase shift v l l ~'~ valves vs mechanism where the high and low pressure gas "source from the compressor was used to phase the gas inside the pulse tube, the equivalent P-V Jill'-diagram at the pulse tube cold end would be '---improved to close to that of the G-M cycle , ... refrigerator as shown in the G-M expansion P-V diagram of Figure 1. It would be noted that, to realize the refrigeration cycle similar to that of the G-M in the double inlet and the four- valve and to Figure 2. Basic configuration of inter-phasing pulse tube keep refrigeration temperature stable, the mass flow rate through either the second inlet tube or V4 Opened two valves at the pulse tube hot end should be changeable through the different process during V3Opened one cycle, which makes operation very difficult, v2 Opened Another problem in the four-valve and the double .... V1 Opened inlet pulse tube is that the additional compressor work is required to perform the time phasing, Opened UnOpened which reduces the overall performance of the refrigerator. Time lone cycle) Here is described a pulse tube refrigeration Figure 3. Typical timing for five valves system with new concept of a phase shift mechanism as shown in Figure 2 to increase the overall refrigeration efficiency and the simplicity. In this pulse tube refrigeration system, there are two sets of the coolers one of which has a pulse tube and a regenerator. The pulse tube hot ends of two coolers are connected by either an orifice valve and an on-off valve V5. Two coolers are operated 180 degrees out of phase with main valves V1, V2, V3 and V4 by a valved compressor generally used for G-M refrigerators. By controlling either a mass flow rate through the orifice or a timing for the onoff valve, the time phase for the gas inside the pulse tubes could be optimized by each other of two pulse tubes. The refrigeration cycle is expressed in the I-P diagram of the Figure 1 with the typical timing for the valves in Figure 3. The equivalent expansion work We produced when a gas at the pulse tube cold end is expanded cyclicly is (per cycle) We = ~ P d V e (1) where P is the pressure in the pulse tube, V~ is the equivalent expansion volume variation in the pulse tube cold end. This quantity can be represented by the area within the closed curve describing the cycle in the equivalent P-V diagram as shown in the fight of the Figure 1.
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ICEC16/ICMC Proceedings VERIFICATION TEST RESULTS
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30-_ _
To confirm the analytical results described before, the experimental device of single stage pulse tube refrigerator was fabricated of which performance was measured for several phase shitters, such as orifice inter-phasing (the pulse tube hot ends of two coolers are connected by an orifice), on-off valve inter-phasing(the pulse tube hot ends of two coolers are connected by an on-off valve), bypass interphasing (the pulse tube hot ends of two coolers are connected by an orifice, and a bypass was used to connected between the pulse tube hot end and regenerator hot end of each cooler), general simple orifice and general double inlet. The pressure oscillation was generated by a valved helium compressor. The compressor was connected to the pulse tube refrigerator by either a rotary valve or solenoid valves. Two regenerators consist of a stainless steel tube with 0.5 mm wall thickness, 39 mm inner diameter and 110 mm length, filled with ~30 no. 300 mesh copper screen disks. Two pulse tubes consist of a stainless steel tube with 0.5 mm wall thickness, 19 mm inner diameter and 160 mm length. At both ends of the pulse tubes no. 80 mesh copper screen disks are brazed for flow straightening and heat exchange. At room temperature the heat exchangers are simply air cooled At the cold ends of the pulse tubes, the temperature is measured with a rhodium-iron thermometer. A heating resistance is provided at a cold end for cooling power measurements at particular temperature. Figure 4 and Figure 5 show the test results of five different phase shitters with solenoid valves at an optimum operating condition for each type of the phase shifter, The test results indicated that the inter-phasing method for the pulse tubes gave better refrigeration performance that of previous phase shitters. 9
CONCLUSION
~,25~ 20N ~ 15=10 ~
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."
120
Simple orifice Orifice inter-phasing
-"
Bypass inter-phasing
--.--
On-offvalve inter-phasing
-i-
Double inlet
Figure 4. Minimum temperature and cooling capacity of five types of the phase shitters
///
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~ 10-
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---e-- Orifice inter-phasing
"Bypass inter-phasing 1. For increase in refrigeration efficiency of the pulse tube refrigerator, an inter-phasing --e-- On-offvalve inter-phasing mechanism was developed to obtain the timephasing for the pulse tube without using an Double inlet additional compressor work in comparison to the four-valve and the double inlet pulse tube. Figure 5. Refrigeration performance of five types 2. Test results for verification and comparison of the phase shitters
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ICECI 6/ICMC Proceedings Compressor
Compressor
I>
Valves V1
V4
to7 fSl
Valves
~ 1st stage cold end
(80K)
1st stage cold end
(80K)
2nd stage cold end (20K)
Thermal Bridge
II
Counterflow Heat Exchanger
f, 2nd stage cold end (20K)
Figure 6. Two-stage pulse tube refrigerator using inter-phasing method
3rd stage cold end (4K)
Figure 7. Three-stage pulse tube refrigerator using inter-phasing method with counterflow heat exchanger for third stage
indicated that the refrigeration efficiency of the inter-phasing pulse tube is higher than that of the simple orifice and the double inlet pulse tube. The refrigeration system using the inter-phasing mechanism for the pulse tube could be made compact by elimination of a gas reservoir. 3. To get much lower temperature by the inter-phasing pulse tube, the two stage and three stage pulse tube refrigerators would be proposed as shown in Figure 8 for 20 K and 4 K applications. In the configuration of three-stage pulse tube refrigerator using the inter-phasing method for the pulse tube, the counterflow heat exchanger could be used for the pulse tube to reach 4 K without using magnetic regenerator materials. REFERENCES 1. Matsubara, Y. and Miyake, A., Alternative methods of the orifice pulse tube refrigerator, Proc. Fitth Intl. Cryocooler Conf., (1988) 127 2. Zhu Shaowei, Wu Peiyi and Chen Zhongqi, Double inlet pulse tube refrigerator: An important improvement, Cryogenics 30 (1990) 514 3. Matsubara, Y., Gao, J.L., Tanida, K., Hiresaki, Y. and Kaneko, M., An experimental and analytical investigation of 4 K pulse tube refrigerator, Proc. Seventh Intl. Cryocooler Conf., (1992) 166 4. Gao, J.L., Hiresaki, Y. and Matsubara, Y., A hybrid two-stage refrigerator operated at temperatures below 4 K, Adv. Cryog. Eng. 41 (1996) in press
Experimental Research on Two-Stage Pulse Tube Refrigerator
Tatsuo Inoue, Takayuki Matsui, Shin Kawano and Yoshimasa Ohashi AISIN COSMOS R&D Co.,Ltd. 5-50, Hachiken-cho, Kariya, Aichi, 448 JAPAN
For the study of multi-staging of pulse tube refrigerator, we have constructed 4valve type 2-stage pulse tube refrigerator that has two sets of phase shifting system. It was found that finite phase difference between 1st and 2nd pulse tube is effective, and the lowest temperature of around 9K was achieved with normal lead shots as 2nd regenerator material. We also evaluated the 1st stage cooling power dependence on the 1st pulse tube size, and we obtained cooling powers of 1st and 2nd cold stage of 22W (at 80K) and 3.0 W (at 20K) respectively, with power consumption of 2.8kW.
INTRODUCTION As a matter of course, a typical feature of the pulse tube refrigerator (PTR) is "no moving parts" in the cold space. Another feature is flexibility of phase shifting and its wide range of regulation, and this feature produces some superior effects especially for multi-stage PTR. Generally, the phase shifter of the PTR has two functions, one is adjusting displacement amplitude of fluid oscillation, and the other is determining the phase difference between pressure oscillation and fluid displacement. Each of these two functions corresponds to the stroke and timing of a displacer (piston) motion of Stirling or GM refrigerator. In the case of multi-stage Stirling or GM refrigerator with single displacer piston, the stroke volumes of each stage is fixed and motion of fluid at each cold stage occurs simultaneously. On the other hand, for multi-stage PTR, both the amplitude of displacement and the phase can be easily adjusted independently each other for each stage. We think this may be the superior merit of PTR to other regenerative refrigerator. We have studied the multi-staging of pulse tube refrigerator, especially focusing on the optimization of phase adjusting for each stage. Some groups have examined multi-stage PTR of orifice type[ 1], double inlet type[2], 4-valve type [3, 4]. We have constructed 4-valve type 2-stage pulse tube refrigerator that has two sets of phase shifting system at the hot end of each pulse tube, because 4-valve type is more active way and easy to evaluate the phase condition quantitatively. In this paper, we report the optimization of phase adjusting for each pulse tube and the 1st stage cooling performance dependence on the size of 1st pulse tube.
EXPERIMENTAL APPARATUS The schematic diagram of the two stage PTR (test apparatus) is shown in Figure 1 and main specifications and operating conditions are listed in Table 1. The type of the PTR is 4-valve type, which has a set of pressure switching rotary valve and a flow controlling valve at the hot end of the each pulse tube. These two phase shifters are driven independently each other, therefore any phase difference of the pressure switching timing can be selected optionally. The pressure oscillation generator consists of a conventional compressor for GM refrigerator and high/low pressure switching valve unit. The cross299
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sectional view of the pressure switching valve unit is drawn in Figure 2 schematically. Although three rotors for the 1st regenerator and each pule tube are driven by one motor, the switching phase of each rotor ( @~, @2 ) can be adjusted independently by turning the housing of each rotor. Table 1 Specification of experimental apparatus and adopted operating conditions Items 1st Pulse tube
Specification or Conditions
q~18~-
q528 • 150 X t 0 . 5
1st Regenerator
38 • 100 • t 0.5 / Material:Bronze Mesh #250
2nd Pulse Tube
13 • 265 X t0.5
2nd Regenerator
20 • 130 •
/ Matefial:PbShot ~0.2
Mean Pressure
1.5 MPa (Working Fluid : Helium)
Operating Frequency
3.0 Hz
RESULTS AND DISCUSSION Switching Phase Dependence Figure 3 and Figure 4 show switching phase (69) dependence of reachable temperature of 1st and 2nd cold stage in case when the 1st pulse tube diameter is 18mm. @ is defined as the phase by which the 1st or 2nd pulse tube's switching valve leads that of the regenerator. The period of high pressure valve opened and that of low pressure is almost equal for each rotary valve. In the case of Figure 3, both 1st and 2nd stage reach their lowest temperature at around 80 degree of O1 when 02 is fixed to 30 degree. As shown in these figures, there are obvious optimum phase (timing) difference @and they are different (not equal) for each stage pulse tube. In these series of experiments, we obtained the lowest temperature of around 9K with lead shot as the regenerator material. Pulse Tube Size Dependence of the 1st Stage Figure 5 and Figure 6 shows the results of the reachable temperature and the cooling power of each cold stage as a function of the diameter of 1st pulse tube. In these experiments, the phase shifters were regulated to suitable condition experimentally. These results show that cooling performance of 1st stage strongly depends upon the pulse tube size (the diameter, in this case). As the diameter is changed from 18 to ~ 28, the reachable temperature simply decreases from 68K to 43K, by contrast, the cooling power at 80K increases from 4W( ~ 18) to 18W( ~ 26). These strong dependence upon diameter of pulse tube may be qualitatively understood that the increase of diameter makes the swept volume of the virtual gas piston large without much increase of heat load via the pulse tube. The Cooling Power of Each Cold Stage The cooling power of 1st and 2nd cold stage are shown in Figure 7 and Figure 8, respectively. The results were achieved only for ~ 26 of 1st pulse tube diameter and two levels of mean pressure (Pm) of 1.5MPa and 1.7MPa were adopted. As both figures show, cooling power is in linear relation to temperature at the examined range, and we obtained a cooling power of the 1st stage of 22W at 80K and that of 2nd stage of 3.0 W at 20K, with the power consumption of 2.8kW. Slight mean pressure dependence of cooling power was observed as in figures, and the obtained cooling power (or efficiency) are not so small compared with that of GM refrigerator. This rather good performance suggests that multi-stage PTR has the potential of high cooling performance comparable to other regenerative refrigerators, because in the case of PTR, the optimization of fluid controlling for each stage can be realized by operating each phase sifter independently, in spite of the ineffective volume of pulse tube.
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CONCLUSION A 4-valve type 2-stage PTR has been constructed and it was found that a finite phase difference between 1st and 2nd pulse tube is effective for achieving higher performance. By the optimization of two phase shifters, around 9K of the lowest temperature was achieved with normal lead regenerator. We also evaluated the 1st stage cooling performance (the reachable temperature and the cooling power ) from the viewpoint of its dependence on the size of 1st pulse tube, and its strong dependence was found. By choosing suitable 1st pulse tube size and phase shifting, we achieved cooling powers of the 1st and 2nd stage of 22W (at 80K) and 3.0 W (at 20K) respectively, with power consumption of 2.8kW.
REFERENCES 1. Tward, E., Chan, C.K. and Burt, W.W. Pulse tube refrigerator performance Advances in Cryogenic Engineering, (1990) 15 1207-1212 2. Ohtani, Y.,Takahashi, M., Kurityama, T. and Nakagome, H. Two stage pulse tube refrigerator for 20K operation Cryocoolers 8 Plenum Press, New York (1995) 337-343 3. Tanaka, M., Kodama, T., Nishitani, T., Araki, T., Kawaguchi, E. and Yanai, M. Two stage pulse tube refrigerator with double rotary valves Proceedings of International Cryogenic Engineering Conference-15 Butterworth, UK (1994) 159-162 4. Tanida, K, Gao, J.L., Yoshimura, N. and Matsubara, Y. Three-staged pulse tube refrigerator controlled by four-valve method, presented at CEC/ICMC (1995)
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or shaft .......... ..... ..i~iiii~....... i~i~:~
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PERFORMANCE OF THE HYBRID TWO-STAGE R E F R I G E R A T O R
K. Tanida, ] J. L. Gao, 1 Y. Hiresaki, 1 and Y. Matsubara 2 1 Suzuki Shokan, Kawagoe, Saitama 350, Japan 2
Nihon University, Funabashi, Chiba 274, Japan
The hybrid two-stage refrigerator with no moving parts at low temperature stage has been developed for cryogenic applications. In this hybrid configuration of a two-stage refrigerator, the first stage is constructed by a Gifford-McMahon (G-M) cycle with a solid displacer and the second stage has the configuration of the double inlet pulse tube with no moving parts. In previous studies, 3.2 K of the minimum refrigeration temperature and 0.37 W of the cooling capacity at 4.2 K was achieved by this hybrid two-stage refrigerator which has a pulse tube of 12 mm inner-diameter and a reservoir of 0.283 L volume. This paper describes the performance of the refrigerator with different sizes of the pulse tube and the reservoir. Measurement results of heat generation at the pulse tube hot end, pressure fluctuation in the reservoir, cooling capacity, minimum temperature and refrigeration temperature stability are presented.
INTRODUCTION In an effort to present a 4 K cooling system without using liquid helium, several types of 4 K refrigerators have been developed to meet requirements from a growing number of superconducting devices and detectors and other cryogenic applications, such as 4 K Gifford-McMahon (G-M), 4 K Boreas and 4 K pulse tube[ 1][2][3]. They are competing with each other in performance, reliability, lifetime and cost. We have developed the hybrid two-stage refrigerator with no moving parts at low temperature stage for cryogenic applications. In this hybrid configuration of a two-stage refrigerator, the first stage is constructed by a Gifford-McMahon (G-M) cycle with a solid displacer and the second stage has the configuration of the double inlet pulse tube with no moving parts. In previous studies, 3.2 K of the minimum refrigeration temperature and 0.37 W of the cooling capacity at 4.2 K was achieved by this hybrid two-stage refrigerator[4]. This paper describes the performance of the refrigerator with different sizes of the pulse tubes and the reservoirs. Measurement results of heat generation at the pulse tube hot end, pressure in the reservoir, cooling capacity, minimum temperature and refrigeration temperature stability are presented. PERFORMANCE OF HYBRID REFRIGERATOR Figure 1 shows the construction of the two-stage hybrid refrigerator. The first stage is the G-M cycle. The displacer installed with regenerator materials reciprocated by a drive motor with a scotch yoke mechanism. The high pressure and low pressure gases were controlled by the slide spool valves having the correct open-close timing with displacer's movement to go through the G-M cycle. The double inlet pulse tube configuration was used for the second stage. The pulse tube hot end was thermally connected with the first stage cold end by a copper plate. A reservoir was connected to the 303
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ICEC16/ICMC Proceedings 4K Cold Head -~--.
Second Stage Regenerator Pulse Tube
J
Orifice Bypass zz ~
Reservoir First Stage Displacer First Stage Regenerator Seal Outlet Valve Inlet Valve
Figure 1. Construction of the two-stage hybrid refrigerator pulse tube hot end through a stainless steel capillary tube as an orifice. Another stainless steel capillary tube as a bypass tube was also inserted between the pulse tube hot end and the second stage regenerator hot end. The flow resistance of the orifice and the bypass was controlled by changing the inner diameter of these two stainless steel tubes to obtain the optimum phase shift effect for increase in the performance of the second stage regenerator. At both ends of the pulse tube at the second stage the bronze mesh was brazed for flow straightening and heat exchange. Rhodium-iron resistance thermometers were used to measure the temperatures at the cold ends of the first and second stage. Heating resistances were provided at the cold ends of the second and first stage for cooling power measurement at a particular temperature. The heat generation at the pulse tube hot end was measured with the same method as that in our previous study[4]. A pressure sensor was used to measure the pressure in the reservoir. The performance of the refrigerator was measured with different sizes of the pulse tubes and the reservoirs. A description of four types is given in Table 1. Figure 2 is the cooling performance at the second stage cold end of four types. Table 2 is the Table 1. Description of pulse tubes and IType I Pulse tube [ 13 X 0.5 x 200 Reservoir 10.283 L in volume
reservoirs in four types Type IV ]Type II [Type III ~bl6 • 0.5 x 200 1 13 x 0.5 x 200 ]#16 x 0.5 x 200 10.40 L in volume 10.283 L in volume 0.40 L in volume
ICEC16/ICMC Proceedings 600
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---
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9
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Figure 2. Cooling performance at second stage cold end of four types 25 24 23 v" ~ 22 ~21 ,- 20 uJ -o 19 o 18 o 17 ~ 1 6 {13 co 15 -o 14 cO o 13 (D co 12 ~11 !,,-, 10 9 8 E 7 k-- 6 5 4 3
Cooldown Time about 6 Hours T e m p e r a t u r e at F i r s t S t a g e 22 K
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measurement results of heat generation at the pulse tube hot end and the pressure difference in the reservoir. The performance of Type I was described in the previous study[4]. In comparison to type I, the reservoir volume in Type II was increased from 0.283 L to 0.4 L without changing the pulse tube size. From test results in Figure 2 and Table 2, it could be indicated that the reservoir volume in Type I was enough for the pulse tube in Type I, because the performance was not changed so much by increase in the reservoir volume in Type II. From the cooling performance in Figure 2 and pressure difference in Table 2 for Type III, the reservoir volume in Type III seems to be not enough for the pulse tube size used in Type III. In the case of Type IV, the gradient of the cooling performance cure was greatly increased with both larger reservoir and larger pulse tube than that of Type I. The minimum temperature in Type IV, however, was slightly higher than that of Type I. It could be improved by optimization of phase shift effect for the pulse tube. To confirm the refrigeration stability at second stage pulse tube cold end, the hybrid refrigerator was operated continually for a period of over one week. As shown in Figure 3 the refrigerator displayed good temperature stability at the second stage cold end even through the mass flow rate from compressor was slightly changed during operation due to change in temperature of cooling water for the compressor. Table 2. Measurement results of pressure in reservoir and heat generation at pulse tube hot end Type II Type III Type IV Type I 0.050 MPa 0.0085 MPa 0.010 MPa Pressure difference 0.015 MPa in reservoir 6.05 W 9.52 W 7.8W Heat generation at 7.5W ~ulse tube hot end 23 K 21.8 K 23.4 K Temperature at 1st 22.65 K stage cold end 4.1K 3.7K 3.2K Temperature at 2nd 3.3K stage cold end CONCLUSION The performance of the hybrid refrigerator was measured with difference sizes of the pulse tubes and reservoirs. It is possible to increase the cooling capacity at 4.2 K by increase in the pulse tube volume. To achieve best refrigeration performance of the hybrid refrigerator, it could be required to optimize the phase shift effect for the pulse tube and to match well with all the volumes of the pulse tube, reservoir and regenerator. It was confirmed that the refrigerator displayed good temperature stability at the second stage cold end even through the mass flow rate from compressor might be slightly changed during operation due to change in temperature of cooling water for the compressor. For cryogenic applications, this hybrid 4 K refrigerator could have the advantages of improved reliability and low cost over other types of 4 K refrigerators. However, lack of knowledge in theory about periodic flow in the pulse tube makes optimization of refrigeration performance very difficult in matching periodic flow patterns through system to obtain a necessary phase shift effect for pulse tube to reach 4 K. REFERENCES 1. 2. 3. 4.
Kuriyama, T., et al., "Development of 1 Watt class 4K G-M refrigerator with magnetic regenerator materials", Advances in Cryogenic Engineering 39B (1994) 1335 Crunkleton, J.A., "A new configuration for small-capacity liquid-helium-temperature cryocoolers", Proceeding of 7th International Cryocooler Conference, (1993) 187 Matsubara, Y. and Gao, J.L., "Novel configuration of three-stage pulse tube refrigerator for temperatures below 4 K", Cryogenics, 34, (1994) 256 Gao, J.L., Hiresaki, Y. and Matsubara, Y., A hybrid two-stage refrigerator operated at temperatures below 4 K, Adv. Cryog. Eng. 41 (1996) in press
Improvement of The Two Stage Pulse Tube Refrigerator Reaching to 4K
Mineo Tanaka*, Tomio Nishitani**, Takao Kodama*, Tohm Araki**, Etsuji Kawaguchi*** and Masayoshi Yanai*** *Faculty of Science, Osaka City University; Sugimoto 3-3-138, Sumiyoshi-ku, Osaka 558, Japan. **Shiga Technology Center, Iwatani International Corp.; Katsube-cho 1095, Moriyama-shi, Shiga 524, Japan. *** Iwatani Plantech Co. LTD.; Katsube-cho 1095, Moriyama-shi, Shiga 524, Japan.
We have constructed a two stage pulse tube refrigerator equipped with three rotary valves. The sy stem can control the amount of the gas passing through the hot end of the first and second pulse tube, and to change their phase independently. The three valves are operated by stepping motor using timing belts. We have investigated the performance of the refrigerator by changing the operational conditions as well as the sizes and the materials of the second regenerator. However, the lowest temperature of the second stage remained rather high than that of 15.6 K obtained by the more simple system.
INTRODUCTION In the last decade, the understanding and the development of pulse tube refrigeration have been progressively advanced, and the lowest temperature obtained by the three-stage pulse tube refrigerator has reached below 4K [1]. For a single stage refrigerator, it is realized by now that the commercial products are available for the practical use at a working temperature of 77 K [2]. This cry ocooler has taken full characteristic advantage of vibration free and the reliability for a long term operation. In the previous work, we have constructed a two-stage pulse tube refrigerator equipped with two rotary valves, and attained the lowest temperature of 15.6 K [3,4]. Rotary valve is adopted as it has actual results for a long term work. In the two valve sy stem, one is located at the hot end of the first stage of the regenerator, and control the main gas flow. The other is located at the hot end of the first and second pulse tube, and works as a phase shifter. Therefore, the phase angle of the first and second stage could not change independently in the system. Apparently, this will limit the performance of the regenerator. Based on the above results, we have constructed new three valved refrigerator so that the phase angle of the first and second stage can be changed independently. In this paper, we report and discuss about the experimental results. EXPERIMENTAL The schematic drawing of the present two stage pulse tube refrigerator is shown in Fig. 1. In this series of experiments, the regenerator and the pulse tube of the first stage was fixed, and only second stage was changed. The size and the material of the first stage is as follows; the regenerator was made of 180 mm long, 0.5 mm thick sus tube of 39 mm OD, packed 1700 sus screens of #200. And the pulse tube was 300 mm long, 0.5 mm thick sus tube of 19 mm OD. For the second stage regenerator, three types are examined, and are shown in Table 1. 307
308 Table 1 type A
B
Table 2
ICEC16/ICMC Proceedings The size and materials of the second stage regenerator. tube size
regenerative materials lead shots (0.25-~0.3mm dia.)380g
stainless steel tube OD=I 8mm, thickness=0.3mm, length=250mm stainless steel tube OD=I 8mm, thickness=0.3mm, length=250mm
copper mesh#325, 825peaces and lead shots(0.3-~0.4mm dia.)260g
stainless steel tube OD=I 8mm, thickness=0.3mm
lead shots (0.3--0.4mm dia.) 173g length = 125mm
The size of the second stage pulse tube tube size
type 1
2 3
sus tube, OD=I 8 mm, thickness=0.5 mm, length = 370 mm sus tube, OD =16 mm, thickness=0.5 mm, length = 480 mm sus tube, OD =18 mm, thickness=0.5 mm, length= 480 mm Compressor
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Figure 1 Schematic diagram of two stage pulse tube refrigerator T y p e A is made similar to the one we reached 15.6 K in the previous experiments, but slightly smaller lead shots are used in this case. As considerable pressure drop was observed in Type A regenerator, some part of the lead shots are exchanged for copper meshes in T y p e B. In Type C, the length was reduced to one half to that of Type A. For pulse tubes, three types are examined as is shown in Table 2. As T y p e 1 gave best results, only this type was used in the following experiments. As for the regenerator, Type A and B did not make any significant difference in their performances. The capacity of each compressor is 40 m 3 per hour, its pressure ratio is 2.3, and the operational mean pressure is 1.6 Mpa. All data were taken at a frequency of 2.54 Hz. The temperature were measured by Au-Fe vs. chromel thermocouples at six points as shown Fig. 1.
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Figure 3 The temperature profile along the regenerator and pulse tubes. The location is the distance from the hot end of the regenerator. The second stage pulse tube is T y p e 1. The solid line shows the first and second regenerator and the dashed line shows the first and second pulse tube. RESULTS AND DISCUSSION As the phase angle of the pressure wave of the gas passing through the hot end of the first and second
310
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pulse tube can be changed independently, their phase angle dependence of the lowest temperature was mainly investigated in this experiments. The amount of the gas passing through the valve was controlled by orifice valve to find the optimum condition. Figure 2 shows the first and second stage cold end temperatures as a function of the phase angle of the second stage rotary valve(V2), and the phase angles of the first stage valve(V1)are shown as a parameter. In this measurements, Type A regenerator was used for the second regenerator. The phase angles of the valves are defined as follows; when the high pressure side of the compressor start to connect to the regenerator through Vr, and at the same time, to the pulse tubes through V1 and V2, the phase angle is defined zero degree. And the positive phase angle means when the connection to the high pressure side by V1 or V2 is advanced to that of Vr. For each setting of V1, the phase angle dependence of V2 become very poor above a certain degree. Similar behavior was observed when the Type A regenerator is replaced to Type B or C. Fig. 2 shows that the phase angle dependence of V1 is larger than that of V2, and the lowest temperature reached in this experiments was about 22 K. These results seem to indicate that the lowest temperature is limited by some other factor rather than the phase angle settings of V2. Figure 3 shows the temperature profile alongthe regenerator and the pulse tube for each type of the second stage regenerator A, B, and C. The feature of the temperature gradient along the pulse tube in Type C is different from those of Type A and B. The impedance of Type C regenerator is smallest among the three. Therefore, the gas flow through the regenerator will be the largest. The temperature distribution near the lowest temperature region seems due to the above reason. We can increase the amount of gas passing through V2 by opening the needle valve. This operation did make change the temperature distribution along the pulse tube, but did not help to reduce the lowest temperature. REFERENCES Matsubara, Y. and Gao, J.L., Novel configuration of three-stage pulse tube refrigerator for temperature below 4K Cryogenics(1994)34 259-262 New Cryo Mini., Iwatani Plantech Corporation, 1095 Katsube-cho, Moriyama, Shiga 524, Japan Tanaka,M. Kodama, T. Nishitani,T. Araki,T. Kawaguchi,E. and Yanai,M., Two stage pulse tube refrigerator with double rotary valves Cryogenics (1994)34 ICEC supple. 159-162 Tanaka,M. Kodama, T. Nishitani,T. Araki,T. Kawaguchi, E. and Yanai,M., Experimental and analytical study of two stage pulse tube refrigerator Advances in Cryogenic Engineering (1995)41 (CEC) to be submitted
Nitrogen Precooled Multi-Stage Pulse Tube Refrigerator Reaching 2.1 K
G. Thummes, S. Bender, and C. Heiden Institute of Applied Physics, University of Giessen, Heinrich-Buff-Ring 16, D-35392 Giessen, Germany
We report performance studies on a multi-stage pulse tube refrigerator (PTR) for operation at liquid 4He temperatures. In the current set-up the 1st stage consists of a liquid nitrogen bath with heat exchangers which serve to precool the oscillating 4He-gas flow passing two subsequent pulse tube stages. The two PTR stages are operated in orifice- and double-inlet mode with the phase-shifting components located at ambient temperature. Er3Ni shot is used for the regenerator matrix in the coldest stage. A minimum temperature of 2.13(1) K has been achieved.
INTRODUCTION The pulse tube refrigerator (PTR) offers the advantage of having no moving parts at low temperatures. This feature results in increased reliability and reduced vibrations of the cold head. Moreover, the extension to multi-stage PTR systems is mechanically quite simple. Recently Matsubara and Gao [1,2] have shown that a three-stage PTR with 4He as working fluid can reach no-load temperatures in the liquid 4He temperature range, when Er3Ni is used as regenerator material in the coldest stage. This progress in refrigeration has been. made possible by the introduction of the double-inlet mode of operation [3], which greatly improved the PTR performance, and by the development of new rare-earth based regenerator materials, such as Er3Ni. These materials exhibit anomalously large volumetric heat capacities below about 10 K, and thus enable an increase in regenerator efficiency at liquid 4He temperatures [4]. We are currently developing a three-stage pulse tube refrigerator for potential use in cooling of low-Tc superconductor devices. In order to facilitate the optimization of the multi-stage system, we have started with two PTR stages and liquid nitrogen precooling of the 4He working fluid [5].
EXPERIMENTAL SET-UP A schematic of the liquid nitrogen (LN2) precooled PTR is displayed in Figure 1. The pressure variation in the system is generated by means of a motor-driven rotary valve which periodically connects the cooler inlets to the high or low pressure side of a helium-compressor (Leybold, model RW 6000). The LNzreservoir contains two heat exchangers made of parallel bundles of thin copper tubes, one of which precools the main gas flow and the other one functions as the hot end heat exchanger of the 1st stage pulse tube. Two regenerative tubes filled with no. 180 mesh stainless steel screens provide thermal insulation between the LNz-cooled heat exchangers and ambient temperature. The pulse tubes and regenerators are fabricated from stainless steel tubes with outer diameter, wall thickness and length (~, x s x L) as follows 1st stage pulse tube 18 x 0.3 x 220 mm 3, 1st stage regenerator 30 x 0.5 x 210 mm 3, 2nd stage pulse tube 9 x 0.5 x145 mm 3, 2nd stage regenerator 20 x 0.5 x 206 mm3. The 1st stage regenerator matrix consists of lead shot (~0.18-0.23 mm) in the lower two thirds of the tube and a stack of no. 220 mesh stainless steel screens in the upper third. The 2nd stage regenerator is filled with Er3Ni spheres (~)0.15-0.3 mm). The two cold end heat exchangers are made of copper cylinders with thin axial flow channels. For a higher cooler performance, the two PTR stages are operated in orifice and double-inlet configuration. The orifice valves, second-inlet valves (needle valves, Nupro, type M), and the two reservoirs (1000 cm 3 each) are located at room temperature. The warm end of the 2nd stage pulse tube is connected to the room temperature assembly via a regenerative tube, as first introduced by Matsubara and 311
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312 .I
OV2
RT2 LNT
Figure 1 Schematic of the two-stage PTR with LN2precooling. RT1-RT3: regenerative tubes; LNT: liquid nitrogen reservoir with heat exchangers; PT1, PT2: 1st and 2nd stage pulse tubes; RG1, RG2: 1st and 2nd stage regenerators; VV: vacuum vessel; C: helium compressor, RV: rotary valve; OV1, OV2: orifice valves; DV1, DV2: double-inlet valves; RS1, RS2: reservoirs.
PT1--'[- I I !1 I ! -1"~RC1 RT3
RG2 PT2
W
Gao [1, 5]. The regenerative tube (RT3 in Figure 1) with dimensions of 7 • 0.5 • 369 mm 3 is filled with lead spheres ((I)0.3-0.5 mm). RT3 is not thermally connected to the LN2-reservoir. A new feature of the system in Figure 1 is that also for the upper stage the phase-shifting components are coupled to the pulse tube hot end by means of a regenerator. Radiation losses are reduced by use of two copper radiation shields (not shown in Figure 1). The temperatures of the cold ends of the pulse tubes are monitored by use of two carbon-glass resistors (Lake Shore, model CGR-I-1000). Net cooling powers are measured by means of resistive heaters which are attached to the cold end heat exchangers of the two pulse tubes.
TEST RESULTS Optimization of the cooler has been performed by varying the frequency f of the pressure wave and the settings of the needle valves (see [5] for more details). The average pressure and the peak to peak pressure variation at the main inlet were approximately held constant at
= 19.5(2) bar and Ap = 9.6(3) bar. Figure 2 displays characteristic cool-down curves measured at f = 1.04 Hz. The temperature of the 2nd stage drops below 2.5 K after 3 hours of operation and reaches its minimum value ofT2 = 2.13(1) K atter 6 hours. Under these stationary conditions the 1st stage cold end is at ~ 18 K. The temperature at the joint between the 2nd pulse tube and the regenerative tube RT3 then is at 46 K. The heat load to the LNz-bath, as determined from the nitrogen evaporation rate, amounts to ~ 12 W. The net cooling powers of the 2nd stage are displayed in Figure 3. Upon increasing the frequency from 1.04 to 1.24 Hz, the net cooling power at 5 K increases from 30 to 45 mW at the cost of higher no-load temperature of 3.28 K. The increase in frequency is accompanied by a strong rise of temperature at the 1st PTR stage from 18 to 42 K, indicating a reduced cooling power of the 1st stage. Investigations of the performance of the 1st stage show that its optimum frequency is at 0.8 Hz [5]. Further improvement of the cooler with respect to a higher cooling power near 4 K therefore should be possible by matching the dimensions of the two PTR stages for the same optimum frequency. = 19.5 bar and Ap = 9.6 bar by use of the NIST12 4He-database show that Td decreases rapidly below 10 K, which is due to the anomalously large volumetric specific heat 9Cp, and it even drops to zero for T = 2.0 K, since 13 disappears. The adiabatic enthalpy variation Ah = dh, however, still remains. This is seen by inserting Eq. (2) into (1) leading to the simple relation Ah = 9 1 Pa 9 = 19.5 bar = 19.5 bar, Pd = Ap/2 = 4.8 bar. In conclusion, the present work demonstrates that a multi-stage PTR working with 4He gas can reach temperatures near the X-line. This result is of fundamental interest concerning the limits of pulse tube refrigeration. Future developments will aim towards higher net cooling powers around 5 K, as required for most practical applications.
DISCUSSION The achieved minimum no-load temperature of 2.13 K, which is only 0.19 K higher than the temperature T~(19.5 bar)= 1.94 K of the )~-line of 4He, is, to our knowledge, the lowest temperature reached up to now by regenerative cryocoolers operating with 4He. The question arises, whether the ;~-line represents a
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Figure 3 Net cooling power vs. cold end temperature of the 2nd PTR stage. Curve a: f = 1.04 Hz, parameters optimized for minimum no-load temperature. Curve b: f = 1.24 Hz, parameters optimized for maximum 0 at 6 K. low-temperature limit for pulse tube refrigeration with 4He. A natural limit should be given by the k-line, since below Tz the large heat transfer by the superfluid 4He will generate excessive heat losses. Neglecting losses, the ideal cooling power of the PTR is given by the time-averaged enthalpy flow < I2i >in the pulse tube (e.g. [1, 3]). It has been argued that < ISI> in the 4 K regime is greatly reduced due to the fact that the amplitude Ta of the oscillating fluid temperature is strongly diminished in the non-ideal 4He gas [2, 5]. However as shown below, the pressure dependence of H(p, T) in the real gas results in < 111> > 0 even when the dynamic temperature amplitude Ta drops to zero. From general thermodynamic relations the change in specific enthalpy dh(p, T) of the working fluid can be written as dh(T, p) = Cp d T + pl(1 - [3T) dp, where Cp, P, and 13 are the coefficient, respectively. For amplitudes Ta and pa = Ap/2. to < I~I > =
(J)
specific heat at constant pressure, the density, and the volume expansion small pressure variations dT and dp can be approximated by the dynamic For an ideal gas 13 - 1/T and then the 2nd term in Eq. (1) is zero, which leads mass flow rate). For adiabtic pressure variations in the pulse tube Td is given
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Td "" (13T/(pCp)) Pal.
(2)
Calculations Of Td at the pulse tube cold end for
(3)
It follows from Eq. (3) that < H > - 1/9 = dh/dp. For 4He below 4 K the density and thus the average enthalpy flow < I2I > is only weakly T-dependent. Figure 4 illustrates the variation of CpTd (first term in Eq. (1)) and Ah (Eq. (3)) below 20 K. Clearly, below 10 K the adiabatic enthalpy variation in the real 4He fluid is higher than in the ideal gas. 12
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ACKNOWLEDGEMENTS We thank U. H~.fner (Leybold) for providing the magnetic regenerator material. This work is supported by the German Ministry of Science and Technology (BMBF) under contract no. 13N6513.
REFERENCES 1 Matsubara, Y. and Gao, J.L., Novel configuration of three-stage pulse tube refrigerator for temperatures below 4 K Cryogenics (1994) 34 259 2 Matsubara, Y. and Gao, J.L., Multi-stage pulse tube refrigerator for temperatures below 4 K, In: Cryocoolers 8 Plenum Press, New York (1995) 345 3 Zhu, S., Wu, P., Double inlet pulse tube refrigerators: an important improvement Cryogenics (1990) 30 514 4 e.g. Kuriyama, T. et al. High efficient two-stage GM refrigerator with magnetic material in the liquid helium temperature region Adv. Cryog. Eng. (1990) 35B 1261 5 Thummes, G., Bender, S. and Heiden, C., Approaching the 4He lambda-line with a liquid nitrogen precooled two-stage pulse tube refrigerator Cryogenics (1996), to appear
Thermodynamic Calculation of Three-stage Pulse Tube Refrigerator*
Guobang Chen, Zhongming Xia, Li Zhao, Liming Qiu, Jiangyao Zheng, Jianping Yu Cryogenics Laboratory, Zhejiang University, Hangzhou 310027, P.R.China
The refrigeration capacity and thermodynamic losses of a three-stage pulse tube refrigerator have been calculated by means of the separate method of gas piston and its boundary layer. The influence of configuration and operating parameters on refrigeration performance is discussed.
INTRODUCTION Past ten years saw fantastic spurt in development of pulse tube refrigerator. The case in point is the liquid helium temperature reached in multi-stage pulse tube refrigerator[ 1]. Otherwise, its theoretical investigation isn't satisfactory, especially in multi-stage pulse tube. In this paper, the method which is always adopted in calculation of cooling capacity in G-M refrigerator is applied to count the theoretical cooling capacity of valved three-stage pulse tube refrigerator. Furthermore, main thermodynamic losses in the procedure are discussed. Finally, the net cooling capacities at each stage are obtained.
THERMODYNAMIC CALCULATION METHOD Thermodynamic calculation methods of a cryocooler include simultaneous and isolated ones. In isolated method, the premises are that irreversible losses do not exert any influence on ideal cycle and do not have any connection with each other, so they can be counted independently. Although the former which takes cycle and all kinds of losses into consideration is more accurate than the latter, it hasn't found wide application in engineering due to its tremendous work load. The application of isolated method in three-stage pulse tube refrigerator is consist of three steps: firstly, refer to certain physical model (isothermal or adiabatic model) and count the theoretical cooling capacity Q~; then, calculate the losses in the refrigerator respectively; finally, the actual cooling power Q~ is obtained.
(1)
Q~j - Q~j - ~ AQj + ~-] AQj+~ j
j+l
in which, Q~j represents the actual cooling power in pulse tube at j stage, Q~j is the theoretical cooling power at j stage, ~ AQj and ~ AQj+ 1 a r e the sums of irreversible losses at j stage and transferable irreversible J"
j+l
losses at j+l stage as its gain cooling. In order to carry out thermodynamic calculation, gas piston and trapezoid pressure wave in the pulse tube are basically hypothesized. The gas piston has the similar function to the solid displacer in G-M refrigerator. * The project is supported by the National Natural Science Foundation of China. 315
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THEORETICAL COOLING CAPACITY The shape of equivalent P-V diagram of the cold volume in a valved pulse tube refrigerator is similar to that of a G-M refrigerator, so its total area can be counted by means of calculation used for G-M refrigerator[2]. Under steady condition, the system energy equilibrium equation to the cold ends of pulse tube and egenerator is, (2)
Qc - ~dW
while the net work which is done by the cold volume to exterior ~ d W should be equal to the area of the equivalent P-V diagram. Therefore, the theoretical cooling capacity Q~ in a valved pulse tube can imitate the calculation of that in G-M refrigerator: Q~ - ~ d W - f (Ph - Pt)V~
(3)
where, V~ is the maximum of cold volume. The cold volume in pulse tube can be considered as the maximum volume of gas flowing into the pulse tube at the compression process.
V~ - I~C
(4)
where, I<. is the volumetric flow velocity at the cold end of pulse tube, t C represents the time of compression. Because there is no solid piston in the pulse tube, the regulation is only realized by the movement of the gas piston which is unsteady and lagged. At the moment of compression working fluid which goes into pulse tube is gradually compressed to be high other than immediately become high as in G-M refrigerator. Similarly, it goes to the cold end and becomes lower by degrees at expansion. Therefore, the theoretical cooling capacity of pulse tube refrigerator may be less than that of G-M refrigerator.
EFFICIENCY OF REGENERATOR The regenerator plays a dominant role in cryocooler since the heat transferred through it is about 10 - 50 times greater than the cooling capacity. Theoretical analysis and experimental results all show the outweighing part of regenerator loss, therefore, the efficiency of regenerator r/r exerts tremendous influence on overall efficiency of the cryocooler. In order to count the efficiency of regenerator, the calculation method applied in G-M refrigerator is employed[2]. The equation set is used as follows: ~mh) ~ + 6x
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Ot Finite difference discretization method is used to convert the differential equations into algebraic ones. Then, we define the enthalpy efficiency of the regenerator as r/i~,h - ( t t h - H) / (H h - H c,) in the warm period
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and q..~, - (H - H~) / (Hi, - H~) in the cold period, where H is the gross enthalpy at the entrance of cold or hot flow. The gross efficiency of regenerator is its algebraic average of the warm and cold periods.
IRREVERSIBLE LOSSES The actual processes of a pulse tube refrigerator unavoidably has irreversible losses including regenerator loss, shuttle loss, loss due to flowing impedance and pressure drop, axial heat conductive and radiative losses. The calculation method of later three losses can refer to reference [3]. The former two losses are chiefly discussed as follows. Regenera_tor loss Regenerator loss can be calculated by following equation A(2. - (1 -
) (:,,m,. (
(6)
- 72. )
where, m,. represents the mass flow through the regenerator in one cycle, q r is efficiency of the regenerator; and Cp is specific heat of helium at constant pressure. Shuttle loss In pulse tube refrigerator, the displacer is not solid but a gas piston whose volume and temperature are varied. When it locates at the cold end, it has the maximum volume, and when it moves at the hot end, it is compressed to the minimum. Its stroke Z approximately is the length of cold volume. In accordance to the boundary layer theory in fluid dynamics, when the gas piston moves in pulse tube, the velocities of the fluid at a cross-section are not equal but distributes along a parabola. The flowing velocity reaches its maximum at the axis of the tube and reduces to zero near the wall. The flowing velocity from the wall to boundary layer is about one tenth of center velocity. We assume the gas in the pulse tube is divided to two parts, i.e. the gas boundary layer where the velocity is zero and the gas piston where the velocities at same cross-section are averagely equal. The gas piston exchanges heat with walls of the tube through the thickness of the boundary layer, consequently causes shuttle loss. Based on the hypothesizes of the infinite thermal capacities of gas piston and pulse tube wall and the linear axial temperature distributions along gas piston and tube, shuttling loss can be expressed: AQ~ - 2cg zrL)Z 2 (7~, - 7;)
(7)
467_; where, 2~ represents thermal conductivity of helium in the gap, D is diameter of the tube, Z is stroke of the gas piston, 7;, and 7~. is temperature at the hot and cold ends respectively, L is length of the tube, ~ is gap 1
between the gas piston and tube wall, 6 - 4.647(0.5r
2.
Calculation results Based on the method above and the calculation conditions in Tab. 1, we obtain the results listed in Tabs.2,3.
CONCLUSIONS 1 .Based on the hypothesizes of gas piston in the pulse tube, a separate method of the boundary layer and gas piston can be adopted in thermal calculation and reasonable results for three-stage pulse tube refrigerator have been obtained.
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2.The chief irreversible losses at each stage of multi-stage pulse tube are regenerator loss and shuttle loss. But the former at first and second stages are smaller than the latter, while regenerator loss is obviously greater at the third stage. This is related to efficiency drop of regenerator at low temperatures. Table 1
Calculation conditions of three-stage pulse tube refrigerator
Pressure (MPa) Frequency (Hz)
High pressure 1.8
1st stage Temperature (K) 80 Pulse tube+regenerative el) 24 x 180 tube
Table 2
Low pressure 0.6 2 2nd stage 3rd stage 40 4.2 cI~ 17 x 210 q:)9 x 140+(D7 x 240
Configuration and efficiency of regenerators
Regenerator Size (mm) Matrix Volume(m 3) = 0.257rDZL Heat exchange area (m 2) A.,. - Ar / V
1st stage q~59 x 130 bronze 3.55E-4 4.8E+4
2nd stage cD45 x 60 lead 0.95E-4 1.3E+4
3rd stage q~ 19 x 190 lead+Er3 Ni 0.54E-4 1.2E+4
Equivalent hydraulic diameter(m) D - 4 p , / A Cross-sectional area (m 2) A r - 0.25reD 2
5.3 E-5 2.8E-3
10.1E-5 1.59E-3
12.8E-5 0.28E-3
Flowing area (m 2) At = q/Ar
1.79E-3
0.52E-3
0.11E-3
Half cycle 0.5f (Hz) Regeneraor efficiency
0.25 0.994
0.25 0.987
0.25 0.980
Table 3
Losses and actual cooing capacities in three-stage pulse tube refrigerator
Item(W) Regenerator loss Pressure drop Conductive loss Radiation loss Shuttle loss Gross loss Interstage heat trans. Cooling capacity Net refrigeration
1st stage 42.6 0.37 2.33 2.57 48.8 96.67
Ratio(%) 2nd stage Ratio(%) 44.07 9.76 33.28 0.38 0.15 0.52 2.41 2.21 7.53 2.66 2.01 6.85 50.46 15.2 51.82 100 29.33 100 27.17 4.81 104.6 30.6 35.10 6.08
3rd stage 293 031 0 22 0 76 1 66 5 88
Ratio(%) 49.83 5.27 3.84 12.92 28.23 100
0
5.92 0.04
REFERENCES Gao, J.L. and Matsubara, Y. Experimental Investigation of 4K Pulse Tube efrigerator, Cryogenics (1994) Vol. 34,No. 1,25-30 Hong, Chaoshen and Xu, Xiangdong, On the Unique Features of the Low Temperature Regenerator in a 4.2K G-M Refrigerator, Proceedings of JSJS-IV, (1993) 6 Walker, G. Cryocoolers, (Part I), Plenum Press, New York, (1983) Barton, R.F., Cryogenic System, 2nd Edition, Oxford University Press, (1985)
Three-staged Pulse Tube Refrigerator with Linear Motor Compressor
N. Yoshimura, Y. Matsubara, Y. Ohtani*, H. Nakagome* and H. Okuda** Atomic Energy Research Institute, Nihon University 7-24-1 Narasinodai Funabashi 274, Japan *Toshiba Research and Development Center, 4-1 Ukishima Kawasaki 210, Japan **The Institute of Space and Astronomical Science, 3-1-1 Yoshinodai Sagamihara 229, Japan
A three-stage pulse tube refrigerator with linear motor compressor has been designed and constructed. As the design concept of multi-staging pulse tube, the parallel configuration of pulse tubes has been selected. This method has been confirmed by the G-M type pulse tube refrigerator to give the cooling temperature down to 4 K. The G-M type cooler, in general, operated at one order lower frequency than the Stirling type cooler. A purpose of this study is to know the possibility of the high frequency operation for the space application with this parallel pulse tube configuration.
INTRODUCTION Pulse tube refrigerators will be classified into two categories, Stirling cycle type and G-M cycle type. In the case of G-M cycle type, multi-staged pulse tube has been successfully demonstrated to give the refrigeration at the liquid Helium temperature[l]. In the case of Stifling cycle type, however, the liquid Helium temperature does not obtained so far. The pulse tube refrigerator of G-M cycle type, in general, operating rather low frequency due to the requirement of the switching valve for generating the pressure oscillation. The Stirling cycle type, on the other hand, can operate much higher frequency because of the direct pressure wave generation by the compressor piston[2,3]. One of the purpose of this study is to know the possibility of the high frequency operation for the space application of the multi-staged pulse tube refrigerator to give the cooling temperature below 20 K. In this study, three-staged pulse tubes of parallel configuration have been used combined with the linear motor compressor.
LINEAR-MOTOR COMPRESSOR DESIGN The linear compressor of the moving coil type having the maximum input power of 600 W was designed as shown in Fig. 1 and fabricated based on the following considerations. The mechanical energy of the moving coil converted from the electrical energy generate the work flow caused by the gas pressure oscillation. To make this energy conversion effectively, the natural frequency of the moving components and the coil driving frequency should almost be same to be the resonant condition. Electrical circuit and the momentum equations of the idealized electro-mechanical system will be given as follows;
L] + RoI + BglXc = Eo sin(cot) Mc J(c' + Ks Xc + Col
J;
c + (:c'2J(c + Cc3 jr
(1) + de(Pc-/]7))
=
Bg/) / ~
(2)
where L; coil inductance, Ro; coil resistance, l; total length of winding 37o; position of compressor piston, Eo; 319
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supply voltage, M~; mass of the moving components, C~ ; dry friction coefficient Cc2 ; viscous friction coefficient, C~3 ;turbulent friction coefficient.A~ ;piston cross sectional area, P~; pressure in the gas compression volume, Po; back pressure of the compressor piston. The necessary condition to get the resonation at the natural frequency of fo is,
1 IKTg / = / ~ = 2--~ Mc
(3)
Linear Compressor Permanent Magnet
Current Coil
where the spring constant Kr will be given as the sum of the gas and the mechanical spring constant, Kg and K~.. The order of Kg will be given as,
Kx
=
Ac AP D
Ax
VOA2
y
;ton
(4)
_(X0Ac)
where AX ; piston displacement, V,., ; total equivalent volume when the piston located at its mid position, D P ; mean pressure of oscillating pressure. To prevent the side force of the compressor piston, K~ should be minimized by increasing the piston diameter. In this design, the ratio Kc, K~ of 0.12 - 0.17 has been selected. The pressure in the compressor volume is the function of the mass flow rate at the regenerator m~, the orifice 1#o and the double inlet line rod. If the pressure oscillation amplitude is relatively low, following simplified equations will be used for single stage pulse tube refrigerator.
,,,. =(:,.(Pc -Pe)
(s)
r/'o = (o(Pe - Pb )
(6)
ri,ci : ('ci (Pc - Pe )
(7)
r
Figurel Schematics of Linear Compressor
10 E
....... I ,! ...... i ..............
4
-a "~-
<E
__@ ......
2 0
i
-2 -4
-6 -8 -10
j
where Pe; pressure in the pulse tube, Pb; pressure in the reservoir. In the case of three-staged pulse tube refrigerator, above three equations for each stage must be involved. In this study, however, the calculation has been done by two set of separated programs. At first, solve the compressor part with single stage pulse tube and get the relation of the pressure and the compressor swept volume. Then, solve the three staged pulse tube refrigerator using this swept volume. Repeating those calculations until the pressure oscillation becomes to coincide each
~
Time Figure 2 Calculated result of linear compressor Table l_C0mpressor Parameter U--~,]c........i 2.65 kg R 4.5 ~2 ~-A-~..........i ~i6cmT . . . . . . . L . . . . . . i7.-8-mH i K~--~ 3-k:~rn ...... ...../ ....... i 3-5 ~)i I--B~...... i-Tesla -;:.... X,, -~ - 3 cm ...................................................... Table 2 Pulse tube & Regenerator Parameter lStage _---i-Pu}_ie.} u b e i i!Regenerator .............-_ -. . . . . /Materials / 1st {~ 14x0.5• ~ 34•215 =; 325 Cu Mesh i 22 • • ~) 8• 154 , 2nd {~ 0.3 Pb shot i 4} 13 • l 3rd 1~} 6 x0.5x180 , 4~ 0.3 Er 3 Ni shot
other, One of the calculated result for the compressor part is shown in Fig. 2. The actual data for each curve can be reproduced as follows by using the amplitude Y; Voltage E ....Y/0.05(Volt), current I--Y (A),stroke X -= Y200(m) and P--(Y• The calculation was done by the constant supplied voltage with the operating frequency of 14 Hz and the initial pressure of 1.5 MPa. Other input parameters for one of the opposed linear compressor are given in Table 1. t
.......................................................
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the current wave form becomes to an out of shape if the dry friction increases at the mismatching operating frequencyf with the resonant frequencyj~. This current wave becomes to a trapezoidal shape Reservoir Volume when the phase difference between the current and the applied voltage becomes to zero by optimizing the resonant frequency.
Fr~ C~176 ][
"e Coo'4
jil ~ ..,ceVa,ves
PULSE TUBES DESIGN
,, i !idi~i!~ ) i i ~ Three staged pulse tube refrigerator combined with 1st Regenerat Cu Me~-~h-~ ~ ;, the linear compressor was designed to give the 2nd Regenerato~r~ ] minimum pressure ratio of 1.5 at the compressor Pb Shot:volume. The parallel configuration of each pulse tube r,.e0ene,at2) has been used as shown in Fig. 3. Each dimension is Er3Ni Shot given by Table 2. The equivalent PV diagrams of each pulse tube cold end are calculated by the following Figure 3 Schematics of Three-staged Pulse tube method. Nine differential equations of the mass flow 1.9 rate (three set of equations (5-7)) can be solved 18 simultaneously by Runge-Kutta method with the boundary condition of the sinusoidal movement of ~ 1 . 7 the compressor piston. Using the concept of the a - 1 6 -4-- . . . . . . i equivalent PV diagram, PV works for each stage, [ itlii ......... i' __ PVI, PV2, PV3 and that of compressor volume, PVc =14 can be obtained as shown in Fig. 4. The calculation ~13 must be repeated by changing the each volume or valve opening rate until the PV diagram at the a - 1 2 compressor volume becomes to almost same with the 1.1 result obtained from Fig. 2. The by-path valves for each stage are opened properly to give the double 0 2O 4O 6O 8O 100 inlet operating condition. In this figure, each PV diagram is normalized to the room temperature. Volume(cm 3) Therefore the difference of PV work at the compressor and the sum of the equivalent PV works Figure 4 Equivalent PV diagram of Double i,~I~l" Pulse tubes and PV diagram of Compressor at each cold end of the pulse tube give the loss due to the pressure drop within the regenerators. 4-----------4
.
.
.
.
.
.
.
.
.
-.
.
.
.
.
.
300 EXPERIMENATAL RESULTS
~.250
An experimental set up having the dimensions of Table 1 and 2 has been constructed and tested. The cool down curves of each cold end of the pulse tube are given in Fig.5. Initial mean pressure is 1.5 MPa and the operating frequency is 14 Hz. It requires fifty minutes to get the minimum temperatures for each stage, 107 K, 61 K and 21 K respectively. Fig. 6 shows the measured wave forms. Each values are as follows; applied voltage to the compressor E--Y• 10.2(Volt), current I=Y• pressure P-(Y• x0.098(MPa) and the piston stroke X=(Yx0.17+l.65)(cm) of the compressor. These results indicate our simplified calculating method described above gives good prediction. PV diagram reproduced from Fig. 6 and
~200
--
~_150
............... - - - i
E 100
/
2nd-Stage
..................
0
0.5
age
i
1
1.5
Time(hour) Figure 5 Cool down curve of each stage
2
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ICEC16/ICMC Proceedings
Fig. 2 with that of Fig. 4 are given in Fig.7. The PV work obtained from the experiment was 320 W and the input work to the moving coil was 550 W. It gives the compressor efficiency of 60 %. The compressor work obtained from Fig. 6 and Fig. 4 are the almost same but there are slight difference with the experimental result. It may caused by the assumption of iso-thermal process for each stage pulse tube.
11)
20 15 10 5
0 E < -5 -10
.
,,..,..
I
I
1.8
,/" \ \
/_xl x'x
13.
\
/
/
~'1.6 1.4 1 0.8
_
0
20
Time
.
.
.
.
Fig. 8 shows the cooling capacity at each stage. X axis is the temperature difference based on each minimum temperature. Actual cooling capacity of 3rd stage pulse tube at 24 K was 0.5 W. It is noted that the first and second stage cooling performances are almost same in spite of different pulse tube sizes. This indicates the phase sift effect of first and second stage or each regenerator size does not well optimized. In the case of multi-staged double inlet method, however, the mutual interaction between each stage could not be avoided. Therefore it is important to find out the new phase sift mechanism instead of the double inlet method in the case of Stirling type multistage pulse tube refrigerator.
__~_
. . . . . .
. . . . . . . .
..
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
} Experimental lesult
40 60 Volume(cm3)
80
1O0
Figure.7 Comparison of PV diagram at the compressor volume
Figure.6 Experimental result of linear compressor
CONCLUSION
.
1.2
15 -20
t
obtained of Fig.2
2
1.5
v
2nd-stage ...................... i ........
i i /
i /
...... !............. .
.
.
.
.
si-sta0e
1
r
...........
E
0
i .............
,. . . .
i
-
---
0.5
0 0
2
4
Temperature
6 Difference
8
10 (K)
Figure 8 Charachteristic curve of Cooling Capacity
Three staged pulse tube refrigerator driven by the linear motor compressor has been constructed and tested at the operating frequency of 14 Hz and the initial pressure of 1.5 MPa. A computer simulation based on the concept of the equivalent PV work for multi-staged pulse tube refrigerator gave good prediction of the experimental result. The minimum temperature obtained in this study has been limited as 20 K, however, it was confirmed that the parallel configuration of the multi-stage pulse tube could be used for the Stifling type pulse tube operating around 14 Hz. Further study for the higher frequency and lower temperature should be done. The authors would like to thank Dr. Ravikumar and Mr. Zhou for their assistance. References 1.Y.Matubara and J.L.Gao "Novel configuration of three-stage pulse tube refrigerator of temperatures below 4K" Cryogenics(1994)34p259 2.Kuriyama, T.et al., "Development of Pulse tube Refrigerator with Linear-motor Drive Compressor", Advance in Cryogenic Engineering Prenum Press, NY, (1994) in print 3.E.Tward, C.K.Chen Zhonogqi,"W.W.Burt, "Pulse Tube Refrigerator Performance," Advances in Cryogenic Engineering, 1989 Proceedings, Vol. 35B,(1990),pp. 1207-1212.
Cryocoolers
G-M
coolers
This Page Intentionally Left Blank
Fffectiveness of Magnetic Regenerator Material with Low Tc below ~20 K -Vew Effective Regenerative Character of Magnetic Matena!sTakasu Hashimoto* Tokyo Institute of Technology, 2-12-1, Oh-okayama, Meguro, Tokyo 152, Japan Since ~1987 application of the high heat capacity magnetic materials to regenerator materials in low temperature range below ~10 K have been developed considerably. In this paper, firstly experimental results of specific heat C in several candidate materials are shown and the possibility of these heavy rare earth magnetic materials as regenerator materials in He temperature range will be discussed. Then, in order to prove usefulness of the magnetic materials as regenerator materials, these materials were applied to the low temperature side regenerator in the low cycle refrigerators such as GM refrigerator. In this case, the peaks of C in magnetic materials is very sharp in comparison with that of He-gas and, therelore,multi-layered type regenerator was prepared. As the results, it is verified experimentally that these magnetic matevals have the superior character as magnetic regenerator matena!s. INTRODUCTION Several novel applications of the metal superconductors have been performed. In order to accelerate considerably those application, however, development of the He liquefier with high efficiency is necessary,because metal superconductor must be used in low temperature range near the boiling point of liquid He. Especially in low temperature range below ~ 10K, so as to make effectively the small power He-liquefier, the regenerator with high efficiency is necessary. Until end of 1980, the non-magnetic metals such as lead,copper and their alloys had been used as the most important regenerator material in low temperature range, which have several suitable physical properties for application as regenerative materials, though the most important characters, heat capacities of those metals are very small in comparison with that of He-gas below -- 1OK. At the end of 1980, very sharp peak of magnetic specific heat near the critical temperature of the order-disorder phase transition, which is the same order of magnitude of He gas near --4K[ ! ], was tried frequently to apply to the regenerator materials near -~4K range. Therefore, in this section,we would show at first"why the magnetic material must be used ior the regenerator materials." The specific heat of normal metals The heat capacities of non-magnetic metals originate from themaal vibration of the atomic lattice and is approximately expressed by the Debye mcxtel[2] as T (1) C = 9Nk,(~-o)-"~i ( ~x~e~ - ]): dr, , where x=hv/kBT, xD=OffT and OD is the Debye temperature. This equation can be well reproducible the specific heat of the normal metal (non-magnetic metals) such as lead and its alloys. * present; Emeritus Professor of Tokyo Institute of Technolory 325
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On the basis of the eq.(1), in the very low temperature range approximately given by the following fomlula, Cv
=
234(i~)3
kBT<
C\ is
(2)
z
This result shows clearly that Cv decreases in proportional to "/"~)as shown in Fig. 1. Moreover, the lowest value of OD in metal is at most --100K, as a result, Cu of those metals become fairly smaller than Cv of pressurized He-gas below 10 K and the lead and its alloys become ineffective for regenerator operation, because these Cv becomes very smail. The specific heat accompanied with the order-disorder phase transition According to thermodynamics, the specific heat C is expressed as
as,) c : T(-Y'i where Sj is the magnetic entropy. In the temperature range below --I()K, the most promising physical mechanism to give a large specific heat is the order-disorder magnetic phase transition at Tc where the value of Sj changes suddenly from the order phase(Sj=()) to the disorder phase (large value of Sj). Considering eq.(3), the magnetic specific heat takes a !ar,,e._ peak near the order-disorder phase transition. According to the thermodynam'icai the<>ry, the magne:ic entropy is obtained from Cv value as
(4)
S, =f (-~ )dT
and the value of ST accompanied with the phase change is determined as S/ = Nk~ ln(2J + i)
(5)
As the J dependence of (c?Ss) have scarcei.~. observed expenmentall~, we must c()nsider it
3T
using eqs.(3)--(5). As the result. It may be concluded that the compounds with large spin density, that is, the compound with the large populauon of heavy rare earth ions (large spin ion) are the most promising for the regenerator matenals. From these point ()f view, several investigating groups have found some hopeful rare earth compounds. "
8
,.-)
"-"
"~
=o
...a 0
"
'
"
./
'
"
Cv=234Nks(T/0 D)3"=CtT 3 /'
oo=lOOK/, //'-- 150K ./
2~0 10 Temperature (K)
/"
/
/'
/."'
.
.""
/
]
30
Figure 1 Temperature dependence of the specific heat of the lattice in low, temperature range. '3 The full line and dotted line are the results calculated using eqs.(1) and (_),respectix ely.
ICEC 16/ICMC Proceedings
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Layer structural regenerator In Fig. 2 are shown the temperature dependences of the specific heat in several magnetic materials near Tc. For comparison the specific heat of the pressurized He-gas and lead metal, which is the representative example of a normal metal,are also shown. The peaks of the specific heat Cv in the magnetic compounds are very sharp and it is supposed that specific heat (_,k,of the pressurized He-gas can not be covered with only one kind of magnetic materials. In order to ascertain the above fact we have shown the reduced hall width (AT~To) of several magnetic materials in table I. This table is clearly shows that the (AT~To) is -( ._~ in contrast with the (AT/Tc)=,-1.O in the He-gas and this fact means that more than three kinds of magnetic matenals with differant Tc in the range from 4K to - 1 5 K are necessary. On the basis of this fact two and three layer magnetic regenerator have been investigated[3]. In this paper, in order to make clear qualitatively the effectiveness of layer structural regenerator, the computer simulation on efficiencies in several types of layer structural regenerators will be performed. Finally,the most probable structure of two or three layer structural regenerators will be produced and their capacities will be compared experimentally. COMPUTER SIMULATION OF EFFECTIVENESS REGENERATORS
OF THE MULTI-LAYER TYPE
First, in order to obtain useful information on the most suitable structure in the double layer type regenerator, the efficiencies of several types of double layer regenerators have been estimated by computer simulation[4]. In simulation, the size of the regenerator was assumed to be 28mm in diameter and 85mm in length, which is about half the length of actual apparatus used in experiment. In operating time of the regenerator the time differential of enthalpies of gas andregenerator materials,
OHg, 8t
OHs
and - - - - , are guided from the concervation law of energy, the concervation law of mass 8/
~
,
,
& o A o
~'~'o.8 ,_,
! -oo
9 "'0.6
~o B / ~i o & \Z
...'; Er3Ni
. ..,..:,."~"
.i"
o
OoO"
000~
(D
i O.
"
.
toO.2
Oo~r o
ao~
""
...'"Pb
~176176 0
0
_.,,""
I
I0
He(10atm~ ,
I
20
,
30
Figure 2 Temperature dependence of the specific heat of several magnetic materials. comparison specific heats of lead metal and He-gas are also shown.
For
Table I Hall width of the peaks of specific heat in several magnetic matena!s and pressurized He-gas[3]. compound EuS ErNi~ DyNi~
T~ (K)
•,T
,AT/T,.
16.3 6 20.5
5.7 3.5 7.0
0.35 0.58 0.35
E r 0 9Yb0. i Ni
9
3.2
0.35
He ( l O a t o m s . )
7.9
7. 1
0.90
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ICEC16/ICMC Proceedings
and equation of thermal transmission as
( -~t ) - ( h A w L / m ) ( T s - T g ) Oils ( "7 ) - (hAwLxJMs)(Tg-rs)
(6)
,
(7)
,
Tg: temperature of gas Hg: enthalpy of gas h : heat transfer coefficient Ms: mass of regenerator matrix
Ts: temperature of regenerator matrix Hs: enthalpy of matrix Aw: heat transfer area per unit length L : regenerator length
m : mass flow rate of the gas
z: half cycle period t : reduced time (t/-c).
x : reduced length of the regenerator(x/L)
The efficiency of the regenerator is defined as the efficiency of the enthalpy (q) and, therefore, q is given by
B - AHreal/AHideal,
(8)
where AHreal and AHideal are respectively the actual and ideal enthalpy changes of the gas. Using three kinds of compounds[5], Ero.9Ybo.lNi, Er3Ni and Er3Co, two kinds of double layer types regenerators, type A and type B, Type A" (Ero.9Ybo.lNi) 1-x(Er3Co)x Type B (Ero.9Ybo. l Ni) 1-x(Er3Ni)x are considered in the computer simulation. The heat capacities of the regenerator matenals not only below 10K but also that in ~10K~20K range give the large effect on the efficiency of the regenerator. Therefore,noticing the latter case where the heat capacity of (Ero.9Yla).lNi) becomes very small, we selected two kinds of meterials, Er3C(~ and Er3Ni which have the different kinds of temperature dependences (3I heat capacities in this range. The condition in computer simulations were as follows: (1) the temperature at the top of the regenerator TH and bottom of regenerator TL were 30K and 4K, respectively; (2) the flow rate of He gas was 2(gr./sec.)" (3) the pressures in the high pressure process PH and that in low pressure PL were 20 and 8 atom., respectively. Figure 3 shows the temperature dependence of specific heats of Ero.9Ybo.lNi, Er3Ni and Er3Co. Using these results, the temperature gradients in two types of double layer regenerators have been calculated as shown in Fig.4. The two kinds of lines at high temperature and low temperature side lines in Fig.4 correspond to the heat absorbing(in high pressure) and the heat releasing(in low pressure) processes in one refrigeration cycle, respectively. Though this oscillation of the temperature of regenerator materials is suggested to be insufficiency of the heat capacity for the regenerator materials, it may be considered that these materials is possible to make regenerative operation as shown late,. However, the most suitable distribution of the specific heat in the regenerator can not yet be undersux)d qualitatively. A typical example of simulation results is shown in Fig.4, which shows the efficiency of the two kinds of regenerators as a function of the ratio of Ero.9Ybo. 1Ni to Er3Ni and Er3Co. This result clearly show that the complex type regenerator using Er3Co is more efficient than that using Er3Ni and also shows that the heat capacity in the range of-9K---15K is also important. On the basis of these results, the experimental apparatus,that is, the multi-layer structural regenerators were designed.
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E
N
I-
..r
//\1
I
Io L =
,Si../'''" /
//...'
L /~""
9:
E,o.,Ybo.,N
.//
0
Er3Co
......
10
20
the
Figure 3 Temperature dependence of materials, Ero.9Yb0. INi, Er3Ni and Er3Co. .
20 9
.
.
.
\
-
.,;.""" /
/ / I
!
9
30
specific . ! .
heats
of
several
candidates
9
Rege;erator m~terial i Er3Ni i
..._,..
::3
E i~
10,
0
Hot End
Position in Regenerator
1
Cold
End
Figure 4 Temperature gradient in the double layer type of regenerator. Two kinds line, high and low temperature side correspond to the heat absorbing and heat releasing process in a cycle. EXPERIMENTAL RESULTS AND CONCLUSION To verify the superiority of the multi-iayer type magnetic regenerator in the low temperature range, using the GM(Gifford-McMahon) refrigerator, the characteristics of the following four kinds of magnetic regenerators are compared[5,6]; (1) Er3Ni only, which is selected as the standard result,(2) Ero.9Yb0.1Ni+Er3Co, which is the representative regenerator of the double layer, (3) Er(Co0.2Nio.8)+Er0.9Ybo. 1Ni+Er3Co,and (4) Ero 9Yb(~ 1Ni+Er;C~+Er0 "75Gd(~.zsNi. Especially, (3) and (4) are Typical three layer type regenerator. In the experiment, so as. to tell clearly the difference of the characters in these regenerators ,these are applied to 2nd regenerator in a usual GM refrigerator and their refrigeration power are compared with each other under the same opperating condition. The structures of four layer structural regenerators used for the experiment are shown in Table !I. In Table III the refrigeration powers at 4.2 K are shown as a function of refrigeration cycle and these result clearly show; (1) the maximum refrigeration power are obtained from type (D) and the power increases as the layer increases, (2)From the comparison of type (C) and type (D),the heat capacity at ~20-10 range seems to be also important character to decide the refrigeration power. Conclusively, in order to increase the refrigeration power we have to use the multi-layer type regenerator. However, the considerations mentioned above are quantitative conclusion and so as to obtain the largest refrigeration power, we have to made clear the most hopeful quantitative information on the distribution of heat capacity in the regenerator.
330
ICEC16/ICMC Proceedings 100 -~
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(A) (B) (C) (D)
The construction of regenerator EraNi only Ero. ~Ybo. INi(50%)+EraCo(50%) Er(Coo. 2Nio. s)(25%)+Ero..~Ybo. INi (25%) +Era Co (50%) Ero. 9Ybo ,Ni (50%) +Era Co (25%) + Ero. 7sGdo. zsNi(25%)..
Table III The refngeration capacities of the GM refngerators with above four types of regenerators at 4.2K as a function of reciprocating speed. reciprocating refrig.power refrig.power refrig.power refrig.power
speed of type(A) of type(B) of type(C) of type(D)
24 0.88 0.96 1.0
. .30 . . . . 36 42 0.66 ( w a t t . ) 0.96 1.05 1.03 1.10 I. 12 1.17 1.17
REFERENCE 1.
2. 3. 4. 5. 6.
For instance, Hashimoto, T., Li ,R., Matsumoto, K., Sahashi, M., Y ayama H., and Tomokiyo, A., Recent Progress in the Materials for Regenerator in the Range from 4.2K to 20K, Prec. Intern. Cryogenic Material Conf. p.667, 1988, Shengyan China. Nagao, M., Inaguchi, T., and Yoshimura, H., Helium Liquefaction by Gifford McMahon Cycle Cooler, Advances in Cryogenic Engineering 35(1990) 1183. Kuriyama, T., Hakamada,R., Nakagome, H., Tokai, Y., Sahashi, M., Li, R., Yoshida, O., Matsumoto, K., and Hashimoto, T., High Efficient Two-stage GM Refrigerator with Magnetic Material in the Liquid Helium Temperature Range, Advances in CD'ogenic Engineenng ( Plenum Press, New York ) 35(1990) 1261. Debye, P., Ann.Physik 39(I912)789. Hashimoto, T., et al.,Teion-Kougaku 29(1994)51. Seshake, H., et al.,Advances Cryogenic Engineering 37(1992)995. Yabuki, M., et al., Recent Progress on Application of High Entropy magnetic matenals to the regenerator in He Temperature Range, Proc.7th Intern.Cryocooler, p.605,Kartland AIZB NM,(1993). Tsukagoshi, T., et al., to be published to Cryogenics in 1996.
Development of a 4K GM/JT Refrigerator for Maglev Vehicle
Satoru Fujimoto, Shoichi Taneya, Toshiyuki Kurihara, Katsuya Miura, Keiji Tomioka Masakazu Okamoto, Tatsuya Yamaguchi, Shinichi Kasahara Motoaki Terai*, Akihiko Miura*, Hiroyuki Nakao**, Takahiro. Fujinami**, Yasutoshi Nakamoto*** Daikin Industries Ltd, 1304 Kanaoka-cho,Sakai,Osaka,Japan *Central Japan Railway Company, 1-6-6 Yaesu,Chuo-ku,Tokyo 185,Japan **Toshiba Corporation, 1 Toshiba-cho,Fuchu,Tokyo 185,Japan ***Toshiba Corporation, 1-1-1 Shibaura,Minato-ku,Tokyo 105-01,Japan
INTRODUCTION In our project, a Gifford-McMahon/Joule-Thomson(GM/JT) refrigerator for Maglev vehicle has been developed in partnership with Central Japan Railway Company and Toshiba Corporation. Commercial GM/JT refrigerators had been used for radio astronomy and other laboratory systems, however they had not been applied to Maglev vehicle because of the small refrigeration capacity due to the low efficiency. On-board helium refrigerators require large capacity, high efficiency and reliability even under severe environmental conditions. Major efforts have been made to improve the refrigeration capacity and efficiency with considering reliabilities.
SPECIFICATIONS Principal Specifications of an on-board refrigerator are listed in Tablel. The refrigeration efficiency had to be increased approximately 80%. In addition, the refrigerator for Maglev vehicle has to be operated under very restricted environment. The constraints for designing the refrigerator are listed below. (1) The refrigerator has to be stored in a limited space and it can't exceed the certain weight limit (2) The power consumption of the refrigerator is restricted. (3) The refrigerator has to withstand the vehicle vibration, magnetic field and variable ambient temperature. (4) The refrigerator has to deal with the excessive amount of heat load momentarily when the superconducting magnet is energized or de-energized Table l Principal specifications of helium refrigerator Refrigeration capacity Input power COP Vibration condition Magnetic field condition Weight
for Maglev 8W (at 4.4K) 8kW 1.0X10-3 exp:-4-10G, comp :4- 5G max 0.1Tesla < 300kg
commercial (CG308) 3.5W (at 4.3K) 6.4kW 5.5X10-4
255kg
STRUCTURE OF REFRIGERATOR The refrigerator consists of an expander unit, a compressor unit and a flexible tube that connects these units. Fig.1 Shows the flow diagram of refrigerator. 331
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ICEC16/ICMC Proceedings
The structure of the expander unit, which is installed on a side of the helium tank on top of the superconducting magnet, is shown in Fig.2. It has a 2-stage GM expander(precooler), three JTheat exchangers, a JT valve, and a radiation shield which will be cooled by 1st stage of G M expander. The displacer inside of expander is driven pneumatically at 2Hz by switching high/low pressure with rotary valve. The major components of the compressor unit equipped on bogie frame has two scroll compressors, oil separators, an adosorber, and an air-cooler for cooling helium gas/oil. They are fixed inside of the aluminum frame like Fig.3. These compressors have different scroll forms which were improved from what have been used for air conditioning. From this fact, it is obvious that the compressors hold high reliability and cost performance.
DEVELOPMENT OF REFRIGERATOR There were mainly three aspects to improve refrigeration capacity and efficiency. (1) Optimizing the pressure condition in refrigeration system (2) Improvement of compressor (3) Improvement of GM expander
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ICEC16/ICMC Proceedings
333
(1) Optimizing the pressure condition It is important to match the compressor's property with the expander's. Because of the fact that the camot efficiency of GM expander changes with pressure since refrigeration occurs in principle of Simon expansion, not adiabatically, it is necessary to design the refrigerator by optimizing the pressure condition. Fig.4 shows the input consumption with constant efficiency of compressors and expander. The pressure condition is determined by intermediate pressure because high and low pressure are fixed by the control valves. When the intermediate pressure is high, the pressure ratio of second compressor becomes small and the input consumption of second compressor will decrease. In the system used, the intermediate pressure is set at 8-9kgf/cmZG to optimize the properties of compressors and expander. In order to adjust this pressure, the expansion volume had to be extended, and suck-in volume of the 2nd compressor had to be enlarged. (2) Improvement of compressor Scroll compressor used for the project is called "fixed compression ratio" type, and its best compression ratio depends on the number of scrolls. If the actual ratio changes, the compression efficiency will decrease because of over-compression or re-compression. The scroll are optimized to get best efficiency as described in the section of "Optimizing the pressure condition" above. As a result, the compression efficiency increased approximately 10% on first compression and 15% on second compression. (3) Improvement of GM expander GM expanders with new regenerator using rare-earth materials such as Er and Ho, have been developing and their capacity and efficiency have shown tremendous improvement in recent years. The specific heat of lead which is a conventional regenerator becomes smaller when temperature is lower than around 15K. On the other hand, new regenerators have an outstanding peak of specific heat between 10K to 20K that is mainly caused by magnetic phase transition. Therefore, Er3Co compound is adopted for the second stage regenerator of GM expander. Er3Co has high specific heat property within the range of 2nd stage temperature. The quality control when they are on manufacturing process is relatively high, and finished product has rigidity. In the range of second stage temperature, 55% of the capacity shown on the P-V diagram is lost and the rest is remained the refrigeration capacity. Some of these losses are caused by shuttle, pumping and so forth, however the largest loss comes from the inefficiency of regenerator. It is estimated that 20% of the regenerator loss can be regained by utilizing Er3Co (Fig.5). Furthermore, the computed simulator that can calculate P-V diagram and some losses of GM expander has been developed. By using simulator, some parameters to design the refrigerator are determined such as displacer stroke, regenerator quantity, and the compound size.
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Fig.5 2nd-stage Capacity
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ICEC16/ICMC Proceedings
CHARACTERISTICS OF REFRIGERATOR (1) Frequency-Capacity property The refrigeration capacity is 8.3W when the input consumption is 7.9kW, and the COP is 1.05x10 -3 at 53Hz. The temperature in helium tank is below 4.5K, and the JT mass flow is 12.512.7Nm3/h. However the capacity without vibration-proof shows 8.5-9.2 watt and the COP i s 1.1x10 -3 at 50-55Hz. This difference of capacity is conductive loss caused by supporting equipment which fixed some elements. COP will decrease at over 55Hz because the capacity of precooler doesn't increase in proportion to JT mass flow. (2) Gas Withdrawal and Load-changing property n| The refrigerators withdraw the helium gas at t 60W/2min" 60W/2mi 70Hz to control the helium tank pressure when ~ H~II Ile~ld superconducting magnets are energized or de, i ! 6W . . . . o!| e~.ergized~ AI].d the refrigerators restrain the i~Iput mode ' ~~.~.~ ~-i ~ ~ ~~.~" consumption and protect the pressure drop in ' [~()~ual(53Hzi , Ehergize helium tank at 40Hz when the heat load is lower En~gize(70Nz) [--]Restraint(z~(JHz) D ~ -i(,TOHz-) or Maglev vehicle stops for more than a certain ._. period of time. The Refrigerators are ordinally ~ 0.:3 I r----g operated in "usual mode" at 53Hz, and can be =~ 0.2 switched from "usual mode" to "restraint mode ~ 0.1 by detecting buffer tank pressure. Fig.6 shows the result of cycle mode test that .~ 15 has two energizing/de-energizing operations each E 10 in a day and switching automatically between "usual mode" and "restraint mode". Pressure in ',,' 5 helium tank is controlled below 0.39kgf/cm2G at 0 0 6 12 18 energizing/de-energizing and 0.2-0.3kgf/cmZG at Time(hour) the other modes.
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has the refrigeration capacity rating under vibration, magnetic field and refrigeration systems are performed to improve the reliability and the
ACKNOWLEDGEMENT This work is supported by Central Japan Railway Company and Toshiba Corporation.
REFERENCES 1) Fujimoto S., Taneya S. and Tamtani I., Development of a GM refrigerator for Maglev(1) 47th Meeting on Cryogenics and Superconductivity (1992) 2) Fujimoto S., Taneya S. and Kurihara T., Development of a GM refrigerator tbr Maglev(2) 49th Meeting on Cryogenics and Superconductivity (1993) 3) Terai M., Inadama S., Tsuchishima H., Suzuki E. and Okai T., Development of new superconducting magnets for the Yamanashi Test Line Maglev'95 14th International Conference on Magnetically Levita.te~ Systems Bremen'Germany (1995) 267-273
Development of 2W Class 4K Gifford-McMahon Cycle Cryocooler
Inaguchi Takashi, Nagao Masashi, Naka Kouki, Yoshimura Hideto Advanced Technology R&D Center, Mitsubishi Electric Corporation, Tsukaguchi-Honmachi 8-chome, Amagasaki, Hyogo,661 Japan
This paper describes the principal design features and performance of the GiffordMcMahon cycle cryocooler by which we could obtain the cooling capacity of 2.2W at 4.2K. The main features of this machine are its large size expansion space, the use of a regenerator packed rectifiable meshes, and the combinational use of Er3Ni and ErNio.gCo0a as regenerator materials.
INTRODUCTION Since the success of helium liquefaction by the Gifford-McMahon cycle cryocoolcr(hereinafter called 4K GM cryocoolcr)in 1989[1], the 4K GM cryocoolcr has been given attention as a cryocooler which will replace the JT cryocooler because the 4K GM cryocoolcr has high reliability and very easy handling[2]. In particular, 4K GM cryocoolcrs actively have been applied to superconducting magnets such as helium free superconducting magnets[3,4,5] and MRI magnets[6] which don't have to supply liquid helium. However, the use of the 4K GM cryocooler has been limited to small magnets or magnets whose heat load is small. In order to further extend the range of applications of the 4K GM cryocooler, the cooling capacity of the 4K GM cryocoolcr must be improved. During this experimental period we obtained a cooling capacity of 2.2W at 4.2K. The main features allowing us to achieve this result are the machine's large size expansion space, the use of a regenerator packed rectifiable meshes, and the combinational use of Er3Ni[7] and ErNio.gCo0.~ [8]as regenerator materials. In this paper we report on the experimental apparatus used and the results. EXPERIMENTAL APPARATUS Figure 1 shows a schematic of the expander of a two-stage 4K GM cryocoolcr which is an experimental machine. The 4K GM cryocooler has two cooling stages through which heat loads are transferred into the cylinder. Displacers which contain regenerators reciprocate in the cylinder. By the displacers and the cylinder two expansion spaces and a room temperature space are formed. The first regenerator has a twolayer structure. We stacked phosphor bronze screens in the high temperature part and lead shots in the low temperature part. In the second regenerator we packed Er3Ni, or the combination of Er3Ni and ErNio.9Co0.1 as regenerator material, then the effect on cooling capacity was investigated. Figure 2 shows a schematic of the experimental apparatus. The cooling stage of the 4K GM cryocooler was installed in a vacuum chamber. A radiation shield was attached to the first stage, and it enclosed the second stage to prevent the radiant heat from room temperature from entering the second stage. The temperature of the first stage was measured using a Pt-Co resistance sensor and the temperature of the second stage was measured using a carbon glass resistance sensor. A cartridge heater was installed at each stage to measure cooling capacity. In order to secure enough flux to the expander, two compressors (each with rated input power of 6kW) were arranged in a row. For this experiment we kept the differential pressure of the room temperature space at 1.2MPa"-1.3MPa by adjusting a bypass valve which was installed between the high pressure pipe and the low pressure pipe. A pressure transducer was installed into the room temperature space of the expander and a linear 335
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ICEC16/ICMC Proceedings Or
Pressure transducer
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Figure I Schematic of expander of two-stage 4K GM cryocooler.
displacement converter was set up on the drive mechanism of the displacers in order to measure the indicated work. Pressure transducers were also installed at the gas exit and entrance of the expander, and inflowing or outflowing gas pressure was measured. EXPERIMENTAL RESULTS Effect of Volume of Expansion Space on Cooling Capacity Table 1 shows the main parameters of the 4K GM cryocoolers which were used in this experiment. We changed the diameters of the second expansion space from 25.4mm to 60mm in order to investigate the effect of volume of expansion space on cooling capacity. The diameter of the first expansion space of #1 through #3 were 61.9mm; that of #4 was 70mm; and that of #5 was 90mm. The stroke of the cryocoolers were all 32mm. We employed Er3Ni as the regenerator material. The cycle frequency employed was the cycle frequency at which cooling capacity at 4.2K became optimum. The optimum cycle frequency was 45rpm in the cases of # 1 ~ # 3 and the optimum cycle frequency was 33rpm in the cases of # 4 " # 5 . Figure 3 shows the effect of volume of the expansion space on cooling capacity at 4.2K. When the volume of the expansion space is 16.1cm 2 (diameter: 25.4mm), the cooling capacity is 0.34W. The cooling capacities increase in accordance with the increase of volume of the expansion space. When the volume of the expansion space is 90.5cm 2 (diameter: 60mm), the cooling capacity becomes 1.4W. The optimum cycle frequency decreases from 45rpm to 33rpm in accordance with the increase of the volume of the expansion space. The rate of increase of cooling capacity is not therefore proportional to the rate of increase of volume.
Table 1 Main parameters of 4K GM cryocooler Diameter of Diameter of Stroke Optimum cycle 2nd expansion 1st expansion (mm) frequency(rpm) space (mm) space (ram) #1 #2
25.4 30.2
61.9
32
45
61.9
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ICEC16/ICMC Proceedings
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/--- Rectifiable meshes Felt mat
Rectifiable meshes Regeneral 9 Er3Ni
Felt mat Felt mats
-'--- Regenerator material 9 Er3Ni Rectifiable meshes
(a)
(b)
Figure 4 (a) Schematic of conventional regenerator and (b) schematic of regenerator packed rectifiable meshes in two additional places of regenerator at .equal intervals Effect of Regenerator Packed Rectifiable Meshes on Cooling Capacity We used two kinds of regenerators shown in Figure 4(a) and 4(b) and investigated the effect of cycle frequency on no load temperature. Figure 4(a) shows a conventional regenerator. Rectifiable meshes are packed only at the ends of the regenerator. Figure 4(b) shows a regenerator packed rectifiable meshes not only at the ends of the regenerator, but also in two additional places of the regenerator at equal intervals. The rectifiable meshes were put between two felt mats to prevent regenerator material from blocking meshes. In both cases Er3Ni was employed as the regenerator material. The experimental machine employed was the 4K GM cryocooler shown in #5 of Table 1. Figure 5 shows the effect of cycle frequency on no load temperature. When we employed the regenerator in Figure 4(a), the no load temperature was 2.78K at the cycle frequency of 30 rpm and it was 4.47 K at the cycle frequency of 45 rpm. The temperature rise was 1.68K. We considered the cause of temperature rise would be that helium gas in the expander flows on one side of the regenerator in the case of the large size 4K GM cryocooler like #5 of Table 1. When we employed the regenerator in Figure 4(b), the no load temperature was 2.63K at the cycle frequency of 30rpm and it was 2.81K at the cycle frequency of 45 rpm. The temperature rise was only 0.18K and it is considered that the rectifiable meshes in two additional places of the regenerator could prevent helium gas in the expander from flowing on one side of the regenerator. Figure 6 shows the cooling capacity of the 4K GM cryocooler using the regenerators shown in Figure 4(a) and 4(b). When we employed the regenerator shown in Figure 4(a), the optimum frequency was 33rpm and the cooling capacity at 4.2K was 1.4W. When we employed the regenerator shown in Figure 4(b), the optimum frequency was 42rpm and the cooling capacity at 4.2K was 1.8W. The optimum cycle frequency of the 4K GM cryocooler using the regenerator shown in Figure 4(b) was 1.3 times higher than that of the 4K GM cryocooler using the regenerator shown in Figure 4(a), and the cooling capacity at 4.2K in the case of Figure 4(b) also improved to 1.3 times greater than that in the case of Figure 4(a).
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338
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Figure 8 (a) Cooling capacity of 4K GM cryocooler using only Er3Ni and (b) cooling capacity of 4K GM cryocooler using combination of Er3Ni and ErNio.9Co0.a Effect of Regenerator Material on Cooling Capacity The effect of specific heat of regenerator material on cooling capacity was investigated, using Er3Ni and ErNiogCoo i as regenerator materials. Figure 7 shows specific heat of the regenerator materials. Two cases were examined. In case one, experiments were carried out using a regenerator packed only Er3Ni. The other case used a regenerator packed Er3Ni in two thirds of the high temperature side of the regenerator and ErNiogCoo 1 in one third of the low temperature side of the regenerator. In both cases rectifiable meshes were packed in the regenerator as shown in Figure 4(b). The experimental machine employed was the 4K GM cryocooler shown in #5 of Table 1. Figure 8 shows the cooling capacity of the 4K GM cryocooler using only Er3Ni and the cooling capacity of the 4K GM cryocooler using combination of Er3Ni and ErNiogCooa. The cycle frequency was fixed at 42rpm. When only Er3Ni was employed, the cooling capacity was 1.8W, and when the combination of Er3Ni and ErNiogCoo i was employed, the cooling capacity was 2.2W. The cooling capacity of the 4K GM cryocooler using the combination of Er3Ni and ErNiogCoo 1was 1.2 times greater than that of the GM cryocooler using only Er3Ni. CONCLUSIONS ,.,.,..W (1) The cooling capacity of "~ '~ at 4.2K was obtained by the 4K GM cryocoolcr. The main features of this machine are its large size expansion space, the use of the regenerator packed rectifiable meshes, and the combinational use of Er3Ni and ErNio.gCoo.1 as regenerator materials. (2) A regenerator packed rectifiable meshes in two additional places of the regenerator at equal intervals can improve cooling capacity 1.3 times greater than a regenerator packed rectifiable meshes only at the ends of the regenerator. (3) The cooling capacity of the 4K GM cryocooler using the combination of Er3Ni and ErNio.gCoo.1 can be improved 1.2 times greater than that of the 4K GM cryocoolcr using only Er3Ni. REFERENCES 1. 2. 3. 4. 5. 6. 7. 8.
Yoshimura, H.,et al.,Rev.Sci.Instrum. 60(1989) 3533-3536 Inaguchi,T.,et al.,Cryocooler 6 (1990)25-36 Watanabe,K.,et al.,Cryogenics 34,ICEC Suppl. (1994) 639-642 Kuriyama,T.,et al,Cryogenics 34,ICEC Suppl. (1994) 643-646 Yokoyama,S.,et al, MT-14(1996)in press Nagao,M.,et al.,Adv.Cryog.Eng. 39(1994) 1327-1334 Tokai.,Y.,et al.,Jpn.J.Appl.Phys.Partl_ 31 (1992) 3332-3335 Onishi,A.,et al.,Cryogenic Engineering 31 (1996) 162-167
Optimization of Intake and Exhaust Valves for 4 K Gifford-McMahon Cryocooler
Rui Li, Atsushi Onishi, Toshimi Satoh, Yoshiaki Kanazawa R & D Center, Sumitomo Heavy Industries, Ltd., 63-30, Yuhigaoka, Hiratsuka, Kanagawa, 254, JAPAN
The influence of intake/exhaust valve timing on performance of 4 K GM cryocooler has been investigated at various cooling temperatures in order to optimize the intake and exhaust valves. The 4 K G M cryocooler employed for this study is a standard two-stage type one with a rotary valve for intake and exhaust. The second stage regenerator in the cryocooler has both lead spheres and spherical magnetic regenerator material, ErNi0.gCo0.,, with the latter arranged at the colder end. The cooling capacities of the first stage and the second stage were measured in wide temperature ranges (-< 80 K for the first stage, - < 1 0 K for the second stage). The results indicate that the optimum intake/exhaust valve timing is dependent on the cooling temperature and is possible to be different for each stage.
INTRODUCTION In the last several years, regenerative cycle cryocoolers, such as G ifford-McMahon (GM) cryocooler especially, developed remarkably because of the successful use of magnetic regenerator materials. A substantial body of literature on applying magnetic regenerator materials to the cryocoolers has been published[l], but the papers describing the improvement in the components other than regenerator material are very few[2,3]. As a matter of fact, 4 K GM cryocoolers are developed from commercially available GM cryocoolers, and the optimization of the other components for the cooling at 4 K is as important as the advances in regenerator materials. Intake and exhaust valves are key components of GM cryocooler, and include several important factors, such as valve type, valve structure and valve timing. In the previous researches[4,5], we investigated the influence of valve timing on performance of 4 K GM cryocooler, and reported that a reasonably early open timing of intake/exhaust valve not only brings about a cooling capacity more than 1 W at 4.2 K on the second stage but also produces a much large cooling capacity at the first stage. References of [4] and [5], however, mainly focused on the cooling capacities at 40 K for the first stage and at 4.2 K for the second stage. In order to optimize the intake and exhaust valves at various cooling temperatures for 4 K GM cryocooler, we measured the cooling capacities of both the stages in wide temperature ranges ( - < 80 K for the first stage, =<10 K for the second stage) in the present investigation. This paper describes the experimental details and the influence of intake/exhaust valve timing on 4 K GM cryocooler performance at various cooling temperatures.
EXPERIMENTAL DETAILS The experimental apparatus used in this work is the same as that described in the previous papers[4,5]. Figure 1 is the schematic of the experimental setup. Both the cold head and the compressor unit are manufactured by Sumitomo Heavy Industries Limited. The first regenerator consists of copper meshes and lead spheres. The second regenerator is also a hybrid one with lead spheres of 360 g placed at the 339
340
ICEC16/ICMC Proceedings Laser Displacement Sensor
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hotter end and spherical magnetic regenerator material, ErNi0 9Co0 ~, of 360 g arranged at the colder end. In order to describe the valve timing perfectly, both the valve open timing and the valve close timing should be used in a general way. In this paper, however, we take the valve interval (at which the valve is open) instead of the valve close timing. Since a rotary type valve is employed for intake/exhaust in our 4 K G M cryocooler, it is preferable to represent the valve open timing and the valve interval by an angular measure as shown in figure 2. The figure illustrates the valve which intake/exhaust open timing is -420/-46 ~ and the intake/exhaust interval is 124~ ~. We define the position at which the expansion volume turns to minimum/maximum as 0~ ~ for intake/exhaust. The open timing of-420/-46 ~ therefore, means that the intake/exhaust valve opens at the angle (--420/-46 ~) before the expansion volume reaches minimum/maximum. Because of the angular expression, the valve open timing will be often designated as valve open angle as the following. According to the results of our previous investigation, we prepared five valves for this work. The interval of these valves is the same (-124~176 and their open angles are from -220/-26 ~ to -620/66 ~
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~ and an interval of
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~
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Table I Comparison of cooling capacity between the 4 K GM cryocooler and a conventional GM cryocooler SRD-210L4 (4 K GM for this work)
SRD-210 (Conventional GM) Maximum second stage cooling capacity
Maximum first stage cooling capacity
o w (@ ~8 K) ~1 W (@lO K)
I. 16 W (@4.2 K) 5.41 W (@7 K) 10.9 W (@ 10 K)
10 ,-, 15 W (@40 K) 58 ~ 70 W (@77 K)
39.1 W (@40 K) 87.8 W (@80 K)
* The cryocoolers compared above are almost the same in size and are manufactured by Sumitomo Heavy Industries Limited.
RESULTS AND DISCUSSIONS For each valve, we measured the load map of the 4 K G M cryocooler in the temperature range of =<80 K for the first stage and < - 1 0 K for the second stage. Figure 3 shows an example of the load maps. In general, the horizontal axis and the vertical axis of the load map represent the first stage and the second stage temperature respectively. It is, however, desired to notice that we took the cooling capacities of the first stage and the second stage as the horizontal axis and the vertical axis in Fig. 3. Figure 4 has the same style as Fig. 3, but gives the details about the cooling capacity at 4.2 K for the same valve. The maximum cooling capacity of the second stage at 4.2 K is 1.16 W, and the cooling capacity keeps its value being about 1 W even though the first stage is heated up to 50 K. Furthermore, the cryocooler employed in this work also has a good performance at higher temperatures. The second stage delivers a maximum cooling capacity of 10.9 W at 10 K, and the first stage has a maximum cooling capacity of 87.8 W at 80 K. Table 1 gives the comparison of the cooling capacity between the 4 K GM cryocooler and a conventional GM cryocooler. It is clear that the 4 K GM cryocooler has a good performance at 4.2 K, but also is superior in cooling capacity though the stages are warmed up to 80 K for the first stage and up to 10 K for the second stage. Figure 5 shows the influence of valve open angle on the first stage cooling capacity at various cooling temperatures. For a constant first stage temperature, there is an optimum valve open angle at
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4
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I -62/-66
Open Angle for Intake/Exhaust (deg.) Figure 6 Secondstage cooling capacity as a function of valve open angle at various cooling temperatures
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which the cooling capacity is maximum. When the first stage is heated from 40 K up to 80 K, the optimum valve open angle shifts from -520/-56 ~ to-420/-46 ~. According to the definition of valve open angle in Fig. 2, the smaller absolute value in the horizontal axis of Fig. 5 means that the intake/exhaust valve opens closer to the position at which the expansion volume turns to minimum/maximum (=070~ Since the intake/exhaust valve opens exactly at the angle of 0~ ~ in the ideal GM cycle, the Fig. 5 indicates that the optimum valve open angle is closer to the condition of file ideal cycle when the cooling temperature is higher. At a higher temperature, the helium gas in the cold head works more like the ideal gas, therefore, the optimum valve open angle naturally shifts to that of the ideal GM cycle. For the second stage, however, the valve open angle dependence of the cooling capacity is different from Fig. 5. Figure 6 gives the second cooling capacity as a function of the intake/exhaust valve open angle. The optimum valve open angle (from -420/-46 ~ to -520/-56 ~) becomes away from the ideal G M cycle condition when the first stage is heated from 40 K up to 80 K. The result is just opposite to that in Fig. 5. The reason is not understood clearly, but we consider that is concerned with the mass distribution of the working gas in the first stage and the second stage.
CONCLUSIONS The influence of intake/exhaust valve timing on performance of 4 K GM cryocooler has been investigated at various cooling temperatures. The following conclusions were drawn from this work. (1) The optimum intake/exhaust valve timing is dependent on the cooling temperature, and is possible to be different between the first stage and the second stage. In other words, the cooling temperature is important for the optimization of the intake and exhaust valves. (2) The 4 K GM cryocooler delivers great cooling capacities not only at 4.2 K but also at much higher temperatures.
REFERENCES Hashimoto, T., Li, R., Kuriyama, T. and Nagao, M., Recent progress in application of magnetic regenerator material, Cryogenic Engineering(Journal of the Cryogenic Society of Japan) (1996) 31 131-149 Kuriyama, T., Takahashi, M., Nakagome, H., Hashimoto, T., Eda, T. and Yabuki, M., Regenerator performance and refrigeration mechanism for 4 K GM refrigerator using rare earth compound regenerator materials, In: Proceedings of 7th International Cryocooler Conference, Air Force Phillips Laboratol 3, Report PL-CP-93-1001, Kirtland AFB, NM (1993) 429-443 Plambeck, R. L., Improved seal for a 4 K Gifford-McMahon cryocooler, In" Cryocoolers 8, Ed. by Ross, R. G., Jr., Plenum Press, New York (1995) 795-801 Li, R., Onishi, A., Satoh, T. and Kanazawa, Y., Influence of valve open timing and interval on performance of 4 K Gifford-McMahon cycle cryocooler, Adv. Cryog. Eng,, Plenum Press, New York (1996) 41 in press Li, R., Onishi, A., Satoh, T. and Kanazawa, Y., Optimization of intake and exhaust valves for 4KGM cryocooler, Cryogenic Engineering (Journal of the Cryogenic Society of Japan) (1996) 31 172181
Investigation of the Performance of a 4.2K G-M Refrigerator
Xiangdong Xu, Linghui Gong, Zhiyong Zhang and Liang Zhang Cryogenic Laboratory, Chinese Academy of Sciences, P.O.Box 2711, Beijing, China 100080 The pressure curves vs time of the 2nd stage cold chamber, admission line and exhaust line of a 4.2K G-M refrigerator were presented and the P-V diagram efficiency (real PV work / maximum PV work) of the cold chamber and the resistance of the helium passage inside the refrigerator at different temperature were obtained. The cause of the change of the P-V diagram efficiency and resistance along with the 2nd stage cooling temperature had been analyzed. We compared the P-V diagram efficiency with the ideal and real cooling capacities and pointed out that the increased P-V diagram area at low temperature ( < 8K or so) doesn't represent the net increased cooling capacity.
INTRODUCTION The performance of two-stage G-M refrigerators had made great progress since the discovery of magnetic regenerator matrix in the end of 1980s[1,2]. Our research group designed and tested a two stage G-M refrigerator in 1993 and obtained 3.0K no-load temperature and 0.515W/4.2K cooling capacity[3]. In order to understand in more detail the working process of the refrigerator, we installed the pressure and displacement transducers in this refrigerator and measured the pressure of the 2nd stage cold chamber, admission line and exhaust line. We obtained the P-V diagrams of the cold chamber and the resistance of the helium passage inside the refrigerator at different temperature. EXPERIMENTAL APPARATUS Two pressure transducers were installed in the admission and exhaust lines of the G-M refrigerator respectively. A displace transducer was connected with the driving bar to recorder the movement of the piston. Thinking of the difficulty of the direct measurement of the 2nd cold chamber pressure at low temperature, we drilled a hole into the end of the refrigerator and connected a room temperature pressure transducer to the cold end via a capillary tube(length 0.6m, inner diameter 1.0mm). The additional void volume and heat leakage introduced by the capillary tube had little influence on the performance of the refrigerator according to our analysis and literature[4]. All transducers were connected to a 386 computer by a A/D converter. Pressure and displacement signals were gathered and processed by the computer. Results were stored and shown on the screen like Fig. 1. EXPERIMENTAL RESULTS AND DISCUSSION Fig. 1 gave the pressure (the 2nd stage cold chamber, admission line and exhaust line) and displacement curves vs time and P-V diagrams of the cold chamber at different cooling temperature (Tc = 3.0, 30K). Fig.2a gave the P-V diagram efficiency curve vs cooling temperature T~. The P-V diagram efficiency,which expresses the ratio of real PV work to maximum PV work, is defined by"
l]PV=
PcoldAdS f(Ph-PI )A S 343
(1)
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ICEC 16/ICMC Proceedings
Where S is the displacement, A is the cross-section area of the cold chamber, Ph is the mean value of the pressure Pi, of the admission line, P, is the mean value of the pressure Pout of the exhaust line, P~o,d is the pressure of the cold chamber.
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Figure 1 Pressure and displacement curves vs time and P-V diagram We can see from Fig.2a that r/pv is almost constant for Tc> 100K and begin to gradually decrease for T~ < 100K. The rbv is still large until T~= 3K, while the real cooling capacity is zero.Fig.2b presents the ideal cooling capacity Q, and real cooling capacity Qox, curves vs Tc of this refrigerator. Comparing Fig.2a with Fig.2b we can know that P-V diagram area doesn't represent the cooling capacity for T~ < 8K or so. This is because the physical properties of helium are far away from that of ideal gas. Especially, even if the admission and exhaust temperature of the cold chamble is constant, the high pressure admission enthalpy is much larger than low pressure exhaust enthalpy. This will cause a large number of cooling capacity losses. In fact, the admission temperature is always higher than the exhaust temperature. Hence, the change of P-V diagram area will lead to the change of the admission and
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exhaust enthalpy. Thus, if we improve the performance of a refrigerator and make its P-V diagram area increase, we can't think that the increased P-V diagram area is the net increased cooling capacity. This situation is different from that of high T~ ( > 8K or so), when P-V diagram area does represent the cooling capacity and the increased P-V diagram area does be the net increased cooling capacity. Therefore, at low temperature the small progress of P-V diagram efficiency has very little contribution to the cooling capacity.
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Figure 3 Resistance of admission and exhaust process vs Tc Fig.3 presents the resistance ( P i . - Pco,d and Pco,d - Pou~) curves vs Tc of admission process and exhaust process respectively. Comparing Fig.2 with Fig.3, we can see that the increase of the resistance had little influence on the P-V diagram efficiency for % > 100K. This shows that the P-V diagrams efficiency is mainly controlled by the admission and exhaust angle for T~ > 100K. For Tc < 100K, the resistance increase rapidly along with the decrease of Tc. Especially for Tc < 20K this is more oblivious. Hence, the decrease of the P-V diagram efficiency mainly stems from the increase of the resistance of
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admission and exhaust for T~ < 100K. CONCLUSION 1.At low temperature( < 8K or so), the increased P-V diagram area doesn't represent the net increased cooling capacity. 2.The decrease of the P-V diagram efficiency mainly stems from the increase of the resistance of admission and exhaust for cooling temperature Tr < 100K. 3.The resistance of admission and exhaust increase very rapidly at low temperature( < 20K). ACKNOWLEDGEMENT This research was supported by the National Natural Science Foundation of China. REFERENCES 1.Inaguchi T. et al., Two-stage Gifford-McMahon Cycle Cryocooler Operating at about 2 K, Proc. 6th Int. Cryocoolers Conf. (1990) 2 25-36 2.Satoh T. et al., Development of 1.5W 4K G-M cryocooler with magnetic regenerator material, Presented at CEC/ICMC 95, Columbus OH, July 17-21.1995. 3.Zhang L. et al., Development and Analysis of a 4.2K G-M Refrigerator, Cryogenics (1994) 34 179-182 ICEC Supplement 4.Plambeck R. et al., A 4K Gifford-McMahon Refrigerator for Radio Astronomy, Proc. 7th Intl. Crvocoolers Conf. (1992) 401-415.
Effect of the Characteristics of the Gap Heat Exchanger to the Real Cooling Capacity of a 4.2K G-M Refrigerator
Linghui Gong, Xiangdong Xu and Liang Zhang Cryogenic Laboratory, Chinese Academy of Sciences, P.O. Box 2711, Beijing, China, 100080
The further numerical analysis of the gap heat exchanger of a 4.2K G-M refrigerator combined with the regenerator has been conducted in this paper in order to understand the process of heat transfer. It was found that the average number of transfer unit of the gap heat exchanger is enough for obtaining the satisfactory cooling capacity when NTU>0.8. Certainly, too small NTU (e.g.NTU<0.2) is disadvantageous.
INTRODUCTION In recent years, the performance of two-stage G-M refrigerators has been made significant progress as the rear earth magnetic materials are developed and applied[l]. While normally, no mention of the NTU of the gap heat exchanger(GHE) are included in papers on low temperature G-M refrigerator. We had found earlier that a low NTU of the GHE could be satisfactory to the performance of a G-M refrigerator[2]. Riedy[3] reported the temperature is down to 6.4K from 7.2K after improving of the structure of the GHE. We aim to find out the quantitative dependence of the performance of the refrigerator on the NTU of GHE. THE PHYSICAL MODEL A numerical calculation of the GHE combined with the regenerator has been done with a pressure variation of Fig.l, using following equations: In the GHE 0.0<x/L<0.025 a(azh) + ~ 4 r -_~ 4 j T,-~, _ _ -.-~, r ~+A --p_ a ( p h - P ) "0
Ox
(1)
at
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ax +A~ at 347
(2)
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ICEC16/ICMC Proceedings
In regenerator 0.025<x/L
IAfI (T,,._Tm)+AflO(ph-P) :0
a(mh) Ox
0
~
(3)
Ot
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_0
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Where Tf is fluid temperature, T m is matrix temperature, T~ is the temperature of cold head wall, x is the length dimension along GHE and regenerator, t is time, L is the total length of GHE and regenerator, h is enthalpy, m is mass flow rate, Af is flow area, a is heat transfer coefficient, A~ is heat transfer area per unit volume, Subscript 1 and 2 represent regenerator and GHE respectively. These two group equations are combined by continuous conditions of mass flow rate and temperature. At low temperature, a numerical solution of the above equations indicates that the NTU of the GHE has very great difference from that of normal heat exchanger. THE RESULT OF NUMERICAL CALCULATION The numerical calculation parameters are Ph=2.0MPa, P~=0.6MPa, Th=50.0K, Tc=4.2K. The temperature profile Tf, inside the GHE(0<x/L<0.025) and the cold part of the regenerator(0.025<x/L
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ANALYSES AND DISCUSSION The GHE is located between the cold chamber and the regenerator. So it has a closed relationship with the processes in the cold chamber and in the regenerator. Its heat transfer rate in GHE is analytically determined by:
Q=rhCe(Tc_Tf)e -NW
(6)
Because Tf is changeable, Q is not only function of NTU, but also that of Tf. The change of Tf can be seen from Fig.3, which presents the temperature profile vs time for two typical NTUs in GHE and regenerator. Comparing Fig.3a with Fig 3.b, we can see that the change of Tf for smaller NTU is much larger than that for larger one. Hence, Q is more strongly influenced by NTU. This can be also seen from Fig.4, which shows the heat transfer power vs time between fluid and wall for different NTU. We can see from Fig.4 that Q is completely different for small and large NTU. Because of cooling capacity being the integration of Q in a cycle, although Q for large NTU is larger than that for small NTU, the difference of the integration of Q in a cycle is not lager. So the cooling capacity is almost same for small and large when NTU is 0.8 - oo. Certainly, too small NTU (e.g. 0.2) is disadvantageous, because heat transfer quantity can not be transferred fully to the cold head. Based on above analysis, we can draw the conclusion that small NTU(about 0.8) is enough to a G-M refrigerate. Too large NTU isn't necessary. Certainly, too small NTU (< 0.2) is disadvantageous.
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ICEC16/ICMC Proceedings 5.0
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ACKNOWLEDGEMENT This research is supported by the National Natural Science Foundation of China. Helpful discussions with Prof. C.S.Hong are sincerely appreciated. REFERENCE 1. Satoh T. et al., Development of 1.5W 4KG-M cryocooler with magnetic regenerator material, Presented at CEC/ICMC 95, Columbus OH, July 17-21.1995. 2. Zhang L. et al., Development and Analysis of a 4.2K G-M Refrigerator, Cryogenics, (1994) 34 ICEC Supplement 179-182 3. Riedy R.C., Low temperature, high performance G-M refrigerator, Cryogenics (1993) 33 653-657 4. Internal Report: Investigation of G-M refrigerator, 45-48 Institute of Physics, Chinese Academy of Sciences(1976), (In Chinese).
Improvement of a 1.5W-class 4K Gifford-McMahon cryocooler
Atsushi Onishi, Rui Li, Hiroshi Asami, Toshimi Satoh and Yoshiaki Kanazawa Research & Development Center, Sumitomo Heavy Industries, Ltd. 63-30, Yuhigaoka, Hiratsuka, Kanagawa, 254 JAPAN
A 1.5W-class 4K Gifford-McMahon(GM) cryocooler has been developed by using magnetic regenerator material ErNi0.9Co0.~. In this paper, the effect of cylinder size for the cooling capaci}y is reported. Various size of cylinders and displacers were tested, and a drastic effect of the size on the cooling capacity were obtained. When these cylinders were tested with same compressor, certain relation was found between the data of the first stage cooling capacity and the second stage cooling capacity. But when the compressor was changed, the newly obtained data formed another relation. The maximum cooling capacity of the first stage at 40K was 53.7W, and that of the second stage at 4.2K was 1.74W.
INTRODUCTION GM cryocooler was invented in 1959[1], and is even now widelyused for cryopump, MRI system, and so on, in almost original design. After 30 years from invention, in 1989, GM cryocooler using magnetic regenerator material, called "4K-GM cryocooler", appeared on the stage[2,3]. From then, 4K-GM cryocooler has been paid an attention for the convenient device to make lower temperature than 4.2K. Our research group has developed 4K-GM cryocooler since 1990, and in the last year, 4K-GM cryocooler producing cooling capacity larger than 1.5W at 4.2K was successfully developed[4]. But then, cylinder size of the cryocooler was as same as commodity of our company, and the effect of cylinder size for the performance of the cryocooler has not to be confirmed. Therefore, cylinder size and regenerator were changed and the effect for cooling capacity were investigated. APPARATUS For our experiments, four sizes of cylinders were prepared. The inside diameter and the depth of these cylinders are shown in Table 1. Cylinder No. 1 is same in size as that of SRDK-408B type 4K-GM cryocooler, the commodity of Sumitomo Heavy Industries, Ltd. Table 1 Cylinder sizes of4K-GM cryocooler
Cylinder No. 1 2 3 4
First cylinder Inside diameter Depth [mm] [mm] 82 199 82 199 82 250 85 199
Second cylinder Inside diameter Depth [ram] [ram] 35 200 40 200 40 200 35 200
Ratio of the cylinder volume (The second volume) / (The first volume) 0.183 0.239 0.190 0.170
Copper mesh and lead sphere are used for regenerator materials in tile first displacer, and lead and ErNi0.gCo0.1[5] sphere were stuft'ed in tile second displacer. Specific heat of lead and ErNio.gCoo.i are shown in Figure l, compared with conventional lnagnetic regenerator material Er3Ni, and the quantity of regenerator materials for each cylinder are described in Table 2. 351
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Figure 1 Specific heat of regenerator materials Table 2 Quantity of regenerator materials
Cylinder No. 1 2 3 4
First regenerator Copper mesh [pieces] Lead sphere [g] 550 560 550 560 875 1,080 550 460
Second regenerator Lead sphere [g] ErNio.9COo.lsphere [g] 350 350 500 500 500 500 350 350
Two types of compressor units were prepared. The construction of these compressor units are indicated in Table 3. Both CSW-51and CSW-71 are products by Sumitomo Heavy Industries, Ltd. Stroke of displacer was 30ram, and cycle speed was selected 60 cycles per minute. Heat load for each stage was impressed from manganin wire wound around each stage, and the temperature was measured by platinum-cobalt resistance thermometer for the first stage, and by germanium resistance thermometer for the second stage. Table 3 Construction of compressor units
Compressor No. Compressor type 1 CSW-71 2 CSW-71 + CSW-51 (connected parallel)
Initial pressure of helium gas [MPa] 1.77 1.57
Input electric power [kW] 7 10
EXPERIMENTAL RESULT AND DISCUSSION First, the difference of the cooling capacity by the combination in each cylinder with compressor No. 1 was tested. The result is shown in Figure 2. When the first stage heat loads of 0, 20, 40W were added, then the first stage temperature and the second stage cooling capacity at 4.2K were measured. The figure indicates when tl~e first stage heat load became bigger, then the second stage cooling capacity became larger. On the other hand, the cylinder having larger ratio of the second cylinder volume to the first cylinder volume in Table 1 brought larger cooling capacity. The reason is considered that the amount of helium gas supplied from compressor is constant, not related to the total volume of cylinder, and the gas is distributed according to the ratio of the second cylinder volume to the first cylinder volume, so the second cylinder supplied much helium gas brings larger cooling capacity.
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Having an eye to the first stage, cylinder No.3 reached the lowest temperature in these four cylinders. Comparing cylinder No.3 with cylinder No.2, the first stage cooling capacity became higher about 15W. The reason is considered as follows (1) By reason of the extension of the first cylinder, heat loss (conduction loss and shuttle heat loss) decreased. (2) Because of increase in quantity of the regenerator material in the first regenerator, cooling capacity increased. To confirm the reason (1) quantitatively, conduction loss and shuttle heat loss[6] on cylinder No.2 and No.3 were calculated. The result is shown on Table 4. Total heat loss of cylinder No.2 was 24.6W, but that of cylinder No.3 was 16.4W. It is considered that cooling capacity of the first stage increases about 8W according to the extension of the first cylinder. About 7W obtained by subtracting above calculated result from experimental result was considered the effect of the reason (2). Table 4 Calculated heat losses
Shuttle heat [W] Conduction [W] Total heat loss [W]
Cylinder No.2 16.2 8.4 24.6
Cylinder No.3 10.8 5.6 16.4
Next, the first stage cooling capacity at 40K and the second, stage cooling capacity at 4.2K were measured. Seven combinations of tbur types of cylinders and two types of compressors (except the combination of cylinder No.3 and compressor No.2) were tested. The result is shown in Figure 3. Classifying for the type of compressor, these data are distributed on the straight line. The maximum cooling capacity of the first stage by the compressor No.1 was 41.3W, and by the compressor No.2 was 53.7W (cylinder No.4, in both cases). On the other hand, the maximum cooling capacity of the second stage by the compressor No. 1 was 1.49W, and by the compressor No.2 was 1.74W (cylinder No.2, in both cases).
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354 CONCLUSION
The results of this work are as follows: (1) In the work on 1.5W-class 4K-GM cryocooler, performance of four cylinders having different size were tested. Comparing with data obtained by using same compressor, the larger the ratio of the second cylinder volume to tile first cylinder volume of the cylinder, the larger cooling capacity was produced. (2) Owing to extend the length of the first cylinder, cooling capacity of the first stage rose up about 15W. The details of this effect were calculated at about 8W for the decrease of heat loss, and about 7W for the increase of regenerator materials. (3) Plotting data of the first stage cooling capacity at 40K and the second stage cooling capacity at 4.2K, these data placed on the straight line. The maximum data of the first stagecooling capacity at 40K was obtained 53.7W by using cylinder No.4 and compressor No.2, and that of the second stage cooling capacity at 4.2K was obtained 1.74W by using cylinder No.2 and compressor No.2. REFERENCES 1. 2.
3. 4. 5.
6.
McMahon, H. O. and Gifford, W. E., A new low-temperature gas expansion cycle, part I Adv. Cryog. Eng. (1960) 5 354-367 Kuriyama, T., Hakamada, R., Nakagome, H., Tokai, Y., Sahashi, M., Li, R., Yoshida, O., Matsumoto, K. and Hashimoto, T., High efficient two-stage GM refrigerator with magnetic material in the liquid helium temperature region Adv. Cryog. Eng. (1990) 36B 1261-1269 Inaguchi, T., Nagao, M. and Yoshimura, H., Two-stage Gifford-McMahon cycle cryocooler operating at about 2K Proc. 6th Intl. Cry0coolers Conf. (1990) II 25-36 Satoh, T., Onishi, A., Li, R., Asami, H. and Kanazawa, Y., Development of 1.5W 4K G-M cryocooler with magnetic regenerator material Adv. Cry_0g. Eng. (1996) 4__.!.1(to be published) Onishi, A., Li, R., Satoh, T., Kanazawa, Y., Makuuchi, H., Aikawa, S. and Hashimoto, T., A-4K-GM cryocooler with hybrid regenerator of magnetic materials Proc. 4th Joint Sino-Japanese Seminar on .C.ryocoolers and concerned Topics (1993) 44-48 Zimmerman, F. J. and Longsworth, R. C., Shuttle heat transfer Adv. Cryog. Eng. (1970) l__fi6342-351
A Simple Method of T e m p e r a t u r e Stabilization for 4 K GM C r y o c o o l e r
Rui Li, Atsushi Onishi, Toshimi Satoh, Yoshiaki Kanazawa R & D Center, Sumitomo Heavy Industries, Ltd., 63-30, Yuhigaoka, Hiratsuka, Kanagawa, 254, JAPAN
A simple method of temperature stabilization for 4 K G M cryocooler has been proposed. A copper pot connected to compressor unit by a capillary tube is mounted on the 4 K stage of a GM cryocooler. The utilization of high volumetric specific heat of pressurized helium in the pot at 4 K region produces a good temperature stability on the 4 K stage. The periodic temperature fluctuation of the 4 K stage is ,43.5 K(peak-to-peak) typically, but the helium pot described in this paper reduces the temperature fluctuation effectively down to ,43.05 K (peak-topeak). This paper shows the experimental details, and discusses the temperature stabilization effect of the helium pot.
INTRODUCTION The remarkable progress in 4 K Gifford-McMahon (GM) cryocooler[l,2] in the last several years provides a great possibility of operating superconducting devices, such as SQUIDs and SIS mixers, at 4 K region without liquid helium or helium liquefiers[3,4,5]. Mechanical cryocoolers like GM cryocooler, however, produce temperature fluctuations, magnetic noises and mechanical vibrations in operating systems. For a highly sensitive measurement, it is necessary to reduce the disturbances mainly caused by these noise sources. Kaiser, G., et a1.[6,7] has proposed a noise reduction method for cryocooler cooled operating systems. They used a conventional GM cryocooler with latent heat collector of liquid nitrogen or liquid neon, and reduced the temperature fluctuations down to 0.04 K ~ 0.15 K. In their system, however, the operating temperature was higher than 50 K, and the operating time for measurement was discontinuous. In this work, we developed a simple method of temperature stabilization for 4 K GM cryocooler. The high volumetric specific heat of pressurized helium at 4 K region was applied to reduce the temperature fluctuation of the 4 K stage. The experimental details and the results are described below.
EXPERIMENTAL DETAILS The schematic of the experimental apparatus employed in this work is shown in figure 1. The cold head (RD-210L4) and the compressor unit (CSW-71 A) are manufactured by Sumitomo Heavy Industries Limited. In the cold head, the second regenerator has both lead spheres and spherical magnetic regenerator material, ErNi0.gCo0.~, with the latter placed at the colder end. The cooling capacity of the cryocooler is about 0.6 W at 4.2 K. The pot used to reduce the temperature fluctuations is mounted on the 4 K stage, and is connected to the supply line of compressor unit by a capillary tube. The stainless steel capillary tube is 3.18 mm in outside diameter, and is arranged to have good heat contact with the first stage of the cryocooler. Figure 2 gives the details of the pot and the second stage. The copper pot is about 900 g in weight, and has a volume of 61 cm 3 which can be filled with helium. The temperature of the pot or the second stage is 355
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measured by a carbon glass resistor (Lake Shore Cryotronics, Inc.). The temperature fluctuations are measured for three cases: (1) there is no pot mounted on the 4 K stage, (2) the pot is mounted on the stage but there is nothing in it, (3) the pot is mounted on the stage and is filled with pressurized helium. We will simply call the three cases at the following as: (1) no pot, (2) vacuum pot, and (3) helium pot. The carbon glass resistor and a electric heater of manganin wire are held on the second stage for the no pot case, and on the copper pot for the cases of vacuum pot and helium pot. For simplifying explanations in the following sections, the temperature measured at the pot will also be called as the second stage temperature.
RESULTS AND DISCUSSIONS Figure 3 shows the test results of temperature stabilization at 4.2 K. For the no pot case, the 4 K stage temperature fluctuates periodically with the working gas cycle in the cold head. The peak-to-peak value of the temperature fluctuation is 0.53 K, and the value is typical for the 4 K GM cryocoolers with a partial
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Figure 6 Cool down curves of first stage and second stage for all cases
copper cylinder for the second stage. For the case of helium pot, however, the fluctuation is effectively reduced down to 0.054 K (peak-to-peak). The initial pressure of compressor unit is 1.76 MPa for both the cases, but the supply pressure is 2.35 MPa for no pot and 2.17 MPa for helium pot at 4.2 K. There is almost no difference in cooling capacity at 4.2 K between the two cases. In fact, the variance of the average temperature (TAv--(TMAx+TMIN)/2)drawn in Fig. 3 is less than 4 mK. The effect of temperature stabilization shown in Fig. 3 is brought about by the utilization of high volumetric specific heat of pressurized helium. Figure 4 gives volumetric specific heats of helium and some cryogenic materials. The volumetric specific heat of helium at 4 K for 2 MPa is ~49 times and -490 times as high as that of lead and copper respectively. In other words, the heat capacity of the helium gas (2 MPa, 4 K) filled in the pot equals that of copper block which is .-0.03 m 3 in volume and -267 kg in weight. Since pressurized helium shows a broad peak in specific heat below 20 K, Fig. 4 gives a hint that the pot can also be used to reduce the temperature fluctuations above 4 K. For all the cases, the temperature fluctuations below 20 K are illustrated in figure 5 as a function of the second stage temperature. The temperature fluctuation for the vacuum pot case is as same as that of the no pot case at 4 K region. It is means that the heat capacity of the copper pot is very poor and can be ignored. At higher temperatures, the heat capacity of the pot becomes larger, therefore the temperature fluctuation of vacuum pot is smaller than that of no pot. It is clear that the effect of temperature stabilization is more superior for the helium pot case. Compared with the no pot case, the temperature fluctuation of helium pot is almost reduced by a factor of l0 below 20 K. The maximum temperature fluctuations shown in Fig. 5 are 1.57 K for the no pot case, 0.49 K for the vacuum pot case and 0.12 K for the helium pot case. The difference in cool down time for the three cases is also an interesting matter. Figure 6 shows the cool down curves of the first stage and the second stage for all the cases. The second stage is cooled more rapidly than the first stage for the no pot case, but is changed more slowly when the pot is mounted on the stage. The total cool down time seems to be 140 minutes for the helium pot case, and 120 minutes for the other cases. The fact indicates that there is no remarkable difference in cool down time whether the pot is mounted or not. Furthermore, because of the low specific heat of helium at higher temperatures, it is similar in cool down curves above 50 K whether the pot is filled with helium. It is the important point of our experiment that the helium pot is not a closed one but connected to the compressor unit. By connecting the pot to the supply line of compressor unit, the pressurized helium naturally goes into or out of the pot, depended on the temperature of the pot. As a result, the pressure in the pot is always equal to the maximum of compressor pressure whether the system is under operating or not, and it is easy to cool down owing to the low specific heat of helium at higher temperatures.
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CONCLUSIONS A simple method of temperature stabilization for 4 K GM cryocooler has been developed. With the copper pot connected to the compressor unit, the temperature fluctuation of the second stage is reduced from 0.53 K down to 0.054 K at 4 K, and from 1.57 K down to 0. 12 K at ~10 K. The system can be operated continuously, and is easy to cool down, furthermore it is safe to use.
REFERENCES Li, R., Onishi, A., Satoh, T. and Kanazawa, Y., Influence of valve open timing and interval on performance of 4 K Gifford-McMahon cycle cryocooler, Adv. Cryog. Eng., Plenum Press, New York (1996) 41 in press Satoh, T., Onishi, A., Li, R., Asami, H. and Kanazawa, Y., Development of 1.5 W 4 K G-M cryocooler with magnetic regenerator material, Adv. Cryog. Eng., Plenum Press, New York (1996) 4_J1in press Plambeck, R., Thatte, N. and Sykes, P., A 4 K Gifford-McMahon refrigerator for astronomy, In: Proc. 7th International Cryocooler Conference, Air Force Phillips Laboratory Report PL-CP-93100 l, Kirtland AFB, NM (1993) 401-415 Takahashi, M., Hatakeyama, H., Kuriyama, T., Nakagome, H., Kawabe, R., Iwashita, H., McCulloch, G., Shibata, K. and Ukita, S., A compact 150 GHz SIS receiver cooled by 4 K GM refrigerator, In: Proc. 7th International Cryocooler Conference, Air Force Phillips Laboratory Report PL-CP-93-100 l, Kirtland AFB, NM (1993) 495-507 Fujimoto, S., Kazami, K., Takada, Y., Yoshida, T., Ogata, H. and Kado, H., Cooling of SQUIDs using a Gifford-McMahon cryocooler containing magnetic regenerative material to measure biomagnetism, Cryogenics (1995) 35 143-148 Kaiser, G., Seidel, P. and Thurk, M., Noise reduction of cryo-refrigerators, Adv. Cryog. Eng,, Plenum Press, New York (1994) 3__991281 - 1285 Kaiser, G., Dorrer, L., Matthes, A., Seidel, P., Schmidl, F., Schneidewind, H. and Thurk, M., Cooling of HTSC Josephson junctions and SQUIDs with cryo-refrigerators, Cryogenics (1994) 3_4 !CEC Supplement 891-894
Numerical Simulation and Test of Regenerator For Two-Stage G-M Cryocooler Zhongping Huang, Yan Chen, Huaiyu Pan, Shimo I.i Cryogenics Lab. ,Zhejiang University,Hangzhou 310027,China
In this paper, the second-stage regenerator of G-M Cryocooler is mainly researched. Numerical calculation of the regenerator has been carried out. The calculution of regenerator effeciency at different temperatures showed that the regenerator effeciency increased when temperature was below 9k. The experimental results of various regenerative matrix in the second-stage regenerator are obtained. If sintered powder metal hydrides is used as the matrix, high strength and suitable ratio of the mixture must be considered.
INTRODUCTION Regenerator effeciency is directly proportional to the ratio of heat capacity of matrix in regenerator to that of helium in void volume. The finite heat capacity of matrix is the main loss of regenerator at low temperature ,so the key point is to search for new types of regenerative matrices with high heat capacity at low temperatures. In this paper, we make the numerical simulation of the gernerator for a two-stage G-M cryocooler. A number of regenerative materials to a two-stage G-M cryocooler have also been studied. R E G E N E R A T O R MODEL AND N U M E R I C A l . S I M U L A T I O N We choose a simplified model of second-stage regenerator in a two-stage G-M cryocooler with the assumptions of one-demensional flow with zero longitudinal thermal conductivity, infinite transverse thermal conductivity, neglect of wall effect, zero pressure drop, stable and periodic change of each parameter. Thus the energy differential equations for the gas and matrix are: h~ L (T
h.Af (T
YI'f OTf dp -~--+-C~pA~ -~-Aoo~ ~--~-=0
(1)
0(CmT.,) at
(2)
and the continuous equation for gas is
0x +Ar
=0
(3)
Using rectanglar differential network, the node numbers in the direction of length and time are N. and N, respectively, the temprature at arbitrary node is T ( i , j ) The differential equations (1) (2) and (3) are turned into: 359
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h. (i ,j) A, i-T, (i ,j)-T~ (i,j)-] I.
+rh(~, j)C~ (~ ,j)
T,(i, j)-T, (i-1 ,j) + C o ( i ,j)p (i ,j)Ac Dx
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h. (i, j) A, i-T, (i, j)-Tm (i, j) 7 = Mm
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(4)
(5)
(6)
BOUNDARY CONDITIONS Temperature Boundary Conditions We suppose the gas inlet temperatures at the hot and cold blow periods are constants. The temperature of matrix and gas of one point at the end of hot or cold blow periods is equal to that of the point at the beginning of next hot and cold blow period. The temperature distribution on matrix is linear at the begining of the first hot blow period. Pressure Boundary Conditions According to the consumption, the pressure of the regenerator is constant at the same time, thus in evaluation process, we can use the experimental pressure wave or the approximate linear or sine function of p (t). Flow Rate Boundary Conditions We assume the flow rate at the inlet of two sections of hot and cold blow periods keeps constant (unvariable). NUMERICAl. SIMUI.ATION RESUI.TS -Temperature Profile of Gas and Matrices Figure1 and 2 show the variation of temperature of gas and matrix with time and position at different blow period by the method described above. The temperatures of the warm and cool ends of the second-stage regenerator(height: 60mm; diameter: 22mm) are T h = 4 0 K and T o = 6 . 4 K respectively. Er3Ni is used as the regenerative meterial. The pressure boundary is linear and the flow rate boundary is constant at the inlet. From Figure 1 and 2, we find the slope of the temperature distribution for gas and matrix near the cool end is always flat. This is because when the temperature is below 20K, the increase of helium heat capacity is almost equal to the decrease of matrix heat capacity, the thermal saturation in the regenerator appears, and then the temperature difference between gas and matrix and heat exchange become very small. So, the temperature varies gently. The result shows that the same situation appears at the cold blow period. The resluts also show that there are some common characters of matrix temperature variation with time in all sections except the middle one. In one period, the matrix temperature variation curve of the hot blow period does not coincide with that of the cold blow period at the same section and they can be further joined to a closed curve. The average matrix temperature at the hot blow period is always higher than that at the cold blow period.
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Profile of gas temperature
Figure 2
Profile of matrix temperature
Distribution of Flow Rate The variance of gas temperature and pressure in the void volume of the regenerator leads to the variance of gas stored in the void volume, thus the flow rate in the regenerator is a quadratic function of poisition and time. Using the same calculation data as above, we get the distribution of flow rate as shown in Figure 3. We find (1) at the same time, the variance of flow rate is flat at the hot end. This is because when temperature is higher than 20K, temperature has little effect on helium density, the flow rate in the void volume varies little with the variance of temperature. (2) The flow rate at the warm section is lower than that at the cool section.
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6 7 8 Cool end t e m p e r a t u r e ( K )
9
Regenerator effeciency at different cool end temperatures
Regenerator Effeciency Considering the whole circulation, the net enthalpy flow entering to the cool end in one circle in the regenerater is a kind of cooling capacity loss. The effeciency of the whole circulation is as follows: fl= 1-
/~HR ,~Cmi.(Th-Tc)ctt
(7)
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In equation ( 7 ) , the numerator is the net enthalpy flow entering to the cool end in one circle and the denominator is the maxium heat transfer in one circle. Figure 4 shows the regenerator effeciency for different temperatures at cool end when the temperature at the warm end is 40K. We can see from Figure 4, the regenerator loss is decreased when Tc is lower than 9K. This is because the gas compressive heat decreases in this case, the effeciency of the regenerator is increased. T h e calculation also indicates that the regenerator effeciency is increased with the increasing of the warm end temperature. That means for reaching the same limit refrigerating t e m p e r a t u r e , the higher the warm temperature is, the higher the regenerator effeciency is. EXPERIMENT RESUI.TS In the experiments we used same cryocooler described in reference(4). We filled 150g lead balls at the cold part in the first-stage regenerator to form combined structure and to replace the whole phosphor bronze in the original cryocooler. The second-stage regenerative matrix is Pb as the original. The results of experiment are showed that after the improvement of the first stage regenerator, the mininum cooling temperature of the first and the second stages decrease from 51.8K to 42K and 9.4K to 7. OK respectively. This is because the heat calSacity of Pb is larger than that of phosphor bronze. When Pb is used as the first stage regenerative matrix, the matrix heat capacity increases, and the regenerator loss decreases. This leads first stage temperature to decrease, and the firststage displacer is located in a further uniform cold environment, the heat load decreases and the minium temperature drops. In this experiment, sintered poweder metal hydrides are used as matrix. The experiment shows that the presieion and intensity become important index for sintered material when it is used as regenerative matrix. But this is still a problem in our s t u d y , and whenever the presicion and intensity of sintered material is resolved, the better regenerator effeciency can be obtained. CONCLUSI()NS From above analysis, we can conclude= 1. Regarding the working fluid as a real gas and considering the effect of gas in the void volume, we get the temperature profile of gas and matrices in the regenerator and gas flow rate distribution. By calculation of the regenerator effeciency under different temperaturses, it is found that the effeciency of regenerator increase with the rapid decrease of gas compressive heat in the regenerator. 2. No-loading cooling temperature of the second-stage is reduced from 9.4K to 7. OK by improvement of the first-stage regenerative matrix on the original two-stage G-M cryocoolor. REFERENCES 1 2 3 4 5
Hong C. S. , Xu X. D. , on the unique feature of the low temperature regenerator in a 4.2K G-M refrigerator Cryocoolors and concerned topics JSJS IV(1993)6-13 Knyiyama et al. , High efficient two-stage G-M refrigerator with magnetic ma.terial in the liquid helium temperature region ACE(1990)35 Ackerman R. A. , A heat-balance analysis of a Gifford-Mcmahon cryorefrigerator CEC(1970)16 221 I.i S. M . , Mao Y. H . , Theory and experiment for rare-earth magnetic regenerative materials ACE(1994)40 625-630 Huang B.J. and I.u C. W . , Dynamic modelling of a Gifford-Me Mahon cryorefrigerator Cryogenics (1993) 33 1046
PERFORMANCE OF A TWIN COLD FINGER GIFFORD-McMAHON CRYOCOOLER J.M. Pfotenhauer, O.D. Lokken, & P.E. Gifford* Applied Superconductivity Center, University of Wisconsin- Madison 1500 Engineering Dr., Madison, WI 53706 USA *Cryomech, Inc. 113 Falso Dr., Syracuse NY 13211 USA A single stage Gifford McMahon cryocooler with two identical cold fingers electrically isolated from each other has been fabricated to cool a pair of 1500 A, HTS current leads. 60 watts of cooling capacity are provided at 70 K on each of the cold fingers. In that the necessary voltage isolation is provided in a region where heat is transferred by the helium gas internal to the cryocooler, thermal inefficiencies associated with an electrical isolator in the cold bus conduction path are avoided. Initial performance results of the cryocooler are reported in terms of measured heat loads at various operating temperatures.
INTRODUCTION In the development of high temperature superconducting (HTS) current leads used in conjunction with low temperature superconducting magnets, it has been recognized [ 1,2] that a thermal intercept is required above the HTS section of the current lead to ensure a minimum total refrigeration load associated with the operation of the current lead. The thermal intercept is often accomplished through the use of a cryocooler thermally connected to the current lead. In order to avoid a short circuit of the current path at this intercept, it is common to include an electrical isolation piece between the single cold finger of the cryocooler and the thermal intercept point on each of the current leads (see figure 1a). Unfortunately, the electrical insulator is typically also a thermal insulator and significant temperature drops can be realized across this member during operation of the current lead. For example, the thermal load from the upper stage of a 1.5 kA conduction cooled current lead operating between room temperature and 70 K is expected to be ~ 60 watts on each polarity. The temperature drop across a G-10 insulator with nominal dimensions of 0.5 mm thickness and 8 x 10 -3 m 2 cross sectional area is in excess of 10 K even without consideration of contact resistance. A decreased cryocooler efficiency will be associated with this thermal resistance since the cryocooler must operate at a temperature 10 K colder to provide the 70 K intercept temperature at the current lead. The added thermal resistance represents a decrease in efficiency for a Carnot refrigeration cycle from 0.3 to 0.25, or a minimum increase of 22% in the equivalently required room temperature power. It is therefore of significant advantage to eliminate this thermal resistance. The present paper presents a means of providing the necessary electrical isolation between the current leads through the use of a twin-cold-finger Gifford McMahon cryocooler which has been fabricated and run through initial performance tests. In this approach, heat is carried from the cold finger to the warm end of the GM cooler via the internal helium gas, while the electrical isolation is built in to the structure of the cooler itself. A pure metal to metal contact is allowed between each cold finger and their respective current leads - thus eliminating the previous thermal resistance, and the electrical insulation within the body of the cryocooler has no influence on the means of transferring the heat away via the helium gas. The internal structure of the GM cooler is designed to avoid an electrical breakdown in the helium gas. 363
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HTS current lead
stage~
copper
I
cryocooler I cold finger I
thermal, intercept
lower stage
Ill
insulation upper stage copper
I I I insulation
HTS current lead
thermal intercept
I
a)
lower stage
b)
Figure 1. a) conventional configuration for voltage isolation between current lead and cryocooler, b) incorporation of voltage isolation within cryocooler. INITIAL OPERATION OF THE TWIN FINGER GM COOLER A scaled representation of the twin cold finger GM cooler is shown in figure 2. This device, jointly developed by the UW Applied Superconductivity Center and Cryomech has been tested both in a standalone configuration at Cryomech and within a current lead test rig at the University of Wisconsin - Madison. In the current lead test rig configuration, the twin cold finger cooler (model PF01), and an additional standard 60 watt cooler (model AL60) are both powered by the 5 kW model CP640 compressor. Expansion chamber volumes have been designed for a nominal cooling power of 60 watts @ 70 K at each cold finger. The refrigeration capacity performance map for both the stand-alone test and the current lead installation test is shown in figure 3. Note that capacity per cold finger is shown here. Refrigeration capacity values in the stand alone test have been obtained by applying a measured heat load directly to each cold finger and monitoring the resulting temperature. The method for determining the refrigeration capacity in the current lead test rig was somewhat more involved and less direct, but the results are in good agreement with the stand alone values. Determination of cooling power from cooldown process The twin cold finger GM cooler has been installed into a mating twin-finger cold well on the UW-Madison current lead test rig to investigate the ability of changing the cryocooler without breaking the vacuum on the test dewar. During the initial operation of the cryocooler in which the copper bus alone was attached to the cryocooler (no current leads installed), it has been observed that a significant thermal resistance is associated with the specific design of the cold well - cryocooler contact. While the cold well will require re-design, the result of this configuration is that the temperatures measured on the copper cold bus between the cryocooler and the current lead are significantly higher than the cryocooler cold finger. Nevertheless, the rate of temperature drop observed during the cooldown process have been useful for determining the cryocooler power. Total cooldown time with the two copper buses attached (14.6 kg) is 2 hours. The thermal resistance of 0.8 K/W determined to exist between the cryocooler cold finger and the copper bus is significantly higher than the thermal resistance associated with the geometry of the copper bus Rt, bus = L / (k A)
(1)
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200
365
i
i
9 150
I
i
Oi ~
oi
UW- Madison
[] Cryomech .....................................................................o...o .........
i:
O
high, low pressure connections
1 o o ........................ i ................................ a = i
:
i,
:
'
::
electrical isolation
Cold Fingers
~o
0 -20
0
20
,
40
[]
. . . . . . . . . . . . . . . . . . . . . . . .
v
9
:
!
60
'
80
'
100120
Cooling P o w e r (W)
Figure 2. Twin Cold Finger GM Cryocooler
Figure 3. Refrigeration performance per cold finger.
allowing a "lumped capacitance" treatment of the copper bus cooldown. Here L and A are respectively the length in the direction of, and cross sectional area perpendicular to, the heat flow, and k is the thermal conductivity of the copper. From the lumped capacitance model, the heat flow (q, in watts) is simply proportional to the cooldown rate, the mass (M) and specific heat (Cv) of the copper bus according to
dT q=MCv d--T"
(2)
The associated values of temperature for each value of q were obtained from the copper bus temperature data by correcting for the thermal resistance of the bus - cold finger contact. As can be seen in figure 3 the agreement is quite good with the values obtained in the stand-alone test. In addition, when combined with the performance of the AL60 driven simultaneously by the same compressor, the total refrigeration power provided at 70 K is 170 watts; a value which compares well with the 180 watt capacity at 70 K of the single AL200 cold head driven by the same compressor. Before installing the current leads into the test rig, the voltage isolation between the two cold fingers was measured using a standard HIPOT device. With all parts at ambient temperature a voltage isolation of 2.8 kV was measured.
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INTEGRATION WITH UPPER STAGE OF CURRENT LEADS The twin cold finger cooler has been used to cool the upper (conventional) stage of the current leads before attaching the lower HTS stages. The goal of this measurement was to characterize the resistance of the leads in their cold operating state and to confirm that sufficient cooling power can be provided by the cryocooler for the upper stage heat load. Voltage taps have been installed which span the distance from the current lead flag to the bottom, or thermal intercept point. A 1000 amp supply was used on the first test of this setup and the resulting voltages across the (+) and (-) leads, as well as the connection bus between them, are shown in figure 4. The data display that the two current leads have significantly different resistances, and that there will be need to monitor the (-) lead closely during subsequent operation. At all values of current, the flag temperature was maintained below 25 ~ demonstrating that sufficient cooling is provided by the cryocooler. A
E (n a.
m q)
a) El 4)
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...... I
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1000
i
i
1200
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Figure 4. Resistance measurements of current leads cooled by conduction to GM cryocooler alone. SUMMARY A single stage, twin cold finger GM cryocooler has been successfully developed for use in cooling the upper stage of HTS current leads. A nominal cooling power of 60 watts is provided at each cold finger at 70 K, and a room temperature voltage isolation of 2.8 kV is provided. The refrigeration performance of the twin cold finger cooler has been mapped both with a direct measurement, and by using the transient cooldown data. Both methods of quantifying the capacity agree with each other. The twin finger cryocooler has been demonstrated to provide sufficient cooling for operation of 1 kA current leads. REFERENCES 1. Wesche, R. and Fuchs, A.M., "Design of Superconducting Current Leads," Cryogenics, 33 (1993), 714-718. 2. Yang, S. and Pfotenhauer, J.M., "Optimization of the Intercept Temperature for High Temperature Superconducting Current Lead," presented at 1995 CEC, Columbus Ohio.
Development of two-stage 4K cryocooler with modified Solvay cycle
H.Torii, T.Kurihara *, H.Morishita, K.Miura Daikin Industries ,Ltd. VC Division Engineering Section 3-12 Chikko-shinmachi, Sakai, Osaka 592, Japan *Daikin Industries,Ltd Mechanical Engineering Laboratory 1304 Kanaokacyo, Sakai, Osaka 591, Japan
During developing two-stage 4K cry ocooler with modified Solvay cycle, regenerator material of 2nd displacer and its packing structure have been studied by computer simulation and experiment. Assuming the specific heat model, cooling capacity has been calculated. Its results as follows; specific heat of hot end does not so effect to the cooling capacity, but of cold end effects to the cooling capacity. It is revealed that there is the suitable specific heat for 2nd regenerator. In this paper, the investigation results of 2nd regenerator are mainly described.
1 INTRODUCTION Superconducting magnet has been cooled by liquid helium. There are many examples of cryocooler (shield cooler) which are used to cool the shield of superconducting magnet to reduce the boil-off rate. Because of its simple structure and high reliability, GM type cryocooler has been used widely. On the other hand, development of cooling superconducting magnet directly by cry ocooler tll has been progressed. Development of the highly reliable cryocooler like shield cooler, is indispensable to direct cooling. Many researches about motor driven 4K-GM cryocooler were carried out. I21'I3]'[41 However there is a few research reports on pneumatic driven 4K-GM cryocooler. I51 Authors have developed pneumatic driven 4K-GM cryocooler and studied the characteristics of regenerator of 2nd displacer and its packing structure. 2
MAIN DIMENSIONS AND CHARACTERISTIC OF PROTOTYPE
Prototype is a modified Solvay cycle(Daikin model V208SCL). [61 In case of experiment and computer simulation, parts such as motor and cylinder except for the displacer are common with original conventional 20K cryocooler. Major improvements have been done on the regenerator and displacer stroke (movement distance). Table-1 presents main dimensions and specification. Table -1
Comparison of dimensions and cycle characteristic prototype
Displacer stroke 1st cylinder diameter 2nd cylinder diameter Reciprocating speed
9mm
original 20K cryocooler 17mm
70mm 32 mm 144 r.p.m 367
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ICEC16/ICMC Proceedings PERFORMANCE OF 2ND DISPLACER REGENERATOR 40
I
o intake start 3.1 Hot End Regenerator _ ~ [] exhaust start By computer simulation program developed by heat station temperature, 3O 1st :35K Daikin 171, temperature distribution in 2nd regenerator, during l refrigeration cycle, is shown Fig 1. Same as motor driven GM type cry ocooler 181, temperature 2o % ~: E amplitude of half of regenerator cold end (B portion) i B portion =) I-is comparatively small, on the other hand, half of hot I0 end (A portion) temperature amplitude is very large. At first , we simulated about hot end regenerator. Relationship between specific heat of the regenerator 0 0.5 1 and temperature is shown in Fig2-(l). 4 kinds hot end cold end 2nd Regenerator Position specific heat characteristic patterns are assumed. For Fig.1 Axial temperature distribution each pattern, specific heat below 10K is constant and of 2nd regenerator has same value of Er3Ni specific heat at 1OK. As for 1.S temperature range above 10K, TYPE 1 specific heat is constant TYPE 4 with temperature. TYPE 2,3 and 4 specific heats are in R" 1 :-"'~'"~ ..-"" TYPE 3 proportion to temperature. Each type of regenerator is packed fully in 2nd displacer. The results of computer simulation are O. 6 ~ TYPE 2 shown in Fig 2-(2). Specific heat greater than TYPE 2 has a u~ O. 3 TYPE 1 .....:-/._ sufficient cooling capacity. The value of specific heat is assumed" 0 .-' ~ I I I I 0.rJ/Kcm 3 at 35K, and it has enough a capability, for hot end 0 10 20 30 40 50 regenerator. Specific heat of various regenerator reported until Temperature(K) Fig.2-(1) Specific heat of regeneratormaterials now is greater than 0.6 J/Kcm 3 at 35K 191 l]01. Therefore it is thought that various regenerator could be used for the hot end. ' TYPE 3 ' TYPE 4 A
v
ID 3
,,
=
O.S
-
TYPE 2
r
,,=:
~Ni
~
+
...I=
o
0.9
.I----'
3.2 Cold End Regenerator Three levels packing structure regenerator are investigated; see Table-2, because specific heat of lead (Pb) decreases below 15K rapidly and the specific heat of regenerator is in proportion to temperature like as the model described in Fig.2-(l ). For the 2nd displacer several different kinds of material in the
o ca.
or .__. ~ o
0.8 heat stage temperature
Ist 93 5 K , ar
0.7
cold end, authers carried out computer simulation in 4.2K and the measurement of cooling capacity . The compressor unit used in the experiment is Daikin U110CW (6.7kW). Table -2 presents the kind of the regenerators which were packed in 2nd displacer cold end. Fig 3-(1) shows relationship between specific heat of the regenerators and temperature. Packing quantity of regenerator is approximately half for motor driven GM cryocooler (total 245 g). 1~] Table -2 Packin~ structure Hot end Intermediate Cold end TEST A TEST B TEST C TEST D
of regenerator Lead sphere Er 3 Ni sphere Er 3 Co sphere Er 3 Ni sphere ErNi sphere ErNi o.9 Co 0.] sphere
co oE
65 g(25%) 55 g(25%) 122 g 124 g 125 g 125 g
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0
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peak
30
Fig.3-(1 ) Specific heat of regenerator materials
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Relationship between specific heat peak value of the cold end regenerator and cooling capacity at 4.2K, by computer simulation and experiment, is shown in Fig. 3-(2). The points designated by X in Fig. 3-(1) indicates the peak of specific heat. Relationship between specific heat of cold end regenerator at 4.2K and cooling capacity is shown in Fig. 3-(3). Experimental results agree with computer simulation to some degree. As shown in Fig 3-(2), correlation between specific heat peak value and cooling capacity isn't found. As shown in Fig 3-(3), it is supposed that cooling capacity ~- ~.z Ist :35K [ ! . , O expenment 2nd : 4.2 13 calculation is in proportion to regenerator specific heat at 4.2K. In Fig 3-(4), vv ~i 1 relation between cooling capacity and temperature by experiment is m 0.8 TEST B TEST A " shown. Table-3 presents the lowest attainable temperature for each o o 0.6 8 _ 1st and 2nd heat stages. n t o 0.4 TEST D TEST C-
•
Table -3
A
1 0.8
o
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e
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~
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-
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1 _-_
-
o
heat stage temperatur 1st :35K / t 2nd:~.2K / 0.15 0.2 0.Z5
Specific Heat at 4.2K (dIK .cm3) Fig.3-(3) Cooling capacity vs 4.2K specific heat
4
o
O2o2.
Fig.3-(2) Cooling capacity vs peak value 3
'
o t~
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0
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Lowest attainable temperature (Experiment) 1st stage 17.7K 2nd stage 2.39K
o o O 0
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4
5
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2nd Stage Temperature (K) Fig.3-(4) Cooling capacity with 2nd temperature
IMPROVEMENT OF COOLING CAPACITY
~ " 0.8
I pc,.ak'temperatu~e ]
E O
Assuming the regenerator models with different specific heat peak, computer simulation was carried out in order to investigate the influence of specific heat peak on the point of peak temperature, size and width.
4.1 Influence of Temperature on Specific Heat Peak Characteristic of specific heat of regenerator for each model is presented in Fig 4-(1). For example, TYPE 5 and TYPE 6 have the same values specific heat peak, though the temperature differs from each other. In addition, each type regenerator is packed in 2nd displacer perfectly. Computer simulation result is shown in Fig. 4-(2). Type, having the specific heat peak from 4.2K to 5K, has greater cooling capacity.
v
~.~
"1" 0.4 0 .,._ ~ tO
0.2
~'
~
0.6
TYPE6 TYPE 5
s lb ls Temperature (K) Fig.4-(1 ) Specific heat of models
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1
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,
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I
5
5.5
% 6
6.5
Peak Temperature (K) Fig.4-(2) Relationship between cooling capacity and peak temperature
4.2 Influence of Specific Heat Pea.k .and Width Characteristic of specific heat of regenerator for each model is presented in Fig 4-(3). TYPE 7,8 and 9 have constant specific heat peak temperature and constant surrounded area by specific heat and temperature. Computer simulation result is shown in Fig. 4-(4). In case of width greater than 4 K, cooling capacity decreases with the increasing width. Region from 2K to 4K is the peak for cooling capacity. On the contrary, cooling capacity decreases below 2K. This peak width is estimated to be sufficient for regenerator.
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4.3 Comparison between Computer Simulation and Experiment As for Er0.sDy0.2Sb material which has sharp specific heat near 3.5K [111 and its peak width about 1.5K, computer simulation and experiment of cry ocooler, in which this material was packed in 2nd displacer, were carried out. Table 4 presents the test results and packing structure is the same as Table-2 except for cold end. Table 4 Comparin~ of cooling capacity ( 3.5K / 4.2K) Cold end materials Experiment Simulation Er0.sDy0.2Sb(50%) 0.27 / 0.52W 0.24 / 0.48W ErNi0.9Co0.1(50%) 0.46 / 0.93W 0.49 / 0.86W
1.6
A
v? 1.2
0.8
TYPE 8
.o 0 . 4
TYPE 7
"T"
0
A
5
Temperature (K)
15
1.1
Od
9Specific heat during 4.2K to 5 K being high 9Temperature width of specific heat peak being between 2K and 4K (in case of surrounded area, described above, being constant)
0
width of peak
I
Fig.4-(3) Specific heat of models ~. ,,e'
Comparing the computer simulation with experimental result, both cooling capacities agree well at 4.2K and 3.5K. According to the computer simulation result at 3.5K, specific heat of Er0.sDy0.2Sb is greater than ErNi0.gCo0.1, but peak width is small (less than 2K). To inprove cooling capacity, the following characteristic is seemed to be needed for the regenerator.
@
Q.
m
TYPE8
TYPE7
I
r
0.9 heat stage temperature 1st "351(
C 0
8
0.8
i
i
~
:4"21(I
2
4 6 8 10 Width of peak (K) Fig.4-(4) Relationship between cooling capacity at 4.2K and peak width
SUMMARY
Two-stage 4K cry ocooler with modified Solvay cy cle has been developed and examined by changing the displacer assemblies from conventional 20K cry ocooler. Maximum cooling capacity of 0.93 W has been achieved at 4.2K. Packing quantity of regenerator in 2nd displacer is approximately half for the reported motor driven cryocooler. By computer simulation for 2nd displacer, specific heat of hot end regenerator doesn't give influence on cooling capacity, but cold end specific heat effects greatly. For the pneumatic driven cryocooler, suitable characteristic of regenerator becomes clear to some degree. 1 W class cryocooler which is not mentioned in this paper will be reported in separate opportunity. 6 l.
2. 3. 4.
5. 6. 7. 8.
9. 10. 11.
REFARENCES M.Urata, Journal of Plasma and Fusion Research, Vol.71, No.11, p1095, (1995) T.Kuriyama et.al., Advance~ in Cryogenic Engineering, Voi.35, p1261, (1990) M.Nagao et.al., Advances in Cryogenic Engineering, Vol.35, p1251, (1990) R.C.Riedy, Cryogenics, Vol.33, No.6, p653, (1993) Guobang Chen et.al., Proceeding of 7th International Cryocooler Conference, (1993) Yoon-Myung Kang et al., Refrigeration, Vol.63, No.733, p50, (1988) T.Kurihara et.al, Cryogenic Engineering, Vol.31, No.4, p 197, (1996) T.Hashimoto et.al., Cryogenic Engineering, Vol.31, No.4, p209, (1996); Y.Tokai et al., Toshiba Review. Vo1.147, No.2, p124, (1992) T.Hahimoto et.al., Cryogenic Engineering, Vol.31, No.4, p 131, (1996); Y.Lon et al., J.Appl.Phys., 77 (1995) 2214
Cryocoolers
Stirling coolers
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Cyclic Analysis of the Two Stage Stirling Cycle Cryogenerator Natu P. V.* and Narayankhedkar K. G.** *Sardar Patel College of Engineering, Mumbai-400 058, India **Indian Institute of Technology, Bombay, Powai, Mumbai-400 076, India
Single stage Stirling cryogenerators are utilised in liquefaction of nitrogen. However, for lower temperatures like 20K, it becomes necessary to use two stage cryogenerator - - an arrangement where a single compression and two expansion spaces are provided. This necessitates use of two regenerators which results in reduction of the temperature gradient across the regenerators and thus the regenerator loss. This paper reports cyclic analysis of two stage Stirling cycle cryogenerator. The comparison of the results shows good agreement between the results of present cyclic analysis and the results reported in the literature.
INTRODUCTION Nitrogen liquefiers based on Stirling cycle operate between 300 to 70 K. For applications like cryopumping the temperatures required are of the order of 20 K. The lowest temperature obtained in a cryocooler is governed by various losses. One of the major losses is due to ineffectiveness of regenerat'or. The loss of refrigerating effect due to regenerator ineffectiveness can be reduced by reducing temperature gradient across the regenerator. Thus for attaining lower temperatures, more than one regenerator need to be used. For example, for attaining 20 K, the range of 300 to 20 K can be divided into two ranges i.e. 300 to Tm K and Tm to 20 K where Tm is the intermediate stage temperature. CYCLIC ANALYSIS Cyclic analysis for single stage Stirling cycle cryocooler has been reported by Martini [1], Walker et al. [21 and Atrey et al. [31. The cyclic analysis starts with the calculation of instantaneous pressure, instantaneous volume in the working space for different crank positions to calculate P-V areas for compression and expansion spaces. The area under P-V diagram gives the ideally required work for compression and also the ideally produced refrigerating effect for the expansion. The next step is to calculate various losses in the system to evaluate the net refrigerating effect available and the net energy supply to the system. The two stage Stirling cycle as suggested by Prast [4] consists of two Stirling cycles combined together whereby one cycle compensates partly for the losses of the other, so that lower temperatures can be reached. The assumptions are similar to single stage cyclic analysis reported by Atrey et al. [3] with suitable modifications as applicable to the two stage Stirling cycle. Pressure-Volume Variations The volume variations in compression space, intermediate expansion space and final expansion space are sinusoidal because of piston and displacer movements. The phase difference between them is a. The complete cycle is considered to be split into "n" equal intervals. 373
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The procedure of calculating instantaneous pressure and volumes in compression and expansions spaces and instantaneous temperature of gas in the compression space is identical to that of Atrey et al. [3]. The pressures and the temperatures at the 1st and (n + 1)th points should match within the tolerance limit to indicate repetition of cyclic process. As all the points are equi-spaced on the P-V curve, and the nature of the curve is symmetrical, the mean pressure is given by:
Pm-
Pro,at_- ~.P(I)
(i)
Where Ptotal is summation of all the pressure values in the working space at different intervals of the cycle. The average existing pressure of the system, Pavg, is as specified by the manufacturer and the ratio of Pavg and Pm would give the correct value of MR. Initially (MR) was assumed to be unity. Thus with the new values of MR, all the n + l pressure values calculated earlier, are corrected by multiplying by a factor equal to (Pavg/Pm) so that Pm would be same as Pavg. Mass Flow Rates The mass flow rates in the three working spaces would be different because of the cooler, two regenerators and two expansion dead spaces. To calculate these mass flow rates, variation of mass fraction of the gas in each space for each interval is considered. Ideal Power Input and Refrigerating Effect of Both Stages The ideal power input to the system, PI, the refrigerating effects, at intermediate stage (Stage I) RE1 and final stage (Stage II) RE2 respectively, are calculated as the pressure variations for the complete cycle are known. Ideal refrigerating effect (Stage II)
RE2 - / P(I)dV E2(I)
(2)
Ideal refrigerating effect (Intermediate Stage)
RE1 = / P(I)dVEI(I)
(3)
The integration can be done by the trapezoidal rule. The pressure taken for each interval would be the average pressure of the interval. The algebraic sum of the product of pressure and volume differences for each interval would give the magnitudes of PI, RE1 and RE2. Loss Analysis The loss analysis of the cryocooler is very important as it affects the performance of the cryocooler. There are various losses in the system because of which the net power supply increases and the net refrigerating effect decreases. The total pressure drop for the whole cycle is calculated. This power loss increases the required power input. The mechanical efficiency takes various transmission losses into consideration. A mechanical efficiency of 80% and motor efficiency of 89% are assumed for calculating the net power requirement for the cryocooler.
RESULTS Table 1 gives values of ideal and actual refrigerating effects of stages I and II and the power input, The various losses responsible for reducing the refrigerating effect and increasing ideal power are also shown. The results are obtained for three average pressures, namely 29, 29.35 and 30 bar respectively.
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Table 1 Performance of two stage Stirling cycle cryogenerator at various average pressures Average Pressure (bar) =:~
30.00
29.35
29.00
Ideal Refrigerating Effect - Stage I, W Pumping Loss Stage I, W Regenerator Loss Stage I, W Shuttle Loss Stage I, W Conduction Loss Stage I, W Temperature Swing Loss Stage I, W P-V Loss Stage I, W Net Refrigerating Effect Stage I, W
588.67 38.82 86.65 20.86 7.71 182.62 15.00 237.01
575.82 37.94 84.08 20.86 7.71 174.79 14.68 235.76
561.53 37.41 82.71 20.86 7.71 166.23 14.55 232.06
Ideal Refrigerating Effect - Stage II, W Pumping Loss Stage II, W Regenerator Loss Stage II, W Shuttle Loss Stage II, W Conduction Loss Stage II, W Temperature Swing Loss Stage II, W P-V Loss Stage II, W Net Refrigerating Effect Stage II, W
240.33 42.72 41.08 5.92 1.13 10.08 92.23
235.12 41.25 40.15 5.92 1.13 45.15 9.92 91.60
232.32 40.46 40.14 5.92 1.13 44.40 9.83 90.44
7244.10
708 7.15
7002.64
141.03 184.14 6.07 48.07 4.43
138.57 183.80 5.96 47.04 4.35
137.26 183.70 5.91 46.48 4.31
383.74 9534.80
379.72 9333.58
377.66 9225.37
47.17
Ideal Power, W Power Loss due to: Pressure Drop in Pressure Drop in Pressure Drop in Pressure Drop in Pressure Drop in
Cooler, W Expansion Space I, W Regenerator Stage I, W Regenerator Stage II, W Expansion Space II, W
Total Power Loss, W Net Power (Mech 7/= 80%), W
Table 2 gives the comparison of the refrigerating effect at intermediate and final stages of expansion as well as the power input reported by Gao et al. [5], the present analysis with electrical motor efficiency equal to 89% and experimental results of Gao et al. [5]. Table 2 Comparison of the results with the reported results
Calculated value [5] Practical value [5] Present Analysis Present Analysis
Power, k W
RE22, W
REll, W
Pmean, bar
10.78 11.00 10.71 10.48
93.50 90.00 92.23 91.60
238.56 250.00 238.01 235.76
29.35 30.00 30.00 29.35
CONCLUSION
The comparison of the results show that the results obtained by the present analysis are in good agreement with experimental results as reported by Gao et al. [5].
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REFERENCES
Martini, W., Stirling cycle design manual NASA Report No. CR (1978) 135-382 Walker, G., Weiss, M., Fauvel, R., and Reader, G., Microcomputer simulation of Stirling cryocoolers Cryogenics (1989) 29 846-849 Atrey, M. D., Bapat S. L. and Narayankhedkar, K. G., Cyclic simulation of Stirling cryocoolers Cryogenics (1990) 30 341-347 Prast, G., A gas refrigerating machine for temperatures down to 20K and lower Philips Tech. Rev. (1965) 26 1-11 Gao Xiang Yuan and Wu Yu Yuan, Analysis of calculation of multistage Stirling refrigerator cycles ICEC 12 (1988), 539-543.
Theoretical and experimental investigations on effects of charge pressure on performance of linear motor driven miniature cryocooler Patel L. N.* and Narayankhedkar K.G.** *S.V.R. College of Engineering and Technology, Surat 395 007, India. **Indian Institute of Technology, Bombay, Powai, Mumbai 400 076, India.
In Electro-Magnetically Driven Cryocoolers (EMDC), a linear motor drives the piston-coil assembly but the movement of the displacer is wholly controlled by the pressure gradient established in the system. Thus the thermodynamic behaviour is coupled to system dynamics. The influence of charge pressure on piston to voltage phase shift, stroke of reciprocating elements and ultimately on the cooling capacity of EMDC is discussed in this paper. The results obtained from the present work are compared with those reported in the literature. Experimental investigations show that the cryocooler performance is sensitive to the charge pressure.
INTRODUCTION
Mechanical cryocoolers have poor reliability mainly due to greater rate of wear and tear, and contamination of the working space. A Free Piston Free Displacer (FPFD) cryocooler is linearly driven either pneumatically or electromagnetically. The moving component is suspended on a mechanical spring which determines its mean position. The FPFD concept has far fewer moving parts and thus increased reliability. Since the driving force is axial, virtually no side loading occurs on the piston or displacer. This eliminates the need for oil lubrication, thus ruling out a major cause of contamination. Seal duty is also substantially reduced and this enables the use of close tolerance dry seals in place of pre-loaded contact seals which is the cause of friction and wear. Thus, the FPFD concept not only preserves to a large extent, the inherent efficiency of the Stirling cycle b u t also increases reliability and life. It is worth noting that while in the case of conventional cryocoolers the strokes and kinematic phase-shift between piston and displacer are fixed, all of these quantities are variable in the case of FPFD cryocoolers. This lends a given FPFD cryocooler to better control and its experimental optimization is easier, as no dimensional changes are required in order to test the influence of variation in strokes and phase-shift. However, introduction of three more variables makes the FPFD cryocooler much more difficult to analyze, theoretically optimize and design. The design of a conventional cryocooler is basically a problem of thermodynamics and heat transfer modelling. The design of an FPFD cryocooler involves additionally the integration of kinematic and dynamic equations of the motions of the piston and displacer into the thermodynamic and heat transfer equation set. de Jonge [1] reported analysis of moving coil linear motor and a d y n a m i c - thermodynamic analysis of an integral Stirling cryocooler. A similar model but considering coil inductance, has been reported by Narayankhedkar et al. [2]. A considerably improved mathematical model for the linear motor with a single coil has been reported by de Jonge and Sereny [3]. 377
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CHARGE PRESSURE The phase shift (r decreases with increase in charge pressure. The increase in charge pressure increases the loading (Ap x pch) and this causes an additional lag of the piston movement. The variation in r with respect to pch could be explained with the help of Equation (8). An increase in charge pressure increases the mean effective pressure. This results in higher values of pressure position coefficients. This further changes the two constants C1 (decreases) and C2 (increases) in terms of which the phase shift r has been defined in Equation (8). The net effect is the decrease in with increase in charge pressure. The amplitudes of both, the piston as well as the displacer pass through their optima. Experimental data by de Jonge [1] for amplitude of displacer (Xd) compare well with the present theoretical results. The plot in Fig. 2 explains the performance trend of the linear motor at varying charge pressure. It is observed that the motor output power and efficiency go through their maxima whereas the motor input goes through its minima. This is because lower the charge pressure, higher is the phase shift. With increase in charge pressure, r decreases. For the charge pressure that results in = 92.5 deg, the motor output is maximum and input power to the linear motor is minimum. This gives maximum motor efficiency. On either side of r = 92.5 deg, output power is less and input power is more. Experimental data for motor input reported by de Jonge [1] and the results from the present analysis reasonably agree. Fig. 3 shows the plot of optimum theoretical refrigeration versus charge pressure. It is observed that the theoretical refrigeration goes through an optima. Fig. 4 shows the theoretical cold production as a function of frequency ratio. It is maximum for wd/w 1.25. For this value of wd/w,r works out 45 deg. 9
!
.
|
-
!
9
!
~
80 • X
60
5
do J o n ~ 1] X
Bg = 0.33 T ~
Pt, m (W) 3
" 40
Rt = 5.2 o | u n | /
"
(W)
2
X
X
x 9~ 20
• NN, de Jonge [ 1]
• 10
0
,
5
I
,
,
,o
I
1;
2;
6
o=
g 1,0
1,2
~
1,4
25
Figure 3 Effect of charge pressure on refrigeration capacity of eryocooler
Charge pressure, Pch (bar) Figure 2 Influence of charge pressure on motor performance
O
2
Charge pressure, Peh (bar)
,
20
1
0 = 45 ~ co = 314 rad/s x=4 K=7.7 Tch = 300 K Pch = 16 bar
1,6
1,8
2,0
2,2
Frequency ratio, rod/co Figure 4 Effect of frequency ratio on refrigeration capacity of cryocooler
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ICEC16/ICMC Proceedings DYNAMIC-THERMODYNAMIC ANALYSIS
Narayankhedkar et al. [4] reported a dynamic model of electro-magnetically driven Stirling cycle miniature crycooler and thus formulated expressions for the instantaneous gas velocity and pressure fluctuation coefficients based on the concept of Reynold's friction factor. Dynamics of Piston-coil Assembly Figure 1 shows the free body diagram of the piston-coil assembly. Various forces acting on it are i) accelerating forces, Mp Sp ii) gas force, ( p c - p ) A p iii) spring force, kp xp iv) electromagnetic force, Bg Lv i. Hence the dynamic equation, neglecting the damping effect is Mps
(1)
+ k p z p + (pc - p ) A p = B g L v i
Thus the following expression is obtained (C1 cos ~ + C2 cos ~)xp - -B~L,,Eo R(I+L2) [cos ~(cos r + L sin r - sin ~(sin r - L cos r
(2)
Where
(3)
~ =a-8 C1
-
Mpw 2 - ApZp-
kp-
BgLvGL
XvR(C~+L2)
- A p Z p F p d cos 2 8
(4)
BgLvG
(5)
C2 - A p Z d F p d cos 8 sin 8 + XpR(I+L2) G = BgLvXpw ,
f-Mp A p
.._.-
s ~
t
! w. d
'E
(a)
FIIEE
8OOY I)IAGIIAt4
R
OF
PlS|Otl COIL A55Et48LY
( b ) I:REE [JOI)Y I)IAGRAt4 OF
I)ISPLACER
Figure 1 Free body diagrams of reciprocating elements Amplitude of Piston The amplitude of the piston-coil assembly is determined from the fact that power input to the piston is equal to the output of the linear motor. Po, m = Pi, i0
(7)
After obtaining expression for the piston stroke, expression for phase shift between piston and impressed voltage is given by
r -
Lc
-c
GI+C2L
(8)
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CONCLUSION Theoretical as well as experimental performance investigations showed that the performance of cryocooler is sensitive to the charge pressure. Best performance is obtained at the charge pressure for which the stroke of piston and piston to voltage phase shift are at their optimum values.
Nomenclature Ap Bg C1 C2 Eo Fpd G i kp L Lv Mp Po,m Pi,p
: : 9 : : : : : : : : : : :
Piston cross sectional area (m 2) Flux density in magnetic gap (T) Constant defined in equation (4) Constant defined in equation (5) Amplitude of impressed voltage (V) Flow loss coefficient ratio Constant defined in equation (6) Current (A) Piston spring stiffness (Nm -1) Coil quality factor Active coil length ( m ) Piston mass (kg) Power output of motor (W) Power input to piston (W)
pc
p R Xp xp ~p Zp
: Compression space pressure : Mean pressure (Nm -2) : Total resistance (w) : Piston amplitude (m) : Piston displacement (m) : Piston acceleration (ms -2) : Pressure position coefficient
(Nm -2)
(Nm -3)
Greek
r 0 tO
:
cot
: : : :
Piston to voltage phase shift Piston to displacer phase shift As defined in equation (3) Angular frequency (fads-1)
REFERENCES de-Jonge, A. K., A small free piston Stirling refrigerator Proc. of 14th IECEC Am Chem Soc, Washington Dc, USA (1979) 1136-1141 Narayankhedkar, K. G., Nagaraja, N., Bapat, S. L. and Patel, L. N.,Analysis of linear motor for electromagnetically driven cryocooler Proc. of International Conference on Cryogenics, Macmillan India Ltd.,New Delhi, India(1988) 284-292 de Jonge, A. K. and Sereny, A., Analysis and optimisation of a linear motor for the compressor of a cryogenic refrigerator Advances in Cryogenic Engineering(1982) 27 631-640 Narayankhedkar, K. G., Nagaraja, N., Bapat, S. L. and Patel, L. N., Dynamic model of electromagnetically driven Stirling cycle miniature cryocooler Cryogenics(1990) 30 231-235
DEVELOPMENT OF MICRO-STIRLING COOLER Norihide Fujiyama, Tadashi Kosaka, Takazo Sotojima, Nobuaki Yagi Daikin Industries LTD., Sakai Plant, 1304,Kanaoka-cho,Sakai,Osaka 591,JAPAN A split Stiding micro cooler with a compressor provided with dual opposed pistions driven by linear motor has been developed. The weight of the cooler is 420g and the occupation volume is about 250cm 3 Operating frequency was 50Hz, and the charged pressure of helium gas was 1.3MPa. For this cooler, 210mW of the cooling capacity at 80K was obtained at the nominal input power of 8W, and 400mW at the maximum allowable input power of 15W. The cooling capacities measured by replacing 10 expanders for the same compressor showed good reproducibility.
INTRODUCTION Recently, in the various fields of applications such as the temperature measurements, the image watchings, the quality controls and the medical diagnostics, the need of an infra-red imaging camera in commercial use has been increasing, and in order to achieve its high sensitivity and high resolution performances, the requirements of a micro-cooler which cools the infra-red sensors upto 80K of a temperature level has been also increasing. Especially a portable imaging camera mainly used for the temperature measurements has a posibility to aquire an extensive market because of its convenience. The demands to the refrigerator used for such a camera are the small size, the light weight and the low power consumption, in addition the long life is also getting important requirement. We have been developing as a target of this use a split Stirling refrigerator with a compressor having dual opposed pistons driven by linear motors which we considered that its principle is fittable to the long life need hear after. THE SPECIFICATIONS OF THE COOLER Table 1 shows the main specifications of the proto type cooler we have developed recently. The reason why we have chosen a split type and also a type of a linear motor driven compressor is that this type of a cooler had been considered to be able to take a fully oil less configuration and to fit the future requirement of a long life. On the other hand, in the case of the integral type Stirling cooler which uses the crank shafts to drive the piston and the displacer, it cannot be evitable to use oil or grease in order to seal and to lubricate the crank shaft of the piston simultaneously. As long as using oil or grease we considered that the problem of the refrigerant contamination could not be solved essencially, and resulted in making the limitation to elongate the life of a cooler. The design goals of the cooling capacity of the cooler was established 150mW under the consideration as follows. As the demand of small size, light weight and low power consumption is very strong on a portable infrared camera, it is necessary to reduce the load to a refrigerator so as to use a smaller refrigerator. So in this type of camera the infrared sensors are usually attached directly to the cold head of a refrigerator, and the capacity of a refrigerator is sufficient with 150mW for this type of a camera. With regard to the weight and the power consumption we settled the design goals to be 450g and 8W respectively as the value allowable standing on the needs of the market and approachable by the latest technology. THE STRUCTURE OF THE COOLER Figure 1 illustrates the structure of the split Stifling cooler we have developed.The cooler is hermetically sealed at all. Outlines are as follows 9the compressor is of 35.4mm in diameter and 89mm in length, and the expander is of 5.3mm in diameter and 94mm in length, and the occupation volume of the refrigerator is about 250cm 3. 381
382
ICEC 16/ICMC Proceedings Table 1. The Specifications of the cooler II I IIII II
Type
Split- S tirlin g
Nominal Operating Temperature Total Weight
<
3 < 300 cm "~( MTBM8000hrs )
Volume Reliability Nominal Cooling Capacity Nominal Input Power Maximum Cooling Capacity Maximum Input Power I
80 K 450 g
=> 150mW 8 W _~ 400mW 15 W IIIll
~9 The design goal (not measured) Refrigerant is helium gas, and the charged pressure was determined 1.3MPa finally as an optimum value for this cooler by experiments. We have chosen the type of the dual opposed pistons so as to eliminate vibrations. The operating frequency was settled at 50Hz. The stiffness of the mechanical spring of the piston has been determined 1500N/m by experiments. We employed dry polytetrafluoroethylene (PTFE) bearings and clearance seal technology for each piston. Moving mass design allows for a very simple, light weight and compact motor which requires no additional mechanisms or guides to support a moving mass. In order to use the narrow gap space efficiently we employed no bobbin coils as the driving coils. The stroke of the individual piston is 5mm at the nominal input power of 8W, and the allowable maximum input power is 15W where the dual opposed pistons don't collide each other.The inner diameter of the connecting tube is 2mm, and are connected to the compressor. We contrived a simple construction at this part and achieved to make a light weight cooler using a copper gasket at this part. We employs 316L Stainless Steal for expander cylinder. Its thickness is 0.1mm, it is mechanically processed to a minimum as possible so as to reduce the heat conduction loss. The cold head which is made of copper, the cylinder and the fin are brazed in vacuum.The stainless steel parts are TIG welded. The displacer is made of epoxy resin, and the regenerator mesh is #180 and its line diameter is 30~m. The number of regenerator mesh is about 1200. We employed dry PTFE bearings for the displacer. The specific frequency in displacer mass-spring system is 1.4times as large as the operating frequency. Compressor Unit Electrical connecting pins ~~
Permanent magnet
./Hausmg --Spring
-
Drive coil\
Expander Unit
] ~ .-, . CopperFin F IVLI ~ompress~on t ~ JAil.-Volume ~fl flflfl[Ifl[~
Cylinder1
lull[
Dual Piston
.
Displacer
-o.
r
Coldhe
,/
_
i
C~ Tube
-
Regeneratormesh ~ ~
I1 Spring ~
/
Ex ansion Volum4 P
I
[~.., I ~
~35.4
,._1 ~
I
Figure 1. The structure of the prototype micro Stirling cooler
(Dimensions in mm)
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EXPERIMENTAL PROCEDURE Schematic diagram of measuring system of a cooling capacities is shown in Figure 2. At the cold head, a load heater and a C-C thermocouples were attached to the cold head. The expander was covered by super insulation in order to reduce ~ e radiation heat loss, and was mounted to a dewar. The pressure in the dewar was kept less than lxl0-"Torr by using a turbo molecular pump. All experiments were performed in a room airconditioned at 293K. The measurements of the cooling capacity was made by measuring the input power applied to the loading heater while the temperature of the cold head was kept constant. The measurements of the input power applied to the cooler was made by taking indicated values of the power meter. To avoid the contamination of refrigerant, the parts of the compressor unit and the expander unit were baked at 373K for 72 hours in vacuum. After assembling the cooler, it was baked at 373K for 24 hours in vacuum. After this the cooler was charged by helium gas. Compressor u n i t A.C. Power source I PCR500 / I
, j
I
Expanderunit Fan
Dewar Super Insulation C-C Thermometer I thermocouplelAdvantestl i 2114H
Power Meterl YOKOGAWA~-~| 2533 ~ 1
]
I( i i IAdvantest/ Heater I ] TR6143 [~ gem~ D.C.source[, Vacuu IMr'~" Turbo Molecular / Gau 150 Pump V,A,W |
frequency,V,A,W PC-9801 , Personal Computer
Temperature
Figure 2. Schematic diagram of the measuring system of cooling capacities
EXPERIMENTAL RESULTS Figure 3 shows the measured cooling capacities of the proto type cooler on three different cold head temperatures at the ambient temperature of 293K. At the cold head temperature of 80K, the proto type cooler had a cooling capacity of 210mW over the design goal when the input power was nominal. And a maximum cooling capacity at cold head temperature of 80K was above 400mW when the input power was maximum. The cool down time from the room temperature of 293K to 80K was about 6 minutes and the ultimate cold head temperature was 51K without loading. The cooling capacities of 10 expanders were measured connecting with a same compressor. Figure 4 shows the results. All measured values had a good reproducibility. Photo 1 shows the proto type micro Stirling cooler. SUMMARY 1. We have developed a micro Stirling cooler, which is 420g in weight and 250cm 3 in volume for an application of a portable infrared camera. 2. We have chosen a split type and a linear drive compressor with dual opposed pistons which we considered to be fittable for long life. 3. The cool down time from 293K to 80K was about 6 minutes, and the ultimate cold head temperature was 51K, and the cooling capacity at nominal input power of 8W was 210mW. 4. We will evaluate the life of the cooler.
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ICEC16/ICMC Proceedings 400
.
.
.
.
.
.
350
_.C.(l.d._H.e~il........~ 85K...:: Teinperatiare 9o 80K
220
.
.
.... ..../~ .
.
.
.
.
.
215 ~
.
210 o
.,, ~~, 25o .
........................ ~ .
~:
o1"'4
~ 2oo ~
_
o
205 o
o
o
o
~ 195
........................
"~=o 100150_ ~ ~ . ~ ~ ............................... . . . . . . .i. li U
i
' Ami)ient ~mperitture : Z93K ............................
50 i
5
~ 190
........................... ',............. -'~'"-"'~"~: . . . . . . . . . . .
r,,) l
6
9
7
9
9
8 9 10 Input Power (W)
9
9
11
12
Figure 3. The coolong capacities of the prototype cooler as a function of input powers
Input Power" 8W Cold Head Temperature" 80K Ambient Temperature" 293K
185 180
9
12
I
9
34
I
9
56
"
9
78
"
910
The serial numbers of replaced displacers Figure 4. The cooling capacities of the cooler measured by replacing expanders only
Photo 1. The proto type micro Stirling cooler
Computer-Aided Design of Split-type Stirling Refrigerator B.J.Huang and H.Y.Chen Department of Mechanical Engineering, National Taiwan University, TAIWAN S.B.Chien and T.F. Shieh Chung-Shen Institute of Science and Technology, TAIWAN The design of split-type Stirling refrigerator mainly relies on engineering experience. In the present study a computer-aided design tool "STC15b" which can be run on PC was developed for the design of split-type flee-displacer Stirling refrigerator. The analytical solution resulted from a linear network model was used in the performance calculation. An empirical relation for the displacer loss coefficient Ca was obtained. The cold-end temperature prediction at various cooling capacity and operating frequency is shown very accurate.
INTRODUCTION
The design of split-type flee-displacer Stifling refrigerator mainly relies on engineering experience. Many researchers attempt to develop a computer-aided design tool using various analysis. The achievement is however quite limited since a complicated numerical method is always involved and the computation requires a super computer for solving the transient governing equations of the various parts of the Stirling refrigerator. In the present study we develop a computer-aided design tool which can be run on PC. The analytical solution resulted from a linear network model was used in the performance calculation. The computation is thus fast and can be done on a PC. To improve the accuracy, the analytical errors due to the unmodeled factors are corrected by an empirical relation collected from field tests. LINEAR N E T W O R K MODEL OF SPLIT-TYPE STIRLING REFRIGERATOR
Stirling refrigerator operates at a cyclic state and the physical process is thus transient. A linear network model is thus developed in the previous study [1] for the system analysis of split-type Stirling refrigerator (Figure 1). The linearly-perturbed models of the components are derived from the governing equations through linearization and approximations. Connecting the equivalent circuits of the components according to the process of Stirling refrigerator, we obtain an equivalent network of split-type Stifling refrigerator as shown in Figure 2. Block diagrams are drawn for the connecting tube and the regenerator in Figure 2 since they belong to the distributed systems. Solving the linear dynamics equations, we obtain two transfer functions of the Stirling refrigerator: Xd (s) [ G3(s)G8(s)/G7(s)+G2(s)][Imp(s)Zti+Imm(S) l+ZtJZ~ ][ sp~R 7~ Ca (s) - (1) Xp (~) Pe (s) a2(s)a6(s']} G ep(S) = --k/G4(S)+ 1-G5(s) G7(s) + 1-Gs(s) • Xp (~) (2) l +Zti/Zc X R Tc G8(s)+G7(s)Zcttanh[F tLt] . where Zti = G7(s)+[G8(s)/Zct]tanh[FtLt]' G1 (s) = Dew(S)+Dde(S)Rpp(S); G2(s) = Dde(S)Rpm(S); G3(s) = G1 (s) + G2(s)Wmp(S); G4(s) - Rpp(S) + R(s)mmp(S); G5(s) - G2(s)mmd(S); G6(s) - Rpm(S)mmd(S); G7(s)ptoS - G8(s) mto (s); Wmp--sVwo/(R ~;'w); Wmd--spoAaw/(R [I'w); Z~ - R T~ /(sV~o); 385
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ICEC16/ICMC Proceedings
[mmd(S)G2(S)]- Eme(S) [ Rpm(S) + G2(s)G6(s)]
G8(s) = Rmm(S) 1 +
mti(s)-
1-G5(s)
Emd(S)G2(s). --
1-G5(s)
'
spoAp/(R Tc ) "" 1-+-~/7_~ X p ( S ) .
It is found that the piston displacement Xp is the system input of a split-type Stirling refrigerator. The system outputs are the displacer displacement Xd and the expansion space pressure Pe. The Stirling refrigerator thus belongs to a single-input-multiple-output (SIMO) dynamical system. The maximum available cooling capacity can be evaluated by integrating the pressure peand volume Ve of the expansion space, where V~ is related to Xd. Since the piston motion as well as the associated pressure waves approaches sinusoidal, Xa and pe can be computed simply from the gain and phase of the frequency response functions, Gdp(jO~) and G~p(jO~). Assuming that adp is the gain of Gdp(jO~) with phase ~d leading the piston; aep is the gain of G~p(jO~)with phase ~e leading the piston. Then, the maximum available cooling capacity is
Qmax- f ~2'~pedV~ =fxa,paapXRpoA aeSin(d~d- dpe)
(3)
where Xpo is the amplitude of the piston; f is the operating frequency; ~ is the piston angle; P,o and Xao are the amplitudes of pe(t) and Xa(t), respectively. The net cooling capacity Qnet c a n be evaluated by subtracting the heat losses from the maximum available cooling capacity Qmax. There are four types of heat losses: namely, heat conduction loss of regenerator Q~o,,a , enthalpy flow loss of regenerator Oenth , shuttle heat loss of displacer Q,h,ttz~ , and hysteresis loss of spring Wir [2].The heat loss due to the gas leakage through the clearance between the displacer and the cylinder wall is related to the seal design, manufacturing process, and material used. In the present analysis, a loss coefficient Cd is included. Cd is the coefficient accounting for the friction and gas leakage of the displacer. Since the theoretical prediction of Cd is very difficult, Cd is analyzed from test results. COMPUTER-AIDED DESIGN TOOL - - T H E PACKAGE "STCI5b"
A computer package "STC15b" was developed in the present study. The package whose configuration is shown in Figure 3 includes three parts: subroutine STC15A, STCCd and STC15s. STC15A is used to evaluate Cdfrom the test results; STCCd is used to derive an empirical correlation for Cd, i.e. Cd(f, TL); STCs is used for the practical design using the empirical relation Cd(f, TL). For design purpose, the subroutine STCs is run to calculate the system performance. After the hardware tests, the subroutines STC15A and STCCd will be run to update the empirical relation Cd(f, TL) in order to obtain a better result than before. The package "STC15b" is run on PC. The computation time takes a few seconds for each design. RESULTS
Some split-type Stirling refrigerators were designed and built in the laboratory for experiments. An empirical relation was obtained: Cd(f, TL): F(TL) • G(/) (4) where F(TL) - 430.66 + 1.0851 T,~ + 77346•176 T~ ' 9G(]) - 0.6653 +0.2604f- 5.212 • 10-4ja The design of a Stifling refrigerator using the package "STC15b" can then be carried out. Shown in Figure 4 is the PV diagram of the expansion space computed from the package.; Figure 5 is the cooling capacity versus frequency; Figure 6 is the cooling capacity variation with the cold-end temperature. The prediction of cooling capacity using "STC15b" is satisfactory as shown in Figure 6. It can be seen from Table 1 that the errors of the cold-end temperature prediction at various cooling load are mostly within_+_+3K.
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CONCLUSION It is shown in the present study that the design process of a split-type Stirling refrigeratoris can be greatly simplified by using the software package "STC15b" The package is also designed to have a learning capability in order to accumulate the field experience and update the program automatically.
Acknowledgment
The present study was supported by the National Science Council, Taiwan, through Grant No. CS84-0210-D-002-027.
REFERENCES 1.Huang B.J. and Lu,C.W. Linear network analysis of split-type Stirling refrigerator". Cryogenics (1994)34 ICEC Supplement 207-210 2.Urieli, I. and Berchowitz, D . M . "Stirling Cycle Engine Analysis". Adam Hilber Ltd., Bristol, UK. 1984 T a b l e 1 C o m p a r i s o n b e t w e e n test results and the design calculation using " S T C 1 5 b " .
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Microcooling of Low Temperature Electronics
H.J.M. ter Brake, J.F. Burger and H. Rogalla University of Twente, Faculty of Applied Physics, PO Box 217, 7500 AE Enschede, The Netherlands
Small scale low temperature electronic applications would largely benefit from the development of a closed-cycle microcooler, with a cooling power in the milliwatt range. We started a microcooling project, and the first cooler configuration under study is the Joule Thomson cycle in combination with sorption compressors. A sorption compressor set-up was built, and a thermodynamic model was developed. Model and experimental results were in good agreement and can be used for future research.
INTRODUCTION Low temperatures provide an excellent operating environment for conventional and superconducting electronics. For conventional electronics, it increases the speed of digital systems, it improves the signal to noise ratio and the bandwidth of analog devices and sensors, and it reduces aging. For superconducting electronics, it is an essential operating condition. There is a broad range of applications making use of cold electronics, such as computers, amplifiers, mixers, fast AD and DA converters, IR detectors and SQUID magnetometers [1]. Many of these systems need only a modest cooling power, as low as 10 mW or even lower. Standard cryocoolers are largely oversized for such devices, and the development of a microcooler would be of great use for low temperature electronics. Walker [ 1] has given an overview of small scale cryocoolers. Little [2] scaled down the conventional Joule Thomson cold stage using photolithographic techniques for patterning gas channels in glass layers. This cooler, however, has the disadvantage that high pressure gas is needed from a relatively large storage bottle. To our knowledge, the smallest closed cycle cooler realised so far is manufactured by Inframetics [3]. They make integral Stirling coolers of 0.3 kg weight, 9 cm maximum size, and 0.15 W cooling power at 80 K with an input of 3 W. A further reduction in size will be limited by the manufacturing techniques and in this respect micromechanical techniques have been suggested as an alternative [4]. We recently started a project with the aim to realise a closed-cycle microcooler. The project is carried out in cooperation with the MESA Research Institute because of their experience in micromechanical technologies. These technologies use special deposition, etching and bonding techniques - originating from the IC,technologywith which very small mechanical structures can be made, in or on planar substrates (mainly silicon and glass). A typical example of a microchannel structure is given in figure 1 [5]. This paper describes the microcooling project. The main requirements and some design considerations for microcooling are given: After that, a sorption cooler is described, and modelling of a sorption compressor is compared with test experiments. Figure 1 Deep anisotropic etching in silicon [5]. This research is supported by the Dutch Technology Foundation (STW). 391
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REQUIREMENTS FOR MICROCOOLING Requirements for microcooling of low temperature electronics can be summarised as follows: 1. Working temperature: 60- 80 K. Many HTc superconducting devices function at LN2 temperature. The temperature of conventional electronics is not very critical. In general, it should not be lower than 40 K due to carrier freeze-out below that temperature. Therefore, LN2 temperature is very suitable for many applications, also for hybrid electronics. 2. Cooling power (net): 5 - 50 mW. The required cooling power is determined by three factors: the power dissipated in the cold device, the heat leakage from the environment and the desired cool-down time. The power dissipated by superconducting electronics is very small, even a microprocessor with thousands of junctions on a substrate of 5 mm x 5 mm may dissipate less than a few milliwatts [6]. Dissipation of conventional electronics is more severe, therefore only small circuitry can be cooled. The heat leakage is determined by conduction through the supporting structure and the wires, and by radiation. Conduction through the wires does not play a major role. Radiation at 60 - 80 K can be reduced to roughly 1 mW via appropriate shielding. The cool down time is determined by the combination heat capacity - cooling power. Furthermore, the required gross cooling power will be much higher in order to deal with losses by conduction through the cooler itself. 3. Closed cycle with long life-time. We aim to realise a life-time of several years. This puts restrictions on the use of moving components, especially if friction is present. 4. Low electromagnetic and mechanical interference. 5. Other issues: fast cool down time, low heat rejection (i.e. high efficiency), small size, low price.
DESIGN CONSIDERATIONS Commonly used cryocoolers can be divided in recuperative (JT, Brayton) and regenerative (Stirling, GM, pulse tube) cycles. Most of these cycles use the reversible expansion of an ideal gas, performing work on the environment and taking up heat at low temperature. Only Joule Thomson expansion is an irreversible process, performing internal work on a non-ideal gas (no moving parts at the cold side). Interesting alternative cycles worth considering are: thermo-electric cooling and the desorption of a gas from a solution. In the field of micromechanics much experience has been obtained in the design and realisation of fluid and gas handling systems. Also, all kinds of elastic membranes and actuation principles have been realised. An important design constraint is that rubbing surfaces, occurring in nonflexible constructions such as pistons and hinge points, are highly undesirable in microtechnology. Some considerations with respect to the miniaturisation of cooling cycles" 1. Stirling. Miniaturised membranes may be an attractive alternative for the use of pistons, but the relatively low pressure differences require highly efficient regenerators. Interferences are likely to occur at the cold side of a Stirling cycle. Bowman et al. [7] recently patented an idea for a microminiature Stirling cooler with membranes running at a high frequency (500 Hz or more). 2. Pulse tube. This cycle is very interesting because no moving parts are used at the cold side. Membrane actuation may be possible. A patent of Cabanel et al. exists on a micro pulse tube [8]. 3. Brayton. A very small turbo expander requires a very high rotation speed, and seems not feasible. 4. Joule Thomson. Little [3] has proved that miniaturisation is possible. However, a small reliable compressor being able to generate high pressure differences is a major requirement for closed cycle operation. A possible candidate for this is a sorption compressor, and the microcooling project therefore started with research in this field.
SORPTION COOLER Basic operation Figure 2 gives the set-up of a sorption cooler [9]. It consists of a compressor unit, a counterflow heat exchanger, and a Joule Thomson expansion valve. The compressor unit contains four sorption compressors and several check valves to control the gas flows. Figure 3 gives a sketch of a sorption
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compressor. Low and high pressures are generated by cyclic ad- and desorption of a working gas on a sorption material, which is accomplished by cooling and heating of the sorption material. Cooling is done with a gas gap heat-switch between the sorption material and a cold source on the outside. A compressor cycle is schematically drawn in figure 4. The compressor is heated during sections A and B, and cooled during C and D. During sections A and C both valves of the compressor are closed, and the compressor is in a regenerating phase. During sections B and D one of the valves is opened; the compressor generates a high pressure gas flow during B, and a low pressure gas flow into the compressor during D. Except for some check valves, the cooler has no moving parts in it. This minimises electromagnetic and mechanical interference, and it contributes to a long life time. Moreover, distributed spot cooling can be realised very well with several small micromachined JT coolers. Important aspects in the development of a very small sorption cooler are: the selection of gas/sorber combinations that are efficient around ambient temperature, the choice of a gas or a gas mixture with a high JT efficiency [ 10], the development of small compressors with an efficient heat management, and the development of small micromachined high pressure check valves. Compressor model and test experiments For predicting the behaviour of a miniaturised sorption compressor, a thermodynamic model for a cylindrical compressor was developed, as well as a test set-up for model validation and gas/sorber research. The compressor consists of five elements: heater, sorption material, inner container, gas gap and outer container - see figure 3. In the model the heat flows between these elements are calculated and used to calculate the temperatures, as a function of time. Also a radial temperature profile in the sorbent is calculated. At each time step the amount of gas adsorbed in relation to that temperature is calculated (using Polanyi's theory [ 11 ]), and by that the pressure in the compressor can be predicted. With the test set-up the following parameters can be determined: the temperatures at four positions in the compressor (see figure 3), the amount of gas that flows into the compressor (and thus the total amount of gas in the compressor), and the pressure in the compressor. The heating rate and the temperature of the heater can also be controlled. Firstly, the test set-up was used to characterise the adsorption of nitrogen on activated charcoal. The theory of Polanyi was confirmed, only at low pressures in combination with low temperatures theory and experiments did not correspond. Secondly, the time dependence of the temperatures and the pressure in the compressor were measured. In figures 5a and b typical results are compared with the model. In this experiment the compressor was filled with a measured amount of gas and cycled from room temperature to a temperature set point of the heater, and back again to room temperature. A good correspondence between the model and the experiments was found. Model and experiments are discussed in more detail elsewhere [12].
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Sorption cooler feasibility Most important for the feasibility of a sorption cooler is an efficient and suitable gas/sorber combination. When such a combination is found, scaling down of the compressor seems feasible. The heat stored in the compressor should be as small as possible, since it is parasitic heat. The heat capacity is determined by that of the sorber and that of the container material, and it can be shown that the ratio of those two is independent of the compressor size. Therefore, the heat required for warming up the compressor scales linearly with its volume. In first order, when the cycle frequency of the compressor is kept constant, the input power and the resulting mass flow (with which the cooling power is proportional) scale linearly with the compressor volume (implying a constant efficiency). Also, essential aspects for the miniaturisation of a sorption compressor are: good thermal conduction within the compressor (from heater to sorption material and from sorption material to heat-switch), and miniaturisation o f the heat switch.
CONCLUSIONS Research was started towards miniaturisation of closed cycle cooling methods. Application of micromechanical techniques for components of a microcooler is also investigated. A miniaturised sorption cooler with JT expansion is a promising candidate for small scale cooling. The developed model of a sorption compressor was in good agreement with measurements. Miniaturisation is feasible, but the availability of a gas/sorber combination that is efficient at room temperature is a crucial requirement. In the future, attention will also be paid to miniaturised check valves, a miniaturised heat switch, and the possibility of miniaturising regenerative cooling cycles.
REFERENCES 1 2 3 4 5
6 7 8 9 10 11 12
Walker, G., Low-capacity cryogenic refrigeration, Oxford University Press (1994) Little, W.A., Microminiature Refrigeration, Rev. Sci. Instrum. 55 (5) (1984) 661-680 Inframetics, Inc., 16 Esquire Road, North Billerica, MA 01862-2598, U.S. Walker, G. and Bingham, R., Micro and nanno cryocoolers: speculation on future development, 6 th Int. Cryocooler Conf. (1990) 363-375 Jansen, H.V., de Boer, M., Burger, J.F., Legtenberg, R., and Elwenspoek, M., The black silicon method II: the effect of mask material and loading on the reactive ion etching of deep silicon trenches, Proc. Workshop on Micro and Nano Eng., Davos, Switzerland, 1994. Kirschman, R.K., Low-Temperature electronics, IEEE Circuits and Devices, 6 (2) (1990) 12-14 Bowman, L., Berchowitz, D.M., Urieli, I., Microminiature Stirling cycle cryocoolers and engines, U.S. Patent 5 457 956 (1995) Cabanel, R., Friederich, A., Refroidiseur ~agaz puls6, demande de brevet european No. 0672 873 A 1 (1995) Wade, L.A., An overview of the development of sorption refrigeration, Adv. in Cryogenic Eng. 37 (1992) 1095-1106 Alfeev, V.N., Brodyansky, V.M., Yagodin, V.M., Nikolski, V.A., Ivatsov, A.V., Refrigerant for a cryogenic throttling unit, UK Patent 1 336 892 (1973) Polanyi, M., Verhandl. Deutsch. Physik., Vol. 16 (1912) Huinink, S.A.J., Burger, J.F., Holland, H.J., van der Sar, E.G., Gardeniers, H., ter Brake, H.J.M., Rogalla, H., Experiments on a charcoal/nitrogen sorption compressor and model considerations, 9th Cryocooler Conf. (1996)
Low Cost Mixture Joule Thomson Refrigerator
A.Alexeev, H.Quack, Ch.Haberstroh Technische Universit~it Dresden, Lehrstuhl ftir K~ilte- und Kryotechnik, D-01062 Dresden, Germany
A closed cycle Joule Thomson Refrigerator for cooling electronic devices, which operates in the temperature range from 70 K to 95 K, has been developed which uses refrigerant mixtures and a single stage oil lubricated compressor. It provides high reliability and no maintenance. This system is compact, has a good thermodynamic efficiency and low levels of vibration and noise. The system and its performance characteristics will be described in this paper.
INTRODUCTION Many future applications of HTSC will be in high frequency electronics techniques. Nowadays thin layer SQUIDs made from HTSC are under discussion for a variety of applications. These include in medical research the measurement of biomagnetic fields (human heart signals etc.) as well as in material science the non destructive testing of sensitive constructions or in geological survey the exploration of the earth natural resources. A cooler should not interfere with the measurement taken and it should have a high reliability. Also handiness and mobility are a necessity for out-of-laboratory applications. The state of the art to this date did not provide a sufficiently good solution. Thanks to recent advances in the development of mixture Joule Thomson refrigerators, this deadlock seems now to be overcome. The commercial application of reliable coolers is within reach to costs which compare to those of the overall system. BACKGROUND The use of multicomponent refrigerants in cycles of Joule-Thomson type in the range of temperatures T>100 K has been known for a long time (W.J.Podbelniak, A.P.Klimenko, M.Fuderer and A.Andrija, D.J.Missimer). Using mixtures in systems, working at temperatures T<100 became possible due to research, which began in the 70ties under supervising of V.M.Brodianski [1]. These technologies were intensively developed at the Moscow Power Engineering Institute and in Odessa. The mixture Joule Thomson refrigerators with oilfree compressors were produced in series in Omsk [2]. At the beginning of the 90ties, due to a more free information exchange, the mixture technologies spread out to other countries too. The APD Cryogenics Inc., in a collaboration with the group of M.J.Boiarski from Moscow Power Engineering Institute developed a mixture Joule Thomson refrigerator on the basis of an oil lubricated compressor with a subsequently connected adsorber. This allowed to increase the maintenance free time to over 20,000 hours [3]. MMR Technologies Inc. also developed a prototype mixture refrigerator [4]. MIXTURE JOULE THOMSON REFRIGERATOR Figure 1 is a schematic of a typical mixture Joule Thomson refrigerator. The compressor unit consists of a single stage oil lubricated compressor, an aftercooler and an oil separation unit. Gas lines connect the compressor to the cryostat. The cryostat consists of a counter flow heat exchanger, throttle valve and evaporator. 395
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PERFORMANCE The first test a run with a rebuilt compressor RW2 Leybold AG. Insulation vacuum was produced with low cost two stage oil lubricated rotary vane mechanical pump, and reached 10-4mbar < P~k < 103mbar. Figure 3 shows a typical cool-down curve, Fig. 4 shows the applied heat load as a function of the cold end temperature for the cooler with the mixture Gem_4. The measured exergetic efficiency of the cycle (assuming isothermal compression) is ~Te ~'~ ~ 43%. An valuable efficiency number can not yet be stated because the compressor used so far has a volumetric efficiency of less then 10 % and is therefore not suitable for series application. The system operates reliably and reproducible. SUMMARY A closed cycle Joule Thomson Refrigerator for cooling electronic devices, which operates in the temperature range from 70 K to 95 K, has been developed which uses refrigerant mixtures and a single stage oil lubricated compressor. It provides high reliability and no maintenance. This system is compact, has a good thermodynamic efficiency and low levels of vibration and noise.
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.
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78
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82
83
COLD END TEMPERATURE, K
Figure 4 Heat load vs. cold end temperature
REFERENCES 1. Alfeev V.N., Brodyanski V.M., Yagodin V.M., Nikolsky V.A., Ivantsov A.V., Refrigerant for a cryogenic throttling unit, UK Patent 1,336,892 (1973) 2. Abakumov L., Dengin V., Ermakov V., Malamyzhev V., Landa Y., Technical level and ways of improvement of Joule-Thomson minicoolers, Proceedings of the third international conference CRYOGENICS 94, Usti nad Labem, p. 36 3. Longsworth R.C., Boiarski M.J., Klusmier L.A., 80 K-closed cycle throttle refrigerator, Cryocoolers 8, Plenum Press, NY (1995) p.537 4. http://www.mmr.com/mmrklim.html at 16.04.1996 5. Longsworth R.C., Cryogenic refrigerator with single stage compressor, U.S.Patent 5,337,572 (1994)
Development of Magnetic Refrigerator Operating in Rc•
Temperature
Toshihiko Ochi, Hisaki Masatomi*, Yoshihiro Hasegawa*, Ryozo Aoki*, Tetsuro Ogushi**, and K~uyoshi Y abu-uchi*** Sumitomo Precision Products Co. Amagasaki Fuso l-10, 660, Japan *Dept. of Electrical Eng. Fac. of Engineering, Osaka University, Yamadaoka 2-1, Suita 565, Japan **Adv. Tech. R&D Center, and ***Manufacturing Eng. Center, Mitsubishi Electric Corp., Amagasaki Tsukaguchi 8-1 - 1, 661, Japan
Without an 5, compressed gas as freon, new type refrigerator is recently much demanded in common use. We have designed a conventional magnetic refiigerator with cascade connection system of Gd magnetics and directional current heat pipes, which are driven by rotating permanent magnets. The serially connected system showed essential characteristics of the heat pumping action. However, Gd has [x~or thermal diffusivity and its bkx:k showed inapplicably long relaxation time o f t = 6min for the heat transfer. Accordingly an improved design was attempted to incorporate it into the heat pipe which brought us a realistic short ~: of l min and the heat current ratio of more than 6 for the following to the counter direction. These characteristics provided us a possibility of real construction of the refrigerator.
INTRODUCTION Great extent of power energy use brings enormous heat exhaust in recent industry and society. Present status of the conventional refrigerator and air conditioner is based on the traditional technology of gas compression and expansion, and new technical developments have been expected[1 ]. For instance, the noble gas ; fluoro carbon is now making a serious problem in global ecology and alternatives are demanded ,and high density integrating circuit system requires an elegant refrigerator more efficient and vibration free in operation. Magnetic refrigeration has been developed since 1926, however, mostly in vel3' low temperature range below 1K, and some attempted up to 20K[2], which restricted the application only in research laboratories. Recently magnetic cooling from room temperature was realized by Brown [3] with use of fenomagnetic Gd down to-I~
However, he used a huge electrical magnet of 7T and gradual displacement of Gd
blocks in liquid heat resel~'oir, whence far from a practical design of the industrial use. Accordingly we have presented a design scheme of a compact refrigerator[4] economical and applicable to real social use e.g. automobile refrigerator or domestic use as follows. 399
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ICEC16/ICMC Proceedings
DESIGNS For refrigerant, rare earth element of Gd or its alloys are adopted. The pure Gd ingot was vacuum melted by Mitsubishi Kasei Co. and was in 3-nine purity with oxygen content less than 800 ppm.
It shows
magnetic phase transition at r(x)m temperature of Tc=292.SK with large entropy change due to 8 effective Bohr magnetons.
For operation, small permanent magnets (Br=-I.2T) namely NdFeCo (Sumitomo
Special Metal Co.) are used, and it provides considerable strong field of H=0.7 T in the tx~le gap[5] without any supply of electrical power.
The merit enables us to design economical and compact
equipments, however, this amount of field only gives temperature change of less than IK on Gd substance by one-shot operation [4,5]. Accordingly for obtaining considerable temperature-span cooling, successive magnetic operation is required for cascade heat transfer process on the serial arrangement ot" the Gd specimens as shown in Figure 1. By driving the magnet yoke over through this serial connection, a heat wave packet with temperature up and down is formed, however, for obtaining a heat pumping operation from a low temperature (TI,) to high temperature (TI0 reservoirs, thermal diodes having heat current rectifying action must be set among the Gd specimens as shown in Figure 1. For this purtx~se, a heat pipe (HP) with asymmetric structure of ethanol evatx)rator has been developed, and the heat conductance ratio ot K(follow)/K(counter)>40 with K(tk)llow)= 1. I x 103(w/m2K) was achieved[ 6,7]. EXPERIMENTS With use of these characteristic elements, a unit series of Cu(TI)-HP(L)-Gd-HP(H)-Cu(Tt0
was
constructed as depicted in Figure 2(a), and step motion of the magnet sweeping was carried out. Induced temperature changes in each part are presented in Figure 3 where we can recognize reasonable temperature changes of Gd by magnetization and demagnetization. The temperature difference
~T=TII-TL is gradually
growing with multiple operations. From this characteristic change it was confirmed that this unit part can play an essential role of the refrigerator element as expected[S]. The temperature variation is expressed by T(t)=T0exp [-t/~:], however the heat transfer through this system takes long time with "r=6min.
Accordingly even at the pericxt of
~t=Smin for the magnet stepping, one finds the Gd temperature change of only AT/T--40-50% remaining large l:x)rtion of the heat. In consequence the attainable temperature difference i~T was limited to 0.43K. Dominant cause of this small extent ~,T is considered followingly ; the first is the small thermal diffusivity D=4.3xl0-t~n2/s in Gd which is almost 1/27 of Cu, as common characters in rare earths. The other is the pc•
thermal contact between the working substance Gd and the diCxte HP, that is also generally to be
considered in the solid substance refrigerators in contrast with the case of fluid refrigerant. As the next investigation, N numbers of the unit element w,ere then connected in series between the hot(Ttl) and cold(Tl,) Cu block reservoirs as seen in Figure 1 and the attaining temperature difference tST>~TII-TI, was observed by the operations, and expressed in Figure 4 with N. Here we find a saturation tendency and
~TN=I.2K at N=6.
In order to analyze this character, the time variation of TI,(N=6) from the level at t=0 was computationally simulated with an ideal condition of no back flow through the thermal diode and no heat income from the atmosphere, and the result is shown in Figure 5.
Substantially different from the
observed T(obs), the computed T(ideal) shows large decreasing without any saturation. However, when the heat back flow is taken into account by a condition of remaining ethanol film of 80 lain thickness on the
ICEC16/ICMC Proceedings
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condenser plate surface, the computed T(back tlow) shows large suppression of the lowering. Still existing difference from T(obs) will be attributed to heat leak from the atmosphere in vacuum by radiation and also conduction via support contact. This heat income per each step operation is proportional to the period ~t, and then more frequent operation with shorter period is required for obtaining low temperature and large heat flux pumping, while the pericxt is limited by the thermal relaxation time "~. Accordingly, an improved construction of the (Gd+HP) unit was designed as depicted in Fig.2(b), where the Gd specimen was cut into sliced plates and incorporated in the heat pipe. This structure will be effective to diminish the heat resistance inside and outside of the Gd substance. Three units of this combined (Gd+HP) system are set in a serial connection as shown in Fig.2(b). RESULT The step magnetization pr(xzess was carried on this combined system and the temperature changes from t=0 are presented in Fig.6, where we can find as first remarkable improvement in the heat relaxation time shortening by the stepping period of ~t in sec. The next point to be noted is the different character in time variation of Tt! and TI~. After rapid cooling due to the demagnetization of the adjacent unit (see Fig.2(b)), TIt starts to increase by the forward heat tlow from the central unit magnetization, while TI~ continues to decrease with little counter flow. This indicates a real evidence of the heat rectifying operation by asymmetrical heat pipe structure. The Tt~ and TI. changes were well expressed by exp[-t/~] and the thermal relaxation time -r through the 3 units was determined at T(lbllow)= lmin. with -r(counter)/l:(follow)>_ 6. This shorter ~ value shows remarkable improvement in the heat conductance of about 6 times to the Gd+HP separated system of Figure 2(a) and we can expect significant progress in a new type magnetic refrigerator designed by use of this incorporated unit of (Gd+HP) type. REFERENCES 1 Barclay, J.A., Magnetic refrigeration Adv. Cry. Eng. (1988) 33 719-731 Barclay, J.A, A review of magnetic heat pump technology Proc. 25th Intesoc. Energy. Convers. Eng. Conf. , USA (1990) 7 222-225 2 Nakagome, H. and Tanji, T. and Horigami, H. and Numazawa, T. and Watanabe, Y. and Hashimoto,T., Adv. Cry. Eng. (1984) 29 581 3 Brown, G.V., Magnetic heat pumping near rcx~m temperature J. Appl. Phys. (1976) 47 3673-3680 4 Aoki, R., Magnetic refrigerator operating from room temperature with permanent magnet Cryogenic En~.(JaDan) (1985) 20 294-301 50chi,
T. and Aoki, R., Proc. 10th Intl.Workshop on Rare Earth Magnets and theirAppl. Kyoto, Japan
(1989) 195-201 6 0 c h i , T. and Ogushi,T. and Aoki, R., Proc. Intl. Workshop on Thermal Invest. of ICs and Microstr. Grenoble, France (1995) 150-153 7 0 c h i , T. and Ogushi, T. and Aoki, R., Development of a heat pipe thermal diode and its heat transport performance JSME Intl. J. (1996) 39B No2(in print) 8 0 c h i , T. and Aoki, R., Magnetic c(x~ling and refrigerating effect by rare earth element Rare Earths (Japan) ( 1991) No 19 19-30
402
ICEC16/ICMC Proceedings _T L
Gd I
l
HP ! eid 2
~
ltP 2
(.id N
,--
1.5[~,--,----,
r
1.0-
Magnetl
i
,
j / / ~
M t ~
..q,
k_~
/
e~ 0.5
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!
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Figure packet blocks magnet
! I
o
0
1 Schematic diagram o1 heat wave transfer through the series of Gd and the heat-pipe diodc(HP) by sweeping
........
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I .........
1
3
2
N
I. . . . . . . .
I ......
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4
5
6
Figure 4 Obtainable temperature difference 6T=TH-TL with connection number N of the units
Malliel
C~(TL) lip(L)
Gd
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":i::~?:~::~i:i:i:.;::~i:i!i'. ;.'i!: .
.
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(a)A unit composed of a working (Gd) bl(x:k and two h e a t pipes (HP) with hcat reservoirs(Cu)
llliIlll]
.I
-////J/
r.lilllill.Ii
[I~
~'~/--/-/~
_I~
-i.
12)
0
5.
X._
_ T(ide,I)
-' ~
, ,~.lep n u m b e r
I1
Figure 5 Comparison of the temperature TL, changes by 6 unit, among the observed T(obs) and the computed with and without the hcat back flow ; T(back flow) and T(idcal)
(I)
(b)Connection of 3 ilnils of the incorpx~raicd Gd slices in heat pipe
Figure 2 Constnlction of the heat pumping unit (l)Ethanol cvapol,aling Cu face with wicks (2)Ethanol condenser of rx)lishcd face (3)Gd blocks and (3')Gd sliced platcs
0.5
-: ~
i-+~-~.<__
- - , ~ ~ ~
0 .....1. . . . . . . . . . . . . . . . I ................... l-
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i/
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r_L.g.~/_///_~
\
(3')
"l'(.h~)
.
rll
r
.
<1
, ~i
........... i
?'~'d~
. .i !
-0.5
0
0
--0.5 ...........
0
I
30
60
90
t (rain)
Figurc 3 Temperature change of each part in the unit of Fig. 2(a) by Ihe magnet step operation
s'0
~o. t(sec)
240
'
320
Figure 6 Temperature changes of TII and TI~ on the unit series of Fig.2(b) by magnet step operation, showing rapid thermal relaxation(x= l min) and asymmetric character between TIt and TL
The Characteristics of Magnetic Refrigeration Operating at the Temperature of 20 K
Katsuhide Ohira, Shinobu Matsuo, Hitoshi Furumoto Nagasaki R&D Center, Mitsubishi Heavy Industries, Ltd. 1-1 Akunoura-machi Nagasaki 850-91 ,Japan
An experimental study of a magnetic refrigeration operating in a Camot cycle at the temperature of 20 K has been conducted. GGG (Gd3GasO12) was selected as the magnetic refrigerant. The magnetic refrigerator consists primarily of a superconducting magnet, a Gifford-MacMahon type refrigerator and a mechanical heat switch. Firstly, it has been confirmed that the calculation of the magnetic entropy of GGG by assuming molecular field approximation was in good agreement with the experimental data for higher temperatures than 18 K. Secondary, the refrigeration power of 0.21 W was obtained from magnetic refrigeration tests at the temperature of 20 K and the cycle frequency of 0.002 Hz, and this experimental results agreed with the calculated results of cycle simulation.
INTRODUCTION Recent years have been expanded demand for liquid hydrogen, not only in the field of space development, but also as a clean energy source in industrial applications, and lower cost methods of liquefaction are being sought. Since magnetic refrigeration operates in a Camot cycle in principle, potentially its efficiency is expected higher than that of gas refrigeration. However, because most magnetic refrigerants exhibit large changes in magnetic entropy below the boiling temperature of helium (4.2 K), magnetic refrigeration has been used only in the liquefaction of helium or the production of superfluid helium. The objective of the research reported here was to develop a high performance magnetic refrigerator to be operated at the boiling temperature of hydrogen (20 K). Single-crystal GGG (Gd3GasO12) was used as the magnetic refrigerant, and the characteristics of magnetic refrigeration at the temperature of 20 K were obtained.
EXPERIMENTAL APPARATUS Overall Structure and Magnetic Refrigerant Figure 1 illustrates the oveall view of the experimental apparatus. It consists of a superconducting magnet, a mechanical heat switch, a magnetic refrigerant and a Gifford-MacMahon type refrigerator (G-M refrigerator) for heat rejection. Figure 2 is a photograph of the test apparatus. A cylindrical shape GGG (diameter of 0.04 m, height of 0.1 m, 0.83 mol) shown in Figure 3 was used as a magnetic refrigerant. Details of Heat Switch Details of the heat switch are shown in Figure 4. The heat switch made of high thermal conductivity copper is placed in proximity to the upper portion of GGG. The heat switch is in contact with the copper block which is thermally connected to the G-M refrigerator, moves up and down, contacts GGG, and rejects heat from GGG to the G-M refrigerator. The cold head of the G-M refrigerator can reach a temperature of 10 K, but, because it is being used as a heat sink for the sake of covenience, an electric heater is wound around the cold head to adjust the temperature to above 20 K. 403
404
ICEC 16/ICMC Proceedings
Measurement of Refrigeration Power Instead of liquefaction of hydrogen, in this experiment an electric heater is wound around the magnetic refrigerant in order to measure refrigeration power. Superconducting Magnet The solenoidal superconducting magnet had an external diameter of 0.136 m, a bore diameter of 0.08 m, a height of 0.195 m, and a maximum magnetic field of 8 T (energizing current of 84 A). Since it was not a pulsed magnet, magnetization and demagnetization were operated at the sweep rate of 0.04 T/sec.
RESULTS AND DISCUSSION Characteristics of Magnetic Entropy of GGG Although the magnetic entropy of GGG can be calculated by assuming molecular field approximation, experimental verification has thus far been conducted only below the temperature of 4.2 K [ 1]. Therefore the authors measured the magnetic entropy of GGG, having ascertained the magnetic refrigeration characteristics near the temperature of 20 K. The magnetic entropy of GGG was measured by following process; The specific heat (Cp) of GGG is dependent on temperature and magnetic field, therefore GGG was heated by using an electric heater at fixed magnetic fields, then the specific heat was first determined from the temperature rise (dT) of GGG and the power (dQ) of the electric heater. The entropy change (dS) was calculated by Cp" dT/T. Figure 5 shows the measured and calculated entropy of GGG. Measured values are in good agreement with calculated values, with error in entropy change of less than 2 % at 4 T and 20 K. This confirms that calculated values by assuming molecular field approximation can be used in practice. Magnetic Refrigeration Tests Time-temperature profiles obtained during the magnetic refrigeration test are shown in Figure 6. When the temperature of GGG rose above the heat rejection temperature (25 K) during the magnetization process, the heat switch was placed in contact for heat rejection. When it fell below the temperature of 20 K during demagnetization, GGG was heated by the electric heater. Table 1 shows the magnetic field patterns and the timing of the heat switch during the magnetic refrigeration cycle. Table 1
Magnetic refrigeration cycle parameters Magnetic Field tl t2 t3
214 s
258 s
464 s
Heat Switch tsl ts2 134 s
276 s
Heater tal ta2 315 s
464 s
The refrigeration power measured at the heater power of 0.65 W and the cycle frequency of 0.002 Hz was 0.21 W. Solid circular symbols in Figure 7 indicate the experimental magnetic refrigeration cycle on temperature-entropy diagram of GGG. Lower conductance of the heat switch and insufficient thermal insulation of GGG prevented the achievement of ideal adiabatic and isothermal processes. Figure 8 shows the temperature variations of the heat switch and GGG during the test. The captions (a) through (e) designate the temperatures of the switch and GGG. (Refer to Figure 4.) When the heat switch is in the ON position, there is a very large temperature difference between the tip of the switch (b) and GGG (a). This can be attributed to poor contact between the switch and GGG due to insufficient precision in the manufacture and assembly of the switch. The heat transfer rate of the heat switch (in temls of rejected heat per temperature difference) was 0.07 W/K, considerably lower than the design value of 1 W/K. Comparison Between Experimental Results and Cycle Calculations Using the heat transfer rate of the heat switch, the power of the electric heater, and other factors measured during the experiment as input data, the magnetic refrigeration cycle was calculated [2]. The results are plotted with open circular symbols in Figure 7, yielding a refrigeration power of 0.22 W. The difference
ICEC16/ICMC Proceedings
405
between the calculated value and the experimental value is within 5 %, and it was confirmed that the calculated cycle drawn on temperature-entropy diagram of GGG was in good agreement with the experimental result.
CONCLUSIONS (1) It was confirmed that the molecular field approximation-based calculations of magnetic entropy of GGG are valid above the temperature of 18 K. (2) The refrigeration power of 0.21 W was obtained from the tests of magnetic refrigeration operating at the temperature of 20 K using GGG as the magnetic refrigerant. (3) Poor contact between the heat switch and GGG caused a lower heat transfer rate for the heat switch than the design value. (4) The calculated refrigeration power for the magnetic refrigeration cycle operating at the temperature of 20 K was in good agreement with the experimental result.
REFERENCES (1) Barclay, J. A., and Steyert, W. A., Materials for magnetic refrigeration between 2 K and 20 K Cryogenics (1982) 22 73-80. (2) Numazawa, T., Kimura, H., Sato, M. and Maeda, H., Carnot magnetic refrigerator operating between 1.4 and 10 K 6th International Crvocooler Conference Vol. lI (1990) 199-213.
G-M Cycle Refrigerator , I .
.
.
.
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I
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,
i
Fig.1
iiiiiiiiil
~
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Cross sectional view of magnetic refrigeration test apparatus
Fig.2
L
m
Photograph of magnetic refrigeration test apparatus
406
ICEC16/ICMCProceedings G-MCycle It R~frigerator II1
1e,,ows-
t
-(cl Thermometer
I[
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C[
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(~) Fig.3
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Photograph of magnetic refrigerant
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Details of heat switch
R'Gas Constant 8. 314 (J/tool 9K) .
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/
Study on Theoretical Refrigeration Temperature of Regenerative Cryocoolers" Guangming Chen, Guobang Chen, Jianping Yu Cryogenics Laboratory, Zhejiang University, Hangzhou 310027, P.R. China
Based on the principles of thermodynamocs, the relation between isentropic expansion coefficient kq and isobaric specific heat Cp is found out. The values of ~, at different temperature and pressures are calculated. From theoretical and experimental values of (Tp of helium in the supercritical area, the limits of refrigeration temperatures which may be reached by traditional regenerative cryocoolers is theorectically demonstrated. It is pointed out that this boundary is just the lambda line of helium. In order to obtain temperatures lower than lambda line, a new type of regenerative cryocooler which can work in helium lI region is proposed. The working principle, schematic diagram and theoratical refrigeration temperature of this new cryocooler are discussed.
ADIABATIC EXPANSION OF FLUIDS The refrigeration effect of a cryocooler can be represented by the value of adiabatic expansion of the working fluid in the expansion space. The temperature change due to a pressure change at constant entropy is expressed by means of the isentropic expansion coefficient p~ 8T /4- (-~)s (I) If the emropy of a refrigeration system is regarded as a function of T and P, then the differential entropy can be expressed as follows:
d s - C,. d r
(2)
dP
T
Applying the condition for an exact differential, two thermodynamic identities here are: 1 (8C r
7
or
82v
)r--(cT2
(By
)e
(3)
1 8C~
-~)P - -I-T ( 8P )rdT+ a(P)
(4)
where c~(P) is a function of P. Applying the condition of ideal gas to Eq.(4), we get or(P) - R. Eq.(2) P changes into: ds-
dr-
dr +
then
The proiect is suported by National Natural Science Foundation of China.
407
dV
(5)
408
ICEC16/ICMC Proceedings
So far, we can calculate the isentropic expansion coefficient #, from Eq.(6) by known values of specific heat of a refrigerating fluid. For an ideal gas, Cp, is independent of P , thus RT v A/,~, =
=
~
(7)
P(Tp ( p Obviously, the value of kt~. is positive, i.e. an isentropic temperature decrease results from a pressure decrease. a RT a ~, therefore For a Van der Waals fluid, (P + -V-T ) ( v - b ) - RT or P = v-b v3v _ R(v-b)
(8)
(-~)P
RT-2.a- (1 - b)~ I.'
1~
As a result 1" c~,
_
I'~.
v ( l - b)
C,-~(o----~)PC,e[l__2a__(lh)2] _ =
v
RTv
(9)
v
which is also positive because v > b. Fig. 1 shows the specific heat ('p of helium as a function of temperature along selected isobar from 1.4K to 4.0K. We can see that the specific heat of helium varies with temperature in an unusual manner for fluids. At the lambda point, the specific heat increases to a large value (The maxima in the plotted curves correspond to I T - l ~ l - 1 0 - ~ K ) ; but it differs on either side of the lambda line. The relation of Cp and P is similar to the behavior of general fluids in temperature region of T > ~ , i.e., its dCp ( 0 P ) r < 0, #~ > 0. However, the relation of ( e and P is quite different from that of other fluids at the C('p region of helium I I ( T < / ] ) , i.e. its ( ~ P )v > 0, #s < 0. It means that there exists a boundary line in HeII across which #s changes its sign. This is just the lambda line. If T < I~, a temperature increase will result from an isentropic expansion of HeII. Based on the above analysis, we can draw a conclusion that the minimum refrigeration temperature of a regenerative cryocooler using isentropic expansion method of helium is limited by the lambda line. Fig.2 shows the minimum refrigeration temperature as a function of expansion pressure.
ISENTROPIC EXPANSION OF HELIUM NEAR LAMBDA LINE V. Arp et al [4] have constructed the equation of state for helium from 1.4 to 4K for pressure up to 2.5MPa that includes the lambda line. They have shown that (]o near the lambda line is given by
- a n o g IT - 7' 1+
(10)
where E is a fianction that remains finite at the lambda line. The value for a is given by Arp as a - - 0 . 6 1 ( + 0 . 1 % ) J~ g . K :
T > I~
(lOa)
a - -0.57(+0.2% ) J / g. K:
T < 7)
(10b)
for helium I and
for helium II, which is independent of pressure within the error limits shown. The exception is that Eq.(10a) does become a function of pressure above 1.5MPa. Rewriting Eq.(10) to separate pressure-
ICEC16/ICMC Proceedings
409
dependent and pressure-independent terms in E, one has (_~ - aT'log[T- T~[+ TEb(P,T)+ TEd(T )
(11)
Then, from Eq.(6) and (11), we have a
OT
_ T Or,)
:
log[r
I dT C Zb _q_Xe(e)
. . . . . . . . . . . . . . . . . . . . . . . . . . .
8P
OP
a l o g l r - T)w[+ ~b (P, T) + ~d (1)
(12)
The functions Z,d(I') and Z,e (P) are constants of integration. Fig.3 shows the isentropic expansion coefficient of helium near the lambda line by Eq.(12). The conclusion from the curves is that there exists strong and opposite temperature dependence of P,. for helium 1 and helium II, respectivelly. HE II REGENERATIVE CRYOCOOLER The minimum refrigeration temperature of a regenerative cryocooler using traditional isentropic expansion method is limited by the lambda line according to the above analysis. However, we can still reach the temperature below lambda line by means of isentropic compression instead of expansion. Fig.4 shows the schematic diagram of this cryocooler and Fig. 5 is the refrigeration cycle expressed on T-S diagram. Here, process 1-2 is an isentropic compression process when the displacer moves down; 2-3 is an isobaric heat-absorbing process when the displacer keep immovable; 3-4 is an isentropic expansion process when the displacer moves up; 4-1 is an isobaric heat rejecting emitting process, the heat is transferred into the precooling stage through the heat exchanger. The area 2-3-b-a-2 represents the net refrigerating effect of a cycle. The temperature drop of isentropic compression process ATe_2 relates to the initial state points and pressure ratio. Fig.6 shows the relation between ATe_2 and pressure ratio at different initial states. We can know from Fig.6 that at the same initial state, the bigger the pressure difference is, the more the temperature decreases; at the same initial pressure and pressure difference, there exists a maximum between the temperature drop and the initial temperature. So we can select the cyclic parameters of the cryocooler. CONCLUSION Based on analyzing thermodynamic properties of helium and characteristic of regenerative cryocoolers, we have got the property of isentropic expansion coefficient near the lambda line. The minimum theoretical refrigeration temperature using expansion method is limited by lambda line. The paper advances a new idea of a regenerative cryocooler in Hell region using isentropic compression method, which lays a theoretical foundation for the study of the regenerative cryocooler in Hell region. REFERENCES Guobang Chen et al., Investigation on a two-stage Solvay refrigerator with magnetic material regenerator, C~_ogenics 32 ICEC Supplement (1994) 5-8 V. Arp, State equation of liquid helium 4 from 0.8 to 2.5K, Journal of Low Temperature Physics 79 (1990) 93-114 V. Arp and K. Agatsuma, The equation of state for liquid helium from 1.4 to 4K and asymptotic limits of the lambda line, Journal of Low Temperature Physics 73 (1984)
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Cryogenic engineering
Cryostats
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A Demountable Type Test Cryostat for Heat Transfer and Thermal Optimisation Studies of Liquid Helium Containers
S. Jacob, S. Kasthurirengan, R. Karunanithi and G. V. Satisha
Central Cryogenic Facility, Indian Institute of Science, Bangalore - 560 012, India
The design of liquid helium containers uses both the highly efficient multilayer insulation (MLI) along with vapour cooled shields attached to the neck tube. We describe a demountable liquid hefium cryostat of 100 litre capacity to carryout optimisafion studies on the number of vapour cooled shields, their positions on the neck tube and the number of MLI layers interposed between them. A Kapton film seal between the neck tube and the top dished end provides demotmtability between the inner vessel and flanged vacuum jacket. The cryostat is insmanented for measuring temperature profiles of neck tube and the insulation layers. The paper discusses the heat transfer analysis for a typical case.
INTRODUCTION Multilayer insulation (MLI) is a highly effective cryogenic insulation, used in cryogenic systems. The design of MLI insulated cryogenic containers for liquid helium, uses the refrigeration of the outgoing vapour to cool the radiation shields, since the ratio of enthalpy of helium vapour to its latent heat is quite high (~ 75). Thus the concept of vapour cooled multishielding has been suggested by Paivanas et al [ 1] for liquid helium containers. Since then, the design of liquid helium dewars and cryostats have been approached both theoretically and experimentally [2-5]. Xu Lie et al [6] has described the standard design method of MLI insulated wide mouth liquid helium dewars along with the detailed heat transfer analysis. However, in view of the interdependent nature of the heat transfer on the position of the vapour cooled shields and the multilayer insulation which are attached to them, optimisation studies still need to be done to determine the number of such vapour cooled shields, their positions on the neck tube and also the number of MLI layers to be wrapped on each of these shields, in fabricating a liquid helium container. However, so far no experimental system has been reported wherein such an optimisafion study can be made, by altering the radiation shield positions and also the number of MLI layers on them. Ours is an attempt in this direction. In this paper, we describe a demountable liquid helium experimental cryostat of 100 litre capacity, using which the above optimisafion studies can be made by varying several parameters. Typical heat transfer analysis is presented here.
DESCRIPTION OF THE EXPERIMENTAL CRYOsTAT The schematic of the experimental cryostat is shown in Figure. 1. The inner vessel of 100 litre capacity is made of AISI 304L stainless steel end has an o.d. of 500 mm and wall thickness 2 mm. The top dished end is connected to a thin walled (0.3 mm) stainless steel neck tube of AISI 321 grade and ends in a bellow. 413
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The bellow top is welded to the bottom flange of the Kapton seal joint. The outer vacuum jacket is made of ,MSI304 stainless steel and has a diameter of 700 mm and wall thickness 3 mm. It is made of 3 pal~, namely the bottom dished end, the cylindrical shell and the top dished end. They are connected together by 'o'-ring sealed flange assemblies. The top flange of the Kapton seal is welded to the central opening on the top dished end. By this arrangement, when this dished end is opened out, the inner vessel can be totally separated from the outer vessel and the necessary modifications of radiation hats and insulation structure can be carried out. S AES Getter pellets are mounted at the bottom dished end of the outer vessel, in such a position so that it can be reactivated b,' external heating to 400~ The first radiation shield around the inner vessel is a copper cylindrical vessel and is tLxed to the neck tube by soldering. On the neck tube, split copper rings have been inserted, which can be anchored at required heights on the neck tube. Copper radiation hats are attached to the rings by brass screws and good thermal contact ensured with Indium wire squeezed in between. Aluminium shields of 1 5 ~ thickness are f'Lxed to these copper shields by aluminium adhesive tape. The support structure consists of a central bracket into wlfich the inner vessel is mounted via an FRP tube. The connection between the bracket and outer cylinder is done by 3mm diameter stainless steel rods connected at 120 ~
Figure 1. Schematic of the experimental clyostat 1.Demountable KAPTON seal 2. Bellow 3. ORing 4. Inner vessel 5. Neck tube 6. Activated charcoal 7. 1 st copper radiation shield 8. Vacuum jacket 9. SS supporting rod 10. FRP tube 11. Pump-out port 12. Getter cartridge E x p e ~ e n ~ l Procedure .~ffter insulating the inner vessel with a specific experimental conliguration it is mounted in the outer vessel and the interspace is evacuated to a pressure of 10 .4 mbar (the insulation structure is heated to 70~ during evacuation) and the getter is activated. 'After the insulation structure is cooled down to room temperature, the experimental dewar is disconnected from the vacuum system. The inner vessel is first precooled with LN2 and then fdled with liquid helium. The evaporation rate of the liquid helium dewar is monitored by measuring the helium gas flow through an integrated gas flow meter as a function of time. The temperature of the inner vessel, the neck tube, the vapour cooled radiation shields and insulation layers are monitored by using thermocouples (copper-chromium Iron and copper constantan types), PT100 and Sificon diodes, mounted at appropriate positions with the help of Delta Bond adhesh,'e. .MI the leads are brought out through two vacuum feed throughs (60 pin and 37 pin) on the outer cylindrical shell. 'Ihese are connected to a DAQ system (HP 3852A with a built in reference junction), and can be scanned at regular intervals using a HP9000/300 computer. Silicon diode sensors and PT 100 are previous .ly calibrated between 4 to 300 K and this calibration data are used in actual measurements with the experimental system.
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HEAT TRANSFER ANALYSIS Before taking up experimental work, a theoretical estimation of the temperature profile of the neck tube and optimum locations of the vapour cooled shields have to be obtained. Temperature distribution along the neck tube For perfect heat transfer between the evaporating vapour and the wall of the neck tube, we can assume that the vapour temperature equals to the temperature of the neck wall at the corresponding section. Using the mathematical model suggested by Xu Lie et al [6], it can be shown that: d.r~
(I)
pT~-O /
with the boundary conditions, at x=0, T=T0 and x=L, T=T, wherein To is liquid helium temperature, T~ is the ambient temperature, L is the length of the neck tube, ~,=~,(T) and A are the thermal conductivity function and cross sectional area of neck tube respectively, m is the evaporation rate of the dewar and Cp=Cp(T), the specific heat of helium vapour. Solving equation (1) by finite difference method, we obtain the temperature distribution of the neck tube. Figure 2 shows the typical temperature distributions for different evaporation rates of liquid helium for the present experimental configuration.
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(2)
where Lo is the latent heat of vapourisafion, Q~ is the radiation heat transfer from the neck tube top plane of the dewar, Qw is the heat conducted by the neck tube, QF is the radiated heat from the neck tube wall to liquid helimn, Qo the heat conducted by the vapour. Qrl is the radiated heat from the first radiation shield, QP1 is the residual gas conduction between the first shield and the inner vessel and Qs is the heat conducted by the supports. The expressions for these terms are given in reference [6].
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Q . and Qpl depend on the first shield position and its temperature while the other terms are independent of the above. (3)
9"- Q , + Qpl = m L o - (Qn + Qw +Qo + Qx~ +Qs) = c
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~o and e~ are emissivities of inner vessel and the first shield, c~ is the Stefan Boltzmann constant, .% and Ax are the areas of inner vessel and the first shield, Trl is the average temperature of the first shield, o~ is the accommodation coefficient,7 = Cp/Cv, R is the univeral gas constant, M is the molecular weight of air, T is the temperature at the measuring point of pressure P.
By iteration procedure, the average first shield temperature Ta can be calculated. For a typical case suggested by Xu Lie et al [6],where the first copper shield average temperature is 60 above that of the joint location at the neck tube, we can now calculate the position at which first shield is attached to the neck tube, from the temperature profile given by equation (1).The position of subsequent shields can be estimated from the heat balance betweeen vapour cooled shields (i = 1, .......... n), and can be written as
(Qr + Qp + Qsr +l : (Qr +Qp + Q,r + Qli
(6)
where Q~ is the solid conduction heat transfer in MLI, and (Q,, h = 0, since there are no MLI layers between inner vessel and first vapour cooled shield. Based on our previous experimental studies [7], Q~is assumed to be ~- 5 times of Q r i .The lateral heat transfer on the shield Q~i is given by [6] Q~i= nKD~L,~ " /(~ +1
(7)
Where K = 0.75 and Kn = 3.81, D~ is the diameter of the i~ shield and I_4 is the effective length of the shield. Using the value of Q~iin equation [6]. T,i. Can be evaluated. Assmning that for aluminium shields, their average temperature is 160 above that of the joint location at the neck tube [6], the positions of these shields are estimated from the temperature profile. Table 1 gives the results of the above calculations on the optimum number of vapour cooled shields and their positions with and without MLI for an assumed evoporafion rate of l litre/day. It can be seen that the number of vapour cooled radiation shields is reduced from 7 to 5 with the introduction of 105 MLI layers. With this configuration, the radiation shields arc mounted on the neck tube of the inner vessel and it is insulated with MLI. Experimental studies using this analytical approach is in progress.
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Table I Theore~ally calculated number o f vapour cooled shields and their positions Shield material Number
Without MLI Position, x (m)
Copper (Electrol3~ic grade thickness O.8mm)
1
0.261
Aluminium foil
2 3 4 5 6 7
0.330 0.362 0.382 0.398 0.411 0.420
(15~m)
Number 1
2 3 4 5
With MLI Position, x (m) 0.261
0.346 0.382 0.407 0.420
CONCLUSION In this work, we have described a demountable experimental cryostat for liquid helium, which enables to carryout optimisafion studies of vapour cooled radiation shield positions and number and the number of MLI layers. To our knowledge, no such experimental system has been reported wherein, the experimental validation of the theoretical analysis of heat transfer to helium dewar can be done, on the same cryostat by varying the number and positions of the vapour cooled radiation shields on the neck and the number of MLI anchored to these radiation shields. ACKNOWLEDGEMENTS The authors wish to acknowledge the help rendered by the staff of Central Cryogenic Facility tor this work. They are thankful for Department of Science & Technology for funding this work.. REFERENCES 1. Paivanas, J.A., Roberts, O.P., and Wang, D.I-J., Multishielding- An ach,anced superinsulation technique Adv. Cryog Eng., (1965) 10, 197-207. 2. Mende, F.F., Gorbunor, V.M., Bondarenko, N. N., Logvinor, V.N., and Zhuravel, T. I., Broad-neck liquid helium cryostat with a long lijbtime Cryogenics (1989) 29 998-1001. 3. Qingfia Li, Eyssa, Y.M., and Meintosh, G. E., Discrete cooling of supports and ~,fulti-layer insulation in Helium dewars Adv Crvoz Eng, (1984) 29, 785-793. 4. Bejan, A., Discrete cooling of low heat leak supports to 4.2 K Cryogenics (1975) 15 290-292. 5. Bora, M. S., Rugaiganisa, Shigetomo Nakagawa, Masao Yoshiwa, Takeo Yoshihara and Akira Hira Experimental im,estigation on heat leak into a liquid helium dewar Cryogenics (1990) 30 942-946. 6. Xu Lie and Wang Ruzhu, Study of the Standard design method of multishields insulatated metallic liquid helium dewars International Conference on Ener_~_ Saving in Refrigeration, Xian, China, (1986), 205-210. 7. Jacob.S, Kasthurirengan.S and Karunanithi.R .Im,estigations into the thermal performance of multilayer insulation.(300-77K) Part 2. Thermal analysis Cryogenics_ (1992)32,12,1147-1153..
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Cooling Performance of a Pressurized HeII Cryogenics System for the Superconducting Magnet Test Facility at KEK Nobuhiro Kimura, Yasuo Ajima, Yoshikuni Doi, Tomiyoshi Haruyama, Norio Higashi, Masahide Iida, Shuichi Kato *, Masanori Kawai, Hiroshi Kawamata, Seog-whan Kim, Yoshinari Kondou, Yasuhiro Makida, Katsuhiro Mimori, Shoichi Mizumaki **, Tatsushi Nakamoto, Toru Ogitsu, Hirokatsu Ohhata, Norihito Ohuchi, Shigekatsu Sugawara, Takakazu Shintomi, Ken-ichi Tanaka, Akio Terashima, Kiyosumi Tsuchiya, Hiroshi Yamaoka, and Akira Yamamoto
National Laboratory for High Energy Physics (KEK), 1-10ho, Tsukuba, Ibaraki 305, Japan * Kyushu University, Fukuoka, 812, Japan ** The Graduate University for Advanced Studies, 1-10ho, Tsukuba, Ibaraki 305, Japan This paper describes a new cryogenic test facility developed at KEK for research and development on high field superconducting magnets. It consists of a vertical double-bath cryostat for l m model magnets in superfluid helium at atmospheric pressure associated with a vacuum pumping system and a refrigerator/liquefier. The system has been successfully operated since 1995 with a refrigeration power of 8 W at 1.9 K, and two superconducting magnets have been tested.
INTRODUCTION A cooperative effort to develop high field superconducting magnets has been carried out between CERN and KEK, as part of the joint research program towards the Large Hadron Collider to be built at CERN. A series of high field superconducting model magnets are being developed at KEK[ 1,2,3]. To test those magnets, a new 1.8 K test facility has been developed at KEK, and a cooling performance test has been carried out by using a dipole model magnet. This paper describes the results of the performance test, of the 1.8 K test facility. TEST FACILITY A flow diagram of the test facilities cryogenics system is shown in Figure 1. The cryogenics system consists of a refrigerator/liquefier, a valve connection box, a vertical double bath cryostat with a heat exchanger and a pumping system (1.8 K refrigerator) which provides saturated superfluid helium inside the heat exchanger. The refrigerator/liquefier was built by Teisan/L'Air Liquide and was previously used for the superconducting solenoid in the AMY detector system at TRISTAN[4]. Its cooling power is 300 W at 4.4 K or its liquid helium production capacity is 100 L/h. During cool down of the magnet, it works as a refrigerator. In steady state, it works as a liquefier. Major parameters of the cryogenics system are listed in Table 1. Table 1. Parameters of the cryogenics system Cold box Liquid He Dewar Liquid Nitrogen Tank 1.8 K Heat Exchanger area Vacuum Pumping He gas recovery
100 L/h (300W at 4.4 K) with LN2 2500 Liters + 2000 Liters 9700 Liters 0.75 m ~ 900 m~/h at 13 kPa (Roots pump) + 360 m-~/h (Rotary pump) 100 m"/h (Recovery compressor) + 20 m ' (Gas bag) 419
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Impure Gas St 2200Nm3 Recovery Compressor
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The power supply for the magnet test has a capacity of 15 kA with 15 V and a stability of l0 -4. The monitoring system provides 96 channels with sampling frequencies of 0.1 to 1 Hz for monitoring cool down or warm up of the magnet. PERFORMANCE TEST AND RESULTS The cryogenics performance of the 1.8 K test facility has been carried out during a test period of a dipole magnet. For precise measurement of the cooling characteristics and the refrigeration power five carbon glass resistance thermometers were installed in the cryostat as indicated in Fig. 2. The five thermometers are positioned in the cryostat at the: inlet of the 1.8 K refrigerator(Ts), after the precooler(Te~o,), at the outlet of the J-T valve(Texb), at the surface of the heat exchanger(T]), and in the 1.8 K pressurized helium bath(T2). The cooling process follows two cooling schemes. In the temperature range 300 K - 80 K, the magnet was cooled down using cold helium gas at about 50 K lower than the magnet temperature to eliminate excessive thermal stresses. Helium mass flow rate and gas temperature were automatically controlled to keep the temperature difference within 30 K between the He gas-inlet and the magnet temperature. Below 80 K where the thermal shrinkage becomes negligible, it was cooled quickly to liquid helium temperature by using the liquid helium which was fed to the cryostat from a 2500 L dewar via the valve connection box. After magnet training at 4.2 K, the magnet was cooled down to 1.8 K by flowing helium through the 1.8 K heat exchanger at low pressure. Figure 3 shows the typical cool-down curve of the magnet from 300 K to 4.2 K, and from 4.2 K to 1.8 K. By using this cryogenics system, the cool-down time of the model magnet from 300 K to 4.2 K took 2--3 days, and the cool-down time from 4.2 K to 1.8 K took 6 hours.
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(1)
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where X is the latent heat of the liquid helium, 9 is the density of the helium gas, and r~ is the mass flow rate of the pumping system. In this figure, the lowest obtained temperature of the cryostat was 1.73 K with no heater load, and the refrigeration power of 8 W was obtained at 1.9 K. A difference of 4W between the calculated value and measured value were estimated by the heat inleak from 4.2K bath to 1.8 K bath through the insulating plate and its electrical feedthroughs. Further difference might be due to superfluid leak around the feedthroughs and the flange of the insulating plate. With this capacity, five training cycle of the magnet have been performed within 10 hours, with a temperature profile as shown in Fig. 5.
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SUMMARY A cryogenics system for a 1.8 K test facility, for testing high field superconducting magnets at KEK was constructed and tested. A magnet was attached to the vertical double bath cryostat and was cooled down by the refrigeration/liquefier with the automatic computer control system. The cryostat has 8 W of refrigeration power at 1.9 K and it took 6 hours to cool down from 4.2 K to 1.9 K with 450 L of liquid helium and 2 tons of the model magnet. The system could be recovered within 2 hours after a quench at 1.9 K. The fast recovery to 1.9 K after a magnet quench permits several quenches per day and a maximum of five have been demonstrated. ACKNOWLEDGMENTS The authors would like to thank Prof. H. Sugawara, Director General of KEK, Prof. Y. Kimura, Deputy Director of KEK, and Prof. H. Hirabayashi and the former Head of Cryogenics Center for their continuous encouragement during this work. The authors also deeply thank Mechanical Engineering Center and Cryogenics group in Physics Division at KEK. REFERENCE Yamamoto, A et al Development of a twin aperture dipole model magnet for the Large Hadron Collider IEEE Trans. Appl. Superc. Vol. 5 (1995) No. 2 1016-1019 Yamamoto, A et al Test Results of a Single Apertt/re 10 Tesla Dipole Model Magnet for the Large Hadron Collider to be published in the Proc. 14th M.T., Tanpere, Finland, 1995 Shintomi, T et al Development of a 56 mm Aperture Superconducting Dipole Magnet for LHC paper presented at Appl. Superc. Conf., Pittsburgh, PA, U.S.A., Aug. 25-30 1996 Doi, Y et al Cryogenics system of a 3 Tesla Superconducting Solenoid for the AMY Particle Detector at TRISTAN Adv. Cryo. Eng. Vol. 33 (1988) 551-557 Mito, T et al Control System for Helium Refrigerators of TRISTAN Detector Magnets Adv. Cryo. Eng. Vol. 33 (1988) 1105-1112
A Supercritical/Superfluid He-Cryostat for STEP
D. Mohr*, A. Seidel*, M. Sander ~ G. Jochimsen ~ A. Wagner ~ J. Wolf +, J. Weber +, K. Petersen + *Daimler-Benz Aerospace AG, Dornier Satellitensysteme, P.O. Box 801169, D 81663 Muenchen, Germany ~ Aerospace AG, Space Infrastructure, P.O. Box 801168, D 81663 Muenchen, Germany +Linde AG, Dept. EVW, Dr. Carl-von-Linde StraBe 6-14, D 82049 H611riegelskreuth, Germany
The Satellite Test of the Equivalence Principle (STEP)-satellite shall measure the universality of free fall to an accuracy of 1 into 10 ~7. Any liquid movement in the required He-cryostat must therefore be avoided. The combination of a supercritical main cryostat with a superfluid shield under equal pressure offers the possibility to meet this requirement due to the avoidance of free liquid surfaces.
INTRODUCTION The Satellite Test of the Equivalence Principle (STEP)-satellite is a candidate for ESA's next medium size scientific mission (M3). On STEP the fundamental physical law of the universality of free fall of any body shall be checked up to an accuracy of 1 into 1017. Squids are foreseen to measure the relative motions of different test mass pairs in an absolute zero-g environment provided by a drag-free satellite control system. The scientific instruments require temperatures below 2 K in a cryostat in order to provide the necessary measurement accuracy. To avoid any disturbance of the scientific instruments by "tidal waves" which would occur on the free surface of a saturated superfluid helium (He II) bath, a combination of a supercritical He (SCHe) main tank and a connected supersaturated He II-shield tank surrounding the scientific instruments was found to be a viable candidate. A Phase A study conducted under ESA-contract established an overall satellite design with a SCHe/He II-cryostat with a main SCHe-tank volume of 346 1. This cryostat provides a cryogenic lifetime of 265days (including 20 % margin) and an operational He II-bath temperature of 1,8 K. The SCHe-tank temperature increases during the mission time from 5,1 K to 28 K. The operational SCHe-tank pressure was found to be optimum at 10 bar.
REQUIREMENTS The requirements to the STEP-cryostat are dictated by the following: Provide an absolutely quiet cooling environment for the scientific payload (test mass pairs) Dimensions and mass are strictly limited by the foreseen launcher (Lockheed Launch Vehicle -2) - The cryostat shall house the scientific payload of 710 mm length, 220 mm diameter and 158 kg mass plus 20 % margin (including tungsten and lead shields) in an ultra-high vacuum "probe vessel" at < 10-9 pa - Provide a "probe" temperature of < 2 K with a stability of 1 mK per one orbit - Provide first fundamental eigenfrequencies of: 9 longitudinal/axial direction >_ 40 Hz 9 lateral direction > 20 Hz 9 all directions on orbit > 10 Hz - Connect the payload by a total of 752 electrical wires (including shielding) - Provide a launch autonomy without He-servicing of 6 days -
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n
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ICEC16/ICMC Proceedings Deliver a He-flowrate on orbit of > 1,5 mg/sec for the satellite's drag-flee control system Guarantee a minimum cryogenic lifetime of 226days (goal 292 days) on a low sun-synchronous earth orbit of 400 km height
CRYOSTAT DESIGN To meet the satellite mass and dimensions limitations a compromise was found by placing the cryostat within a toroidal Service Module (SVM), being thermal disadvantage for the cryostat, but placing the cryostat behind a Sunshield (Solar Array) and by providing a large radiator plate to the bottom of the cryostat. This radiator, although seeing a part of the earth's surface, has as well a large field of view to free space. By this means the Cryostat Vacuum Vessel (CVV) temperature is lowered considerably. Figure 1 shows the overall STEP-satellite layout, while Figure 2 shows the outer view of the cryostat. V102
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Figure 1 Overall STEP-Satellite Radiator Panel Figure 2 STEP-Cryostat The extreme scientific payload sensitivity to any mass movement inside the satellite at signal frequency (equaling one orbit) excludes the use of a "conventional" He II-cryostat. The liquid surface in it would be subject to earth and moon gravitational influences with the consequence of the occurence of He-tides at signal frequency. This would jeopardize the whole scientific task qf STEP. Electrostatic He IIorientation might be a candidate for orienting the He II in the cryostat but severe doubts that He-tides might still exist prevent this kind of He II-orientation/fixation. The way out to solve the problem was found in the use of a hybrid supercritical He (SCHe)/He II-cryostat. Since SCHe has no free liquid surface tidal He-surface waves don't exist. On the other hand the scientific payload in the "probe vessel" requires a temperature of < 2 K. This is achieved by surrounding the probe vessel to a large extent with a supersaturated He II-shield tank of only 5 1 volume operating at the same pressure as the SCHe-tank. The pressure equality in both tanks is achieved by a connecting capillary. The shield tank He II is produced on orbit from SCHe in it and then maintained at He II-temperatures by the following way explained by figure 3 The SCHe coming from the main tank is throttled down in a Joule-Thomson Valve (JT) into the twophase He-region to a temperature of approximately 1,7 K. This two-phase He, being now colder than the
ICEC16/ICMC Proceedings
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"probe", flows through a heat exchanger (HX101) at the He IIshield tank. The two-phase liquid is evaporated there and picks up the parasitic heat of the "probe" (instruments heat dissipation, cabling and mechanical support flux)of less than 10 mW total, while keeping the temperature constant. The He which still has some small liquid rest or is just saturated vapour then cools the SCHe coming from the main tank ~. 2 @ "~_.* " " " IT@ ~JT in a counter flow heat exchanger (HX301) to a lower temperature than in the tank. The rest of "cold" still in the low-pressure He cools the SCHe in the main tank by a heat exchanger (HX201) in it. sf He CVV HS 3 HS 2 HS 1 Having passed HX201 the lowpressure He cools the 3 vapourFigure 3 He-Flow Schematic with Heat Exchanger cooled cryostat heat shields and finally (but only on orbit) the CVV. The necessary "probe" temperature is maintained by regulating the downstream pressure of the Heventline system by use of the "proportional He-thrusters" of the satellite's drag-free and attitude control system which uses the He-boiloff gas. By keeping the He-downstream pressure at the JT-exit constant within a bandwidth of + 20 Pa the temperature of the two-phase He in HX101 is kept sufficiently constant to provide the required "probe" temperature stability. The cryostat resulting from a NASTRAN and an ESATAN models analysis is shown by figure 4.
| | |174
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ICEC16/ICMC
Proceedings
This cryostat has a cylinder-toroidal SCHe-tank of 346 1 cold volume, operating at 10 bar pressure. This tank is, via a lower and an upper carbon-fibre-compound Spatial Framework, hung to the CVV on 16 support straps, each consisting of 4 GFC- and 2 CFC-chain loop elements. Three vapour-cooled shields, each with MLI (as well as the main tank), form the thermal insulation system together with the low conductive straps. The structural analysis requires for a total mass of suspended items inside the CVV of 397 kg strap cross sectional areas of 36 mm 2. The main SCHe-tank houses in its central cylinder the "probe" vessel (with the payload) with a feedthrough head for electrical connectors and its He IIshield tank and heat exchanger HX101, and the heat exchanger HX301 and the JT-valve below the probe. To insulate the cryostat effectively from the SVM, it is suspended on the SVM via 16 low-conductive GFC-struts and its outer surface, except of the radiator on its lower bulkhead, is equipped with MLI. The thermal analysis reveals a He-flowrate from the SCHe-tank as shown by figure 5 as function of mission time. Figure 6 shows the SCHe heating during the mission time. At a final temperature of approximately 28 K the JT-valve is no longer able to provide enough cold He to maintain the probe temperature. The integrated flowrate results, together with a SCHe-tank content of 49,2 kg at launch in an actual lifetime of 265 days.
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CONCLUSION With an actively cooled Cryostat Vacuum Vessel and equipped with a radiator a hybrid SCHeMe IIcryostat can be realized which provides the necessary quiet environment at < 2 K for the scientific payload of STEP while fulfilling stringent mass and volume requirements and providing a lifetime of over 6 months with 20 % margin.
Design of Thermal Shield for the ITER Cryostat
Kazuya Hamada, Kazuhiko Nishida, Takashi Kato, Tadaaki Honda, Hiroshi Tsuji, Akira Itoh, Masahiro Nakahira, Susumu Shimamoto, Michael E.P Wykes*, Robert Bourque*, Isamu Ohno** and Yasuyuki Miyauchi** Japan Atomic Energy Research Institute, Naka Fusion Research Establishment, 801-1, Muko-yama, Nakamachi, Naka-gun, Ibaraki, 311-01, Japan, *ITER Joint Central Team, Naka Joint Work Site, 801-1, Muko-yama, Naka-machi, Naka-gun, Ibaraki, 311-01, Japan, **Ishikawajima Harima Heavy Industry Co., Ltd., 6-2, Marunouchi, 1-chome, Chiyodaku, Tokyo, 100, Japan
Japan Atomic Energy Research Institute is studying a thermal shield in a cryostat, whose diameter is 36m and height is 34m, for International Thermonuclear Experimental Reactor. The thermal shield is cooled by 80-K helium gas. The designed radiation heat leak from room temperature to 4-K superconducting coil and the structure is less than 0.24 W/m 2 by using a multi-layer metal plate. Total heat load of thermal shield is 147 kW at 80K. The thermal shield is a segment structure in order to allow a maintenance and replacement by remote handling system.
INTRODUCTION Japan Atomic Energy Research Institute (JAERI) participates in design activities and research and development for the International Thermonuclear Experimental Reactor (ITER) in collaboration with European Union, Russian Federation and United States of America[1 ]. In ITER, superconducting coils, such as Toroidal Field (TF) coils and Poloidal field (PF) coils are used and the total weight is around 25,000 tones. The coils are forced flow cooled by using supercritical helium at 4.5K, 0.6MPa and are installed in the cryostat for vacuum insulation. In order to reduce a radiation and conduction heat load from room temperature to 4-K coil, a inner surface of cryostat is covered by thermal shield system cooled at 80 K. The PF coils are operated at pulse current and induce an electro-magnetic force and a Joule heat caused by the induced current on the thermal shield. The thermal shields are designed to safety withstand these loads and to meet the heat load and cooling condition requirements. Additionally, they are designed to be maintained using remote handling techniques. DESIGN CONDITION AND STRUCTURE The cross section of Cryostat is shown in Figure 1. The outer diameter and the height of cryostat is 36 m and 34m, respectively. Thermal shield are installed over the inner surface of cryostat and around of the many access port to the plasma vacuum vessel, and to coils gravity support. Usage of liquid nitrogen for cooling of thermal shield is avoided because of radioactivation. Therefore the thermal shield is cooled by 80-K, 1.6-MPa helium gas. The requirement and the design criteria for thermal shield are as follows; @Thermal shield is cooled by 80-K helium gas and the inlet temperature/pressure is 80K/1.8MPa and outlet temperature is less than 100K. Pressure drop of overall thermal shield is less than 0.1 MPa. @ Heat flux from thermal shield to 4-K objects is less than 0.24 kW/m 2 @ Heat flux from room temperature to thermal shield is less than 9 W / m 2 for cryostat and 3.6 W / m 2 for 427
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ICEC16/ICMC Proceedings
port and gravity support. @to allow a maintenance and replacement work by remote handing techniques @to withstand thermal contraction and electromagnetic force resulting from pulse operation of PF coils @to withstand an accumulative neutron and gamma dose of 1 X 107 Gy The thermal shield is a segmented structure, as shown in Figure 2. The estimated total surface area of thermal shield is 20,230 m 2. The dimension of one segment is determined as 4 m X l m, to be compatible with remote maintenance and replacement. Each segment is connected by a cover plate which bridges of the gap between neighboring segments so as to block the heat leak from room temperature. On the shield panel, cooling pipe are welded. 36m
._nCryostat v]
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Figure 1
Cross section ofITER cryostat
Figure 2
Thermal shield of upper ceiling
THERMAL DESIGN To accomplish the design requirement of heat flux, a conventional multi-layer aluminum deposited polyester (super insulation) and multi-layer metal plates are compared. The characteristics of each insulation method is shown in Table 1. Table 1
Comparison with insulation method
Multi-layer metal plates Fabrication procedure is complicated Quality control is easy Depend on the work procedure Performance good weak for mechanical property Accumulative Neutron and Gamma On account of superior radiation resistance, multi-layer metal plates structure are selected as thermal Fabrication
Aluminum deposited polyester Available
ICEC16/ICMC Proceedings
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shielding for the cryostat. In order to accomplish the design condition of heat flux, the needed number of layer and the available emissivity are estimated as 45-layers and 0.3, respectively. The shield material is a stainless steel plate, which the thickness is 0.1 mm, and aluminum is deposited. The schematic diagram of multi-layer metal shield is shown in Figure 3. The thickness of shield is 95.4 mm and the shield is supported by Titanium rod. The static heat load and the helium gas mass flow rate are calculated and listed in Table 2. Total heat load is 147 kW and removed by 1427.2-g/s helium gas flow. A block flow diagram of cooling path is shown in Figure 4. The cooling path is distributed for 3 cooling objects, such as cryostat wall,, port and gravity support. Each cooling path consists of 4 blocks for circumferentially. Such a paralleled configuration results in a low pressure drop for cooling down operation and small temperature difference within thermal shield. The pressure drop is estimated as less than 0.1 MPa at 80K. It is assumed that the friction factor is expressed as Blasius formula. S h i e l d Support
q5 5 m m P o l y m i d e rod
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Schematic diagram of cross section of thermal shield (from vertical direction) Static heat load condition of thermal shield Part of thermal shields Cryostat Port
Gravity Support of Coil Total
Heat load of thermal shield 41.6 91.1 14.3 147
Mass flow rate 404 g/s 885 g/s 138 g/s 1427 g/s
kW kW kW kW
Cryostat Upper Ceiling Cover h Cryostat Side wall Cryostat Bottom Cover
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~-~
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ICEC16/ICMC Proceedings
The helium gas is supplied by warm helium circulation compressors located at room temperature and the gas is cooled by liquid nitrogen. W E I G H T LOAD CONDITION The weight of one segment of shield panel is estimated as 270 kg and total weight of thermal shield is estimated as 1,370 tons. 0.2 G is assumed as seismic load condition. The PF coil is operated by pulse .current and the most serious condition for the thermal shield is a discharging operation such as a quench. The most adverse electro-magnetic force occurs at the location induced in Figure 1, during discharge following coil quench. At the thermal shield panel, nearly 1 T is applied by PF-4 coil and dumped at 15seconds time constant. The calculation results are summarized in Table 3. Table 3
Load condition per one segment of shield
Self-weight 270kg
Seismic load Vertical direction 954. lkg Horizontal direction: 80.1 kg
Electro-magnetic force Vertical direction" 22kg Horizontal direction: 65.4kg
In case of a segments located on the vertical wall of the cryostat, the self weight and seismic load are supported by two Titanium ( qb 10 mm) rods from cryostat wall. For a segment which is located at top and bottom of cryostat, the segment is designed as self-standing. CONCLUSION The conceptual design of thermal shield for ITER cryostat is carried out. It is estimated that the required shield performance is less than 0.24W/m 2 and the thermal shields is designed by conventional technique. ACKNOWLEDGMENT The authors would like to thank Drs. M. encouragement and support on this work. frame work of the ITER EDA agreement. ITER Director, the parties to the ITER EDA
Ohta, T. Nagashima and S. Matsuda for their continuous This report is an account of work for undertaken within the The view of authors do not necessarily reflect those of the agreement, or the International Atomic Energy Agency.
REFERENCE
1 Thome, R., Design & Development of the ITER Magnet System, Cryogenics (1994) Vol.34 ICEC Supplement 39-45
DESIGN OF SUPERFLUID-COOLED CRYOSTAT FOR 1 GHz NMR SPECTROMETER
Akio Sato*, Tsukasa Kiyoshi*, Hitoshi Wada*, Hiroshi Maeda*, Satoshi Itoh** and Yoshio Kawate** * National Research Institute for Metals, 1-2-1 Sengen, Tsukuba 305, Japan ** Kobe Steel, Ltd., Electronics Research Laboratory, 1-5-5 Takatsukadai, Nishi-ku, Kobe 651-22, Japan
The basic design of one gigahertz NMR spectrometer is being carried forward. A magnetic field of more than 23.5 T for this spectrometer will be achieved by the superfluid cooling technology. This paper will describe the basic design for the superfluid helium cryostat for an outer superconducting magnet with a cold bore of about 150 mm. Some technical points have become clear. Safety of the cryostat involving a magnet with a huge stored energy of 50 MJ has been checked. The amount of cryogen in the magnet vessel should be less than 100 L. The consumption rate of 708 cc/hr has been estimated.
INTRODUCTION Tsukuba Magnet Laboratory (TML) of National Research Institute for Metals has started the second Multicore research project for the development of a one gigahertz NMR spectrometer. A magnetic field of more than 23.5 T for this spectrometer is a challenging target, but will be achieved, using a newly developed oxide superconducting coil in a backup field of 21.1 T. The cryogenic system consists of two sections: one is a 4 K pool boiling part for the oxide superconducting insert coil, and the other is a superfluid helium cryostat for an outer superconducting magnet with a field over 21 T in a cold bore of about 150 mm. Superfluid cooling is one of the key technologies necessary to achieve a high field of more than 21 T in such a large bore, effectively increasing the critical current for superconducting wire. The magnet design is being carried out for several types. The final design will depend on superconducting wire development status in the coming few years. We took the following magnet size for granted in the cryostat design. The magnet size was supposed to Nb3Sn SC joint be 1200 mm in diameter and 1500 mm in height. The total weight of the magnet will be 8 tons. The magnet has a peculiarity in its superconducting joints; inner Nb3Sn coils have superconducting joints 400 mm above E E the upper surface of the coils as shown in Figure 1. 0 0 This paper will describe a basic cryostat design for this magnet, and point out problems to be solved for the superfluid helium cryogenic system for an NMR spectrometer.
1
NbTi SC joint
BASIC DESIGN OF CRYOSTAT Pressurized superfluid helium cooling has been adopted to achieve long term operation and cryogen reduction in a magnet cryostat, resulting in cryostat compactness and safety insurance in case of a magnet quench. A helium I vessel that supplies coolant to the heat exchanger for cooling helium II is located in the doughnut space around Nb3Sn superconducting joints as shown in Figure 2 and 3. The helium I vessel capacity is 650 L. 431
SC shim coil
SC coil
Figure 1 Arrangement of the NMR magnet and superconducting joints
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Power leads will be removed after energizing the magnet and current transfer to the Persistent Current Switch (PCS). The PCS will be situated in the helium I vessel, preventing heat Mainterminal loss during the energizing process, whereas protection resistance and diode will be located MainPCS in the helium II vessel. Maindumpresister ~ ~ Shimcoildumpresister The superfluid helium acts as a superconductivity stabilizer especially for bare superconducting wire because of its excellent Helium II vessel heat transfer characteristics. In our case, howi:i: ever, all of the coils are impregnated solenoid, Shim coil SC joint so the magnet is almost in an adiabatic condiMainSCjoint (NbTi)----.T tion in the superfluid helium. The superfluid i helium thus acts as only a temperature stabi[ /lI~'~l Shim coil Maincoil l lizer. The amount of helium II coolant is a kind of trade-off between temperature stabilization and cryostat security. In this design, a helium II vessel coolant capacity of less than Figure 2 Schematic view of the cryostat. 100 L has been selected. The heat capacity of The helium I vessel is located in the doughnut 100 L superfluid helium at 1.8 K is 43.4 kJ/ space around Nb3Sn superconducting joint K, so an accidental decrease of cooling power of one Watt would increase the superfluid bath temperature by only 1.38 mK/min. It means that the heat capacity is large enough for temperature stabilization. 100 L of coolant is very small compared with the volume of the magnet. Reduction of the coolant amount in the helium II vessel to less than 100 L will be achieved by filling some compound in the magnet vessel.
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SAFETY ANALYSIS IN CASE OF QUENCH In the case of a quench the magnet will release its storage energy of about 50 MJ. The storage energy will be consumed as Joule heat loss in the magnet, and the magnet temperature will increase up to about 80 K in a few seconds. The superfluid helium around the magnet will begin to increase its temperature and will then evaporate at saturated temperature in the closed helium II vessel. The only exit of evaporated gas is a cold safety valve connecting to the helium I vessel. The flow rate will achieve to a maximum value of 340 L/s in the case where release pressure is 0.14 MPa. Therefore the maximum flow rate through the cold safety valve is estimated to be 4.3 m/s when the valve diameter is 50 mm. All of the liquid coolant inside the magnet vessel will be changed to gas in 1.6 seconds if helium II coolant is 80 L under film boiling assumption, and the release flow rate will decrease to 267 L/s. In this estimate, a cold safety valve of diameter of 50 mm has sufficient capacity for the relief valve. For redundancy, in case of an accident with the cold safety valve, a rupture disk that will release high pressure gas to adiabatic vacuum space
l_
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Figure 3 Overview of the designed cryostat
ICEC16/ICMC Proceedings and a drop-off valve on the outer shell of the cryostat will be fitted.
433
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THERMAL DESIGN
-13I . F,F'b I The superfluid heat exchanger in the magnet vessel cools down the helium under conditions of at4 K shield I ... ..... .~!i mospheric pressure. The superfluid heat exchanger is exhausted by a 200 L/min vacuum pump in a steady state at 1.8 K. An alternative vacuum pump is used when maintenance is required. An auxiliary pump with an exhausting rate / of 6000 L/min will be used in the precooling pro/i cess from 4 K to 1.8 K, resulting in a reduction of 1 | J ] [did d~ "] .... u cooling time to half a day. Coolant supplied from the helium I vessel 80K shield Gas Coold Shield (GCS) through the Joule-Tomson (JT) valve flows in a mist state through the heat exchanger, and evapoFigure 4. Arrangement of the magnet vessel and rates on the inner surface of the heat exchanger to the thermal radiation shield. cool down the outside helium in the magnet vessel. Temperature differences between pressurized helium and saturated helium in the heat exchanger are controlled by the JT valve in a certain value to keep the magnet temperature steady. This temperature difference control method was adopted in the Tohoku Univ. magnets and the stable controllability has been confirmed [ 1,2]. The supporting rods and thermal radiation shields location is shown in Figure 4. There are three types Pumpout of thermal radiation shield including Gas Cooled Shield (GCS). The 80 K- shield is cooled by liquid nitrogen, and the 4K-shield for the 1.8 K magnet 'Q=" vessel is cooled by helium I. The magnet vessel is I t Qc.'2 supported directly from the outer vessel by an FRP supporting rods thermally-anchored at 4 K and 80 K and the intermediate GCS temperature, respectively, considering mechanical stiffness. Thermal flows are shown for this case in Fig~ Q r Qr ure 5. An example of the thermal balance calculation result is summarized in Table 1. Total consumption rate of liquid helium is 708 cc/hr. Liquid ".t.Qo.i helium is mainly consumed in the insert Dewar 9~,2 ii ~ that is designed so as to exchange an oxide superOuter v a c c u m e shell conducting magnet, that would be under developQch Conduction through neck tube (helium) ment even in near future. Qcn Conduction through neck tube (nitrogen) We plan to supply liquid helium at the rate Qcs Conduction throuh supporting rod Qcl Conduction in liquid helium of one 500 L Dewar every month. The consumpQr Radiation Qsv Super-leak of liquid helium tion rate almost satisfies this specification. ::
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Figure 5 Thermal flow model for calculation. PROBLEMS IN LIQUID HELIUM SUPPLY SYSTEM The liquid helium supply will seriously affect the temperature distribution in the helium I vessel, especially in the bottom part of the vessel where lambda point superfluid helium layer is formed in a steady state. Any temperature increase around the communication channel such as the safety valve will induce an effective cooling power decrease. Any temperature change would affect the persistent current of the superconducting magnet with many superconducting joints and induce a magnet quench in the worst case. Therefore one of the key technologies necessary for a stable long term operation of an NMR spectrometer is a reliable liquid helium supply system. A detailed design of the liquid helium supply line exit and the helium I vessel struc-
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Table 1 An example of the thermal balance calculation result, where OVC and GCS are the Outer Vacuum Can and the Gas Cooled Shield, respectively; LN 2 means thermal shield of liquid nitrogen or the stage of 77K; He I and He II also represent its temperature stage ( See Figure 4 and the thermal flow model of Figure 5). The result was calculated for the case of Figure 4.
Heat Flow Path
OVC
~ LN 2
Radiation 50.6 Supporting Rods 2.04 Neck tube (He) 0.519 Neck tube (N2) 0.229 Insert Dewar 8.20 Feedtrough Communication Channels Total
61.6
LN 2 ~ GCS 0.997 0.162 0.045
LN 2 ~
He I
8.45 xl0 -3
0.0190 0.133 0.0164
He I ~ He II 7 . 5 3 x l 0 -6
2.99 xl0 -3
0.216
0.593
1.80
GCS ~ He I
8 . 4 5 x l 0 -3
9 . 1 2 x l 0 -3
5.76 xl0 .3 0.112
0.394
0.121 ( in Watt)
Liquid helium supplied to the superfluid heat exchanger 5.75 xl 0 .3 g/s 166 cc/hr @ 4.2 K; 1 atm Exhausting rate from the superfluid heat exchanger 90.8 L/min @ 300 K Boil-off rate of liquid helium ( Helium I vessel ) 184 cc/hr Boil-off rate of liquid helium ( Insert Dewar ) 359 cc/hr Total consumption rate of liquid helium 708 cc/hr @ 4.2 K; 1 atm Total consumption ra~e of liquid nitrogen 1450 cc/hr @ 77 K; 1 atm
ture will be carded out in the next stage, avoiding disturbance around the communication channel.
SUMMARY In this basic design some technical points have become clear. The amount of cryogen in the magnet vessel is designed to be less than 100 L, which will ensure cryostat insurance in the case of a magnet quench. Safety of the cryostat involving a magnet with a huge stored energy of 50 MJ has been checked. Precooling of the magnet in the restricted space is a remaining problem. The estimated consumption rate of 708 cc/hr for liquid helium will allow the monthly supply system by one 500 L helium Dewar There still remain other problems to be solved. Precooling of the eight-ton magnet will take more than three weeks. Magnet charging will take a few days. In all processes, including normal operation mode, the cooling system operate automatically and stably. We will continue the cryostat design and research and development works to obtain design data.
REFERENCES Watanabe, K., Noto, K., Muto, Y., Maeda, H., Sato, A., Suzuki, E., and Uchiyama Y., Research and Development for Pressurized He II Cooled Superconducting Magnets Sci. Rep. RITU (1986)A-33 297- 306 Noto, K., Watanabe, K.,, Muto, Y., Sato, A., Horigami, O., Ogiwara, H., Nakamura, S., Suzuki, E., Proc. ICEC- 10 (1984, Helsinki), 181 - 184
A Variable Temperature Cryostat to Measure
Jnoncu(T) of ITER Strands up to 20 Teslas
Jager B.*, Bocquillon A.*, Chaussonnet P.*, Martinez A.*, Nicollet S.*, Serries J.P*. and Vallier J.C.** *Association Euratom-CEA DRFC, CE Cadarache, F 13108 Saint Paul Lez Durance, France **CNRS - High Field Laboratory, BP 166, F38042 Grenoble, France
The determination of the non-copper critical current density Jno,,cu is crucial for the development of Nb3Sn strands for ITER coils. The temperature dependence of J,,onCu in a magnetic field plays a leading part in determining the operating parameters and stability of the ITER coils. We describe the development of a variable temperature cryostat based on a new concept. This new cryostat is able to measure critical currents up to 600 A in fields up to 20 T in a temperature range of 4.2 K to 20 K.
INTRODUCTION The different conductors of the ITER magnets are made of cables of superconducting strands. The ITER specifications on the critical properties of these strands mainly concern the critical current density to be achieved at 4.2 K and 12 T and the level of hysteretic losses. The model predicting the critical current density of Nb3Sn as a function of the field and the temperature has been historically presented by Summers [1]. This model involves several parameters and coefficients which need to be carefully characterised for the practical industrial wires used for ITER. Indications already exist on the important variations which take place from one strand to another and this is not surprising in view of the great variety of processes used 9internal tin, modified Jelly Roll, bronze route. More precisely the behaviour of the strands as a function of the temperature has to be checked for every strand. In fact the operating temperature taking into account the design of the cables is never 4.2 K, the usual test temperature for strand benchmarks but generally temperatures greater than 5 K. The new cryostat presented here must help in obtaining information in this direction. It will be able to measure critical currents up to 600 A in the 4.2 K to 20 K range and in magnetic fields up to 20 T. The design allows to insert the cryostat inside the bore of the hybrid magnet of the CNRS High field laboratory.
PRINCIPLE OF THE CRYOSTAT The difficulty in measuring the critical current of ITER type superconducting strands, in magnetic fields which can reach 20 teslas, is to supply a current of about 600 A to the strand to be tested without temperature modification. To solve this problem, we used a new principle (cf. Figure 1). Two current leads are cooled by helium circulation, their two cold ends are connected by a sample support making up a thermal shunt. The superconducting strand to be tested is wound around this support and thermalized by helium which is at the same temperature. The electrical resistance of the support must be high enough for the current going through it to be insignificant, when the strand to be tested is in superconducting state.
DESCRIPTION Test cryostat This is an insert-cryostat which is placed in an existing 80 K cryostat. The general diagram is shown in Figure 1. It essentially includes a helium tank of about 18 liters (height = 1.08 m, external diameter = 0,1643 m, internal diameter = 0.0761 m). This annular tank is pressurised up to a maximum pressure of 435
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0.29 MPa. It therefore allows to perform measurements in supercritical helium. A pressure/backpressure regulator system regulates the pressure to the desired value. The helium extracted through the lower part supplies the two current leads. If liquid helium is used, an evaporator/heater helps to supply the current leads in helium gas at an adjustable temperature. Two control valves, placed at the outlet of the current leads, regulate the temperatures of the two cold ends at the same set point (5 K < set point < 20 K). The stability of the temperature is then better than 0.05 K. The sample holder is connected to the cold ends of the current leads. All of the test cryostat is maintained under vacuum by a cryopump. This pump is made by sticking coconut charcoal grains to the bottom of the helium tank. The lower part of the cryostat is removable so as to allow the sample to be introduced. An indium seal ensures tightness versus vacuum. Current leads We used the same technique as that described before [2]. The main characteristics of these current leads are described in Table 1. They consist of a copper braid cooled by helium. This braid is introduced into an insulating tube. This tube is then introduced into a stainless steel tube ensuring tightness versus the external vacuum. This technique helps to avoid cold insulating electrical breaks on the helium supply tubes. The ends of the braids are soldered in OFHC copper end pieces. The temperature of the cold end of these current leads is regulated by the TCV valves (of. Figure 1), which act on the helium flow rate. The electrical and thermal connection with the sample is ensured by a pinched connection on the cylindrical ends of the current leads. The surfaces in contact are machined with care and gilded. Magnetic field temperature measurements The sample temperature must be precisely known whatever the magnetic field. We used new Cernox sensors (type CX 1030) specially developed by Lakeshore for magnetic field measurements. Three sensors are mounted on the sample holder to verify the possible temperature gradient in the presence of current. Two other sensors are mounted on the cold ends of the current leads and help to ensure their
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Table l Main characteristics of current leads copper wire diameter number of wires A, copper cross section L, length of the cooled braid L I/A at I = 600 A
Electrolytic Tough Pitch - 85 < R R R < 100 0.3 m m 624 44.11 10 -6 m 2 1.925 m 2.62 107 A/m
temperature regulation. We measured the effects of a magnetic field up to 20 teslas on these sensors. These measurements were made at constant temperature in a boiling helium bath at 4.21 K. Figure 2 shows, depending on the magnetic field, the errors in terms of temperature for the five sensors. The m a x i m u m error made at 4.21 K remains inferior to + 0.02 K for a field of about 15 teslas. For a AllenBradley carbon resistor, this error is o f - 0.14 K at 20 teslas and continues to increase. Each sensor is mounted in a calibrated hole with grease. Sample The sample (Figure 3) is connected to the current lead ends by two high-purity copper bus bars (RRR 400) which are as symmetrical as possible to minimise the temperature gradient. The reaction mandrel is a V A M A S titanium mandrel which is also kept for the test without any transfer and possible associated degradation. Two copper ends are soldered on the titanium mandrel for the sample terminations. A thin copper cylinder is added after reaction so as to ensure a very uniform temperature of the mandrel. This piece is not in electrical contact with the strand so as to avoid any low resistance short circuit. Two Cernox sensors are installed at the top and the bottom of the sample. The third one is installed in the very middle of the sample. The insertion of the sensor in front of the central turn of the sample is possible by drilling a transversal channel in the mandrel. Two voltage couples, one across the central turn and the other across the 7 turns monitor the resistive transition of the sample.
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with Tcom= 18 K
CONCLUSION This new cryostat allows the measurement of critical current (I < 600 A) to be performed in ITER-type superconducting strands within a large temperature and magnetic field range. It uses a new principle of thermalization of the superconducting strand. We validated its principle by measuring the critical temperature Tco of a Nb3Sn strand in its self field. The measurements under magnetic field will start end of 96 at the CNRS High Field Laboratory in Grenoble
REFERENCES Summers L. T. et al, Characterization of internal tin Nb3Sn superconductor for use in the proof of principle coil, IEEE trans, mag., Vol. 27 N~ March 1991, 1763-1766. Jager B. et al., Test facility for joints of subsize cable in conduit conductors of NET/ITER winding studies, Advances in Cryogenic Engineering, 41, Plenum Press, New York, to be published (1995)
Tests of the Cryostat for 1.3 GHz Superconducting Cavity at T < 1.8 K
Filippov Yu.P., Kovrizhnykh A.M., Batin V.I. and Uchaikin S.V. Joint Institute for Nuclear Research, Particle Physics Laboratory, Dubna 141980, Russia A system for experiments with the 1.3 GHz superconducting cavity has been designed, built and tested. The major part of the system is a horizontal cryostat which includes a helium vessel for the cavity, He-gas and LN2cooled shields, two nodes for RF couplers, a device for mechanical stretching/squeezing the cavity and a control system. Cooling down to 2.0 K and below is performed with saturated superfluid helium pumped by 500 1/s unit. Operation experience collected up to date is discussed. The feature of the control system is the usage of 16 channels cold multiplexer to measure temperatures within the liquid helium vessel. Its components are tested for the radiation environment performance. The obtained results are presented for the various dose values.
INTRODUCTION Our previous report [1] described the design of the cryostat for the special 1.3 GHz RF-cavity and the preliminary test results. There are two forthcoming stages of the project completion. The first is the cryogenic test that is the main subject of this report. In the beginning of this year, the cryostat was fully assembled and prepared for these tests. At the second stage, it is planned to connect the system with the electron gun at the Particle Physics Laboratory (PPL) for a beam test. At the beginning, we are going to use the S-band fivecell cavity prepared by our colleagues from PPL. Such a test requires a comparatively big number of sensors: temperature, pressure, level and so on. In order to reduce the number of the wires coming into the helium vessel and, perhaps, avoid the signal deterioration, we have an intention to use the cold multiplexer, MUX, if the MUX operates under the corresponding radiation fluences, that is the other subject of the present report.
CRYOSTAT AND TEST RESULTS The side view of the cryostat and the sequence of assembly procedures are presented in Reference 1. Figure I shows separate nodes of the cryostat before the final assembly. The cavity has to be gathered with two special bellows nodes to decrease the heat leak into the helium vessel from the ambient space: one of these nodes is seen at the beam pipe going into the helium vessel. At the foreground, one can see two rods to tune the cavity by its stretching or squeezing along the beam line. The cavity should be inserted into the helium vessel. The helium shield cooled with the pumped vapor, is mounted around the helium vessel. The cap of the helium vessel is supplied with the window to connect the corresponding wires within the vessel: it is located on the opposite cap (invisible in the picture). At the background, one can see the vacuum shell with the nitrogen shield inside. The nodes of inlet and outlet of liquid nitrogen, feeding line of liquid helium and the line to pump the helium vapor, two couplers located under the angle of 25 degrees, are mounted on the vacuum shell. The helium vessel with the cavity inside and the helium shield outside of it, have to be put into the nitrogen shield" its position is regulated with the corresponding flange supports. The centering is 439
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Figure 1 Cryostat during assembly. provided by means of adjusting screw of the supporting node which is at the bottom of the vacuum shell. The outer dimensions of the cryostat are: diameter - 680 m m and length - 920 mm. The position of the cryostat may be regulated with four screws in the legs of the supporting platform. To seal the flanges at cryogenic temperatures, we use the indium or annealed copper gaskets allowing one to replace or remove the cavity if necessary. After pumping the helium vessel down to the given pressure, the necessary compensation of the changed force may be performed with the help of the tuning device (four screws in this case). Figure 2 shows the assembled cryostat during testing. The scheme of the cryogenic system was described in Reference 1. To maintain and regulate the temperature regimes of the cavity, the stabilizing pressure system is used. It consists of the PID controller and a pneumatic valve. T h e rhodium/iron resistance thermometer and the resistance thermometers of the TVO type [2] are used for cryogenic tests. The Rh/Fe thermometer was fabricated and calibrated at the CRYOMET VNIIFTRI (Russian National Scientific & Research Institute for Physical and Radiotechnical Measurements, Mendeleyevo, Moscow region). The tests have shown that the heat leak into the helium vessel from ambient space, Qa~, is about 3.5 W when the temperature of liquid helium is of 4.25 K. The thermal energy emitted with the cavity, Q ~ , has been simulated with the heater: its range is from 1 to 4 W. At the total heat load of 7.5 W (Q~m + Q ~ ) , the 0.5 m3/s pump easily maintained the temperature of 1.8 K in the helium vessel: the minimum temperature of 1.43 K is achieved at the 4.5 W heat load. In our experiments we need to operate about half of an hour in the quasi-static conditions for the diagnostic purposes. So, no liquid helium can be fed into the cryostat during this time. The volume above the cavity is about 10 liters for the L-band case and a bit more in the case of the S-band cavity. So, at the specified value of thermal inleak, such a volume of He allows sufficient operation for the S-band and L-band cases, correspondingly.
COLD MULTIPLEXER AND ITS RADIATION TESTING Thin film rhodium-iron and Cernox TM resistance thermometers are well known as suitable sensors in order to operate in the conditions combining rather high radiation fluences and cryogenic temperatures [3]. However, their costs are rather high. So, we are going to use the TVO resistance
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Figure 2 Assembled cryostat under cryogenic test. thermometers for beam testing because of their high tolerance to radiation environments with reasonably good accuracy [2] and comparatively low price. As mentioned, the aim of using multiplexer is to reduce significantly the number of leads entering the cryostat. It is achieved with two high-speed 16 channels HCMOS-multiplexers of PC74HC/4067 type. These chips are placed within the helium vessel and switch the potential wires from one sensor to another. In the result, only 7 wires are used instead of 34 to measure the resistance of 16 sensors connected inseries. These wires are for biasing current, output register connected with the counter in the helium vessel, supply, and amplifier. One can find another way of measurement in Reference 4. The tests of the MUX board have been performed under the radiation environment conditions. The experiment was arranged at the Pulse Reactor IBR-2, Laboratory of Neutron Physics of Joint Institute for Nuclear Research. Scheme of the board for testing the components under irradiation is shown in Figure 3. The parameters of testing are as follows: temperature in the cryostat 77 K; 3 . 7 - l 0 s neutron/(cm2s); fluence of neutrons gamma radiation 400 Gy/hour; "~1,.~ MeV. average energy During the "on line" experiment, the following characteristics of multiplexer were measured: the resistance of the open channel, Ro,~, matching of the channel resistances, ARon, the dynamic current of supply, I~=pv. d The obtained results are presented in the Table 1. The data have shown that the main characteristics became worse gradually with the dose increase of radiation. A sudden failure of the multiplexer occurred at the total dose of about 14,500 Gy. As Table 1 Test results of MUX for gamma plus neutron irradiation experiment. gamma radiation total dose, Gy neutron total fluence, neutron/cm 2
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consumption increases correspondingly. It is difficult enough to determine The leakage currents of simultaneously closed transistor keys seem to appear influence of the neutron flow mainly. In conclusion, one can note that the requirements completely for the planned beam test.
CONCLUSION The horizontal cryostat for experiments with superconducting S- and L-band cavities has been assembled and tested down to 1.43 K. The measured heat leak into the helium vessel from ambient space is about 3.5 W at 4.25 K. It allows one to operate about 40 minutes in the quasi-static regime when no liquid helium can be fed in the cryostat. In the environment combining gamma and neutron radiations and cryogenic temperatures, up to the total doses of approximately 10,000 Gy and 3. l013 n/cm 2, correspondingly, the cold multiplexer based on the HCMOS PC74HC/4067 chips, can be used successfully. To define these boundaries more exactly, further explorations are needed.
ACKNOWLEDGMENT We would like to thank the personnel of the Nuclear Reactor at the Laboratory of Neutron Physics of our Institute for their assistance with this work. We express our deep gratitude to our colleague Dr. A. P. Cheplakov for his help to arrange this experiment.
REFERENCES 1. Filippov, Yu. and Kovrizhnykh, A. Cryostat for beam test with 1.3 GHz SC-cavity at 1.8 K Cryogenics (1994) 34 ICEC Supplement 769-772 2. Datskov, V. Cryogenic metrology and investigation of breakdown heat regimes in superconducting accelerators. Cand. Sci. Dissertation (1985) Joint Institute for Nuclear Research, Dubna (in russian) 3. Scott Courts, S. and Scott Holmes, D. Effects of cryogenic irradiation on temperature sensors. In: Advances in Cryogenic Engineering (1996) Plenum Press, NY, (to be published) 4. Filippov, Yu., Sergeyev, I. and Uchaikin, S. Multichannel system for fast measurement of temperatures at 1-4.2 K Cryogenics (1994) 34 ICEC Supplement 413-415
Comparison of Floating and Thermalized Multilayer Insulation Systems at Low Boundary Temperature
Gerard Ferlin, Berthold Jenninger, Philippe Lebrun, Guillermo Peon, Germana Riddone and Balazs Szeless LHC Division, CERN, CH-1211 Geneva, Switzerland
The Large Hadron CoUider (LHC) will be a 26.7 km circumference particle collider using high-field superconducting magnets operating in superfluid helium. An efficient and robust thermal insulation system is therefore required to minimise the residual heat inleak to the large surface area at 1.9 K constituted by the stainless steel wall of the helium enclosure. The baseline solution uses "floating" multilayer reflective insulation. An altemative consists of a combination of multilayer reflective fdms and a soft screen, partially thermalized to the 5 K level and supported away from the cold wall by net-type insulating spacers. We establish the improvement potential of the alternative over the baseline solution, and compare their insulation performance on the basis of measured characteristics of thermal contacts and spacers.
INTRODUCTION Large cryogenic projects, such as superconducting particle accelerators and colliders, expose large areas of cold surface to residual heat flux from the ambient-temperature environment, and therefore demand efficient and robust thermal insulation systems. This is particularly true of the Large Hadron Collider (LHC), presently under construction at CERN [1], with 1600 high-field superconducting magnets operating in superfluid helium [2] distributed around the 26.7 km circumference of the machine tunnel. The residual heat inleak from the 75 K thermal shield, which reaches the 50'000 m2 of surface area at 1.9 K presented by the stainless steel wall of the helium enclosure, represents the single largest contribution to the thermal budget of the cryostats [3]. This was identified early in the project, and an experimental program initiated, to investigate and qualify thermal insulation systems at low boundary temperature on representative samples and geometries. First results [4] confirmed the advantage of multilayer systems over single reflective surfaces for coping with degraded insulation vacuum, a situation bound to happen locally over the circumference of the machine, and further tests [5] sought to improve the performance of "floating" multilayer systems by limiting parasitic conduction between layers. Cryostats for prototype cryomagnets [6] and a full-size thermal model [7] were built in industry and operated with such an insulation system, bringing a real-scale verification of the thermal performance measured on samples and thus validating thermal budget estimates for the LHC [8]. In view of the importance of this source of heat inleak to the 1.9 K level, we investigate in the following the potential of improvement presented by an alternative combining multilayer reflective films and a soft screen, actively thermalized to the 5 K level and supported away from the cold wall by net-type insulating spacers. After establishing the range of thermal impedances of interest for thermal contacts [9] to the 5 K pipe, and insulating spacers [ 10], we show how these can be realised in practical systems, and present results of measurements on test samples.
POTENTIAL OF A THERMALIZED MULTILAYER INSULATION SYSTEM Figures 1 and 2 show the principle and the thermal flow scheme for floating and thermalized multilayer superinsulation systems. Heat flux through multilayer insulation is given by the combination of solid conduction, residual gas conduction and radiation. 443
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THERMAL CONTACT TO COOLING PIPE A dedicated experiment (figure 4) has been performed to measure the impedance of thermal contacts on a geometry, materials and assembly techniques of industrial nature, representative of heat intercepts to be implemented in large numbers in LHC cryostats. The heat intercept is constituted of a conductive sleeve, made of aluminium, shrink-fitted onto a DN65 stainless steel tube and connected to an isothermally heated block by means of aluminium strips. The thermal impedance between the aluminium sleeve and the stainless steel tube is evaluated from temperature measurements with a precision of +10 %. The temperature difference shows, as expected, a linear dependence on the heat flow Q. The thermal contact between the aluminium sleeve and the stainless steel tube has been tested at different temperatures and performs as follows: 1.2_+0.1 K/W at 4.5 K and 0.5_-+0.05 K/W at 20 K. Samples with different diameters and materials will be tested in order to investigate the consequence on thermal impedance.
NET-TYPE INSULATING SPACERS The insulating spacers considered consist of a set of piled-up nets of non-metallic, low-thermal conductivity materials which provide support at low residual heat flux. Different type of nets are alternatively inserted to avoid large contact area due to accidental superposition. A specific experiment (figure 4) has been developed to measure, by means of a heatmeter [ 11], the heat flux between two circular aluminium plates separated by spacers. The cold plate temperature is at 2 K and that of the warm plate, which simulates the soft screen, can be varied from 10 K to 30 K. The compressive force can be increased from 11 N to 94 N. The spacers are made of a combination of two types of glass fiber nets (N 1 and N2) from industry. N1 has a 10xl0 net grid with a thread thickness of 1 mm. N2 has a 6x6 net grid with a thread thickness of 0.5 mm. A spacer composed of 3 layers of N 1 and 2 layers of N2 have been tested. For a warm temperature of 25 K and with a compressive force of 34 N, the measured heat flux to 2 K at 10-4 Pa is 10_+1 mW, which gives a thermal impedance of about 70 K.m2/W.
COMPARATIVE PERFORMANCE Considering a thermal contact resistance of 5 K/W (contact over 30 % of length of soft screen at about 10 K) and an insulator spacer thermal impedance of 100 K.m2/W at 10-4 Pa we can calculate the variation of heat inleak at 1.9 K as a function of residual gas pressure and compare it to the performance of the reference solution. The change of thermal resistance between soft screen and cold mass with insulation vacuum has been evaluated. Figure 5 shows that the thermalized version has lower residual heat flux both for nominal and degraded vacuum. At 104 Pa the residual heat flux to the cold mass as low as 30 mW/m2 can in principle be achieved. An actively cooled screen will be mounted in a full-scale thermal model [7,8] and heat inleak measurements will be carried out in order to confirm sample measurements and predicted performance.
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|
i
i i
...............
|
l
,
,
1E-02
|
l_
i
9
1E-01
Residual He pressure [Pa]
Figure 5: Comparison of thermal performance between floating and thermalized systems.
REFERENCES 1 2
Evans, L.R., The Large Hadron Collider project, presented at this conference. Lebrun, Ph., Superfluid helium cryogenics for the Large Hadron Collider project at CERN, Cryogenics (1994) 34 ICEC Supplement 1-8 3 The LHC Study Group, The Large Hadron Collider, Conceptual Design, Report CERN/AC/95-05 (LHC) (1995) Lebrun, Ph., Mazzone, L., Sergo, V. and Vullierme, B., Investigation and qualification of thermal insulation systems between 80 K and 4.2 K, Cryogenics (1992) 32 ICMC Supplement 42-47 5 Benda, V., Lebrun, Ph., Mazzone, L., Sergo, V. and Vullierme, B., Qualification of multilayer insulation systems between 80 K and 4.2 K, Proc. Kryogenika'94 Usti nad Labem (1994) 107-110 6 Brunet, J.C., Kerby, J., Lebrun, Ph., Rohmig, P., Szeless, B. and Williams, L., Design of LHC prototype dipole cryostats, Cryogenics (1992) 32 ICEC Supplement 191-194 Dufay, L., Ferlin, G., Lebrun, Ph., Riddone, G., Rieubland, J.M., Rijllart, A., Szeless, B. and Williams, L., A full-scale thermal model of a prototype dipole cryomagnet for the CERN LHC project, Cryogenics (1994) 34 ICEC Supplement 693-696 Benda, V., Dufay, L., Ferlin, G., Lebrun, Ph., Rieubland, J.M., Riddone, G., Szeless, B., Tavian, L. and Williams, L., Measurement and analysis of thermal performance of LHC prototype dipole cryostats, presented at CEC'95 Columbus (1995) Riddone, G., Theoretical modelling and experimental investigation of the thermal performance of LHC lattice cryostats, Doctoral thesis, Politecnico di Torino, Italy (1996) 10 Pe6n, G., Thermo-mechanical study and optimisation of vapour shielded cryostats, Design of components of the half-cell LHC superfluid helium cryostat, Doctoral thesis (in preparation), University of Zaragoza, Spain 11 Danielsson, H., Lebrun, Ph. and Rieubland, J.M., Precision heat inleak measurements on cryogenic components at 80 K, 4.2 K and 1.8 K, Cryogenics (1992) 32 ICEC Supplement 215-218
Cryogenic engineering
Cooling technique
This Page Intentionally Left Blank
A LHE ECONOMISER AT 1.8 K S. BUHLER, Institut de Physique Nucl6aire, F - 91406 ORSAY c6dex, France Cryogenic tests with superconducting cavities at 1.8 K produce a substantial amount of cold He vapour whose enthalpy may be recovered efficiently with an economiser cycle. The working principle, limiting conditions and theoretical performance of a simple economiser operating with cold vapour at 4.2 K are presented. A new economiser with a dual loop for cold vapour produced at 4.2 and 1.8 K level is described. Its energetics is compared with those of other refrigeration modes. First measurements of the performance are reported.
INTRODUCTION Refrigeration at LHe temperatures produces generally a substantial amount of cold He vapour which may be either usefully employed (e.g. for radiation shielding of current lead cooling) or whose enthalpy may be recovered in the LP (low pressure) return line of an autonomous refrigerator. In situations, however, where no refrigerator is available and the LHe boil-off dominates, the use of an economiser looks attractive [ 1, 2 ]. In particular R & D on superconducting cavities at 1.8 K [3] often requires a wide range of refrigeration power which is generally supplied from a crude vaporisation of LHe. BASICS OF AN ECONOMISER AT 4.2 K The principle of a simple LHe economiser at 4.2 K is shown in Fig. 1. A given heat load ~ on a LHe bath produces a boil-off mass flow rh,~p = Cl/1, (1~ = latent heat of vaporisation). The liquid level is maintained, not only with a continuous LHe supply rh~:ofrom an external storage dewar, but also with the liquid fraction (1 - X) of a expanded high pressure flog darn, previously cooled down in a heat exchanger against the escaping low pressure vapour flow rhLp. From a mass and an energy balance on the reservoir we find
rh rhvap = - - = rhne (1 - X ) + &:q
(1)
rh i_iP
t
Introducing [3 = rhLe/rhHe we define a gain factor f f=
~p rh,iq
=
1- X 13-1
i
i-1
(2)
and a specific high pressure flow rate Y=mne = f-1 rhea, / ( 1 - X)
1~ LP
EXCH.
Trr
rialiq
haurx
et
(3)
Figure 1 9Principle of a simple economiser
The liquid fraction (1 - X) depends on state (T, p) of the high pressure flow at the J.T. valve as shown in Fig. 2, whereas Trr is essentially determined by the previous LP/HP heat exchange. 449
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ICEC16/ICMC Proceedings LIQUID
FRACTION 1-X
~
1.0
0.8
VAPOUR FRACTION X mLp(Va p)
=•
mLp(t.iq)
=
--
l-X
.
0
.
-
0.2
--
0.4
( 1 bar )
Figure 2 9Liquid or vapour fraction after isenthalpic expansion of He from high pressure HP to 1 bar.
0.6
0.4
0.6
0.2
0.8
0
10
I
20
1
30
v
-
-
-
-
HP ( bar )
40
For the ideal heat exchanger of Fig. 3 we notice a strong influence of [3 on the temperature profile. The minimum temperature difference ATrm. = 0 is normally located in the heat exchanger region of some 7 to 15 K, but exceptionally can also be found at its cold end (Fig. 3b).
ip t~l~ - " ~-
H
"
I-I
At, .gsa*c
aT..~.SX::
Ale
H
'HI
L P - ~
H
baT mm
4.2 [
14
[ TIT. 6K ]
TIK)
4.2 I
300
[TjT=4.2K]
Figure 3a" Temperature proffie (perfect heat exchange) for [3 < 1,51
=4 r
7.5
I00 T(K)
Figure 3b" Temperature profile (perfect heat exchange) for 13= 1,51
Assuming a perfect heat exchange we can vary Trr for any given pressure and calculate the corresponding gain factor f and its specific HP flow rate y. Table I gives an numerical example for HP/LP = 4/1 bar. When we increase Trr to its limiting value where no liquid is produced anymore the gain f improves steadily, however at the expense of a strongly increasing HP flow rate. Table 1 9Characteristics of an economiser HP/LP = 4/1 bar
(K)
fl = thu'
AT~n
at T(ex)
(K)
mHp
Liquid fraction
(1 - x )
Vapour fraction X
Gain factor f
HP flow rate
m Hp mliq
_P_
"
7'- m~
m,ap
4.2
1.515
7.5
0.95
0.05
2.84
1.94
0.68
4.5
1.470
7.5
0.90
0.10
2.91
2.12
0.73
5.0
1.351
8
0.77
0.23
3.19
2.84
0.89
5.5
1.220
9
0.62
0.38
3.82
4.55
1.19
6.0
1.053
14
0.30
0.70
6.66
18.9
2.84
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ECONOMISER AT 1.8 K Figure 4 shows the simplified flow scheme of this process. Refrigeration is provided with LHe fed from an external, storage into the reservoir R1. At 4.2 K some cooling power Q4.2 may already be required, but the bulk flow is subcooled in E1 in order to reduce the flash losses during the following J.T. expansion to the very low pressure (VLP) of 16 mbar. The main heat load Q~.8 produces an important VLP vapour flow that warms up in E1 to approximately 4 K. The economiser loop is now dual and consists in fact of an exchanger line in three sections : 9 the lowest part from 4 to 80 K with two separate and thermally very efficient heat exchangers E20 and E21 9 the warmest part with a rather crude heat exchanger E4, especially devised for a small pressure drop at VLP 9 a LN pre-cooling bath E3 tbr the incoming HP flow in order to compensate the possible imperfections of FA and to guarantee stable conditions at the warm ends of E20/21. Control. A variable heat load 04.2 or t)x.8 produces cold LP or VLP vapour, its cold enthalpy is permanently absorbed from a matching HP flow through E20 or E 21, adjusted on both lines with a temperature control valve TCV which maintains a preset temperature at each J.T. valve.
I V
I"*'"I 1,": "" I 1 bar
bar
E4 I
t
110 m b a r ~
,t.2K RI
LCV
E1
d4 Q1.8
Figure 4 9Flow scheme of a dual economiser at 1.8 and 4.2 K
Performance The lower temperature of the VLP retum flow allows in principle a full liquefaction of the HP flow in E21 and even some sub-cooling of the produced LHe (Fig. 2). In practice however one may be more interested in bigger output of LHe by setting Trr at a higher value. ENERGETICS Let us consider the specific power consumption W/Q of two conventional refrigeration processes and compare it to those of a simple or dual economiser cycle. A) For an ideal Camot refrigerator, operating between a low temperature T and ambient temperature T o the specific power consumption W A/ Q is W A/ Q = (T O- T) / T B) A simple vaporisation of a cryoliquid with no use of its cold vapour requires an energy investment corresponding to the ideal work of liquefaction, therefore W B / Q = w~iq / lv C)
Refrigeration produced with an economizer requires the total energy W~ W C = Wliq (LHe). Mli q + Win,. Map + Wliq (LN). MLN
(4)
with its components Wliq (LHe) specific work of liquefaction for the extemal LHe supply ; w ~ specific work for isothermal compression LP ~ HP ; W~q (LN) specific work of liquefaction for an optional LN pre-cooling and its corresponding mass flows. For an ideal economiser without any LN pre-cooling, the above equation reduces to 9
Q _
w~iq(LHe ) + wn~,. ~ - ~
(5)
Numerical values of W~ / Q for some selected configurations of a simple or a dual economiser and its comparison with a Camot refrigeration (W A/ Q) or LHe vaporisation (W B/ Q) are given in Table 2. For the
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economiser we find a specific power consumption which is situated generally between those of our two reference processes A and B, except if we approach the limiting conditions of a high Trr or for a very crude design where no counter-current heat exchanger from 300 to 80 K is employed. Table 2 9
Specific power requirements W/Q for refrigeration at 4.2 K and 1.8 K (T o = 300 K)
MODE
DETAIL
CARNOT REFRIGERATOR
WM/Q = ( T o - T ) / T
LHE V,N:~ISATICN
WB/Q = w l t J l v
PRESSURES VI.P/LP/HP (bar)
TjT
(K)
SPECIFIC ~ ~TION W/Q FOR REFRI(IB:tATION AT T = 4.2 K T = 1.8 K 70.4
1.0 / 0.016
-
165.7
326,3
449.3 .
SIMPLE EO3M3MISER
ideal heat exchange
-1114
4.2
142,4 138.7
ideal heat exchange
-11/,4
5.0
ideal heat exchange
-/114
6.0
165.0
LN pre-cooI ing for HP from 300 4 80 K
-11/4
5.0
327.8
DUAL E O : : ~ I S E R
285.7
~(4,2) : o
o.o16 / I / 4
4.o 1 4.0
(~(4.2) = 0
0.016 / I / 4
5,o / 5.o
276.6
o.o16 1 1 1 4
6,0 / 6.0
293.1
(~(4.2) = 0
EXPERIMENTAL DEVICE AND RESULTS For preliminary tests of a cryostat with a superconducting cavity we have built an ancillary cold box [3] which was equipped with an optional economiser, providing the opportunity for some R & D with such a device. Measurement of the gain factor The overall gain factor has been measured in an easy and convincing way as follows. The continuous LHe feed in R1 (Fig. 2) is temporarily changed to an intermittent supply mode between two preset minimum/maximum levels. For a given but stationary heat load the frequency of these intermittent transfers will distinctly change if the economiser is switched on or off thus enabling one to quantify its effect. At the time of writing we have had only a few occasions to run the economiser reliably, but we have already measured an overall gain factor of 4 with Trr = 6 K, HP = 5 to 10 bar and (~.8 = 5 W. CONCLUSIONS An economiser is a relatively simple, efficient and versatile device capable of increasing substantially (a factor of four was already achieved) the autonomy of a cryosystem refrigerated with LHe supplied from an external storage dewar for an important heat load at 4 K or lower. ACKNOWLEDGMENTS The present paper is the result of a common effort of the cryogenic group SBT/IPN which succeeded in inserting the officially marginal experience of the economiser into an already tight test schedule. In particular the author is indebted to R. Chevrollier, D. Grolet and R. Martret for assembly, B. Arapoglou for process control and J.P. Thermeau, P. Brunot for valuable help in discussion and calculations. REFERENCES (1) Avenel O., DerNigohossian G., Roubeau P., A liquid helium saver, Proceed. ICEC6, Grenoble (1976), (2) Buhler S., Subcooled cryogenic targets Proceed. ICEC5, Kyoto (1974) (3) Buhler S., Blache P., Chevrollier R., Junquera T., Colombel N., Panvier R., Gastebois J., Status report of the T1T capture cavity cryostat, Proceed. CEC, Columbus OH (1995).
LOW NOISE GAS FLOW CRYOSYSTEM FOR COOLING HIGH-T~ SQUID
J. TROELL and C. HEIDEN Institute of Applied Physics, Heinrich-Buff-Ring 16, University of Giessen, D-353 92 Giessen, Germany
Conventional coolers which cause disturbing signals must be separated from SQUIDs to achieve the maximum sensitivity of the sensors. For this purpose we have developed a gas cycle system for the thermal contact between cooler and sensors. First a rigid system of stainless steel tubes was used and later replaced by a flexible system of PTFE tubes. This offers a decrease of vibrations and an easy scan of three dimensional fields. The total power losses (2 W) allow the use of a miniature cryocooler and consequently the development of a portable system YBCO-HT~-SQUIDs were successfully operated in an tmshielded environment.
1. THE GAS CYCLE SYSTEM The cycle system is shown in Figure 1. The gas, Helium or Neon, is driven by an oilfree diaphragm pump (DP) and passes a flow controller (FC). Then, the gas is pre-cooled in a cotmterflow heat exchanger (CFX) and in the ideal case adopts the temperature of the cold back streaming gas. The counterflow heat exchanger consists of two coaxial stainless steel tubes (4 m long, 0.2 mm thick, diameters: 5 mm and 2.5 mm) bent to a helix. The gas passes a second heat exchanger (HX) which provides thermal contact to the cooler. This heat exchanger consists of a copper tube (length: 1 m) bent to a helix and soldered to a copper cylinder which is in thermal contact to the first stage (C1) of a two stage Cfifford-McMahon cooler (Typ Leybold RGD 210).
T3 ....
SH
TS TS' ..........
DP
vc
I
I
~
m
CF'X 1TI~C2 ,,I............
72
Cl
T4-'".... 74-"-
T1 ....T5: resistance thermometers
Figure 1 Gas cycle system After that, the cycle gas flows through the transfer system (TS) of 2 m length to the sensorhead and again back to the cotmterflow heat exchanger. 453
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In the first (rigid) version, the transfer system consists of four coaxial stainless steel tubes. The inner tube leads the gas to the sensor head (SH). The second and third are cooled by the backstreaming gas and are used as radiation shields for the cold gas in the inner tube. The outer tube is the vacuum enclosure.
a)
b)
gas flow to the sensor head (Q: 3.5 mm) backstreaming gas v a c u u m enclosure ( ~ : 4.5 and 7 mm)
gas flow to the sensor head P T F E hoses (~" 3.2 m m / /
back streaming gas
2 mm)
superinsulation
corrug
stainless steel tube ( Q 16 mm)
Figure 2 Cross-section of the transfer system; a) rigid version, b) flexibel version. To obtain a better mobility of the sensor head, the rind transfer system was replaced by a flexible one. Two parallel PTFE hoses lead the gas to the sensor head and back. They are wrapped with 12 layers superinsulation as a protection against radiation. Both hoses are placed in a corrugated stainless steel tube which serves as vacuum enclosure. In the flexible system Neon is used as cycle gas, because the PTFE hoses showed a high permeation of Helium. The flexible transfer system with a length of 2 m can be bent in a radius smaller than 25 cm. A straight line scan of 1.6 m length can be made with a distance of 1.8 m between the SQUIDs and the cooler. The sensor head was made of epoxy resin. Metallic and magnetic materials were not used near the sensor head to avoid disturbing eddy currents and distortions of the magnetic field. Two SQUIDs in a distance of 45 mm can be mounted to the lids of the cylindrical sensor head which allows gradiometer measurements.
2. EXPERIMENTAL RESULTS The only difference in the thermal performance of the rigid and the flexible system is in the minimum temperature of the sensor head (55 K rigid, 60 K flexible). Therefore, we only present the results of the flexible system. The cooling down time is about 2 hours (Figure 3). The temperatures of the system components were measured with Platinum resistors (Pt 100). Temperature gradients and losses in the cycle depend on the mass flow rate. From measuring the temperature differences between different locations of the cycle (AT) the power loss P~, depending on the mass flow rate rh and on the specific heat Cp of the cycle gas, has been evaluated by the following equation Pl = thAT Cp. Figure 4 shows the temperatures of the system components and Figure 5 the calculated losses PI. It is evident that the losses in the cotmterflow heat exchanger increase with a higher mass flow. This is caused by its efficiency of about 96 %. The losses in the remaining part of the system are constant (= 1.6 W). With a mass flow of 44 mg/s (Neon) the total losses can be reduced to 2 W, with a temperature of 73 K at the sensor head. The only known closed cycle gas flow system similar to ours [1] is a rigid system which requires more than 4 W cooling power.
ICEC16/ICMC Proceedings
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A High-T~ SQUID gradiometer with YBaCuO microwave SQUIDs was successfully operated in flux locked loop in an unshielded environment. The flux white noise is about 3 o10.5 d~o/~/Hz (Figure 6).
300
behind the cooler (T2)
250
v
sor head (T3,T4) -behind the transfer system (T5)
200-
,.~_--before the cooler (T1)
150E
.
100 -
50-
I
,
,
I
30
I
,
t
9O
60
150
120
time (minutes) Figure 3 Cooling down behaviour
90-
' V ~
I
80-
(I) Q.
E
'
I
'
I
'
I
'
before the cooler (T1)
er dem Transfersystem (T5)
75-
~==
I
"~'--~~
85-
L_
'
7065 sensor head (T3)
6095 5
~
,
-
504540
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40
behind the cooler (T2) []
-m
[]
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!
50
'
i
60
'
i
70
'
I
80
'
I
90
mass flow [mg/s] Figure 4 Temperatures of the system components as fimction of mass flow
"
100
456
ICEC16/ICMC Proceedings .0
,,.
'l
1'
j
.....
i
'
'
9
'
' 9
9
total power loss 2.5
_•
A~
2.0
(1)
o
1.5
,.,,,,
sensor head and transfer s y s t e r ~ . . . . . . . . . . ~. " .
--= .
i ~
.
.
.
.
L_
(1) 0
counterfl
1.0
0.5 L
40
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I
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|
50
,
,
,
,
60
I
~
..
70
I_
,
I
80
,
i
,,
90
100
mass flow rate [rag/s]
Figure 5 Power losses of the system components as fimction of mass flow
10 "1
10
:::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::::
.2
,
"
,
,
"
,
,
N
-1" v
!!!!_
~ ._ or X
= ,
!!! !!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!!,
04
_
0
,,,
20
,
~
40
60
80
,
100
frequency ( Hz )
Figure 6 Flux noise spectrum of an YBaCuO SQUID gradiometer operated with the rigid gas cycle system
3. CONCLUSION A low noise cooling system for high-Tc SQUID was developed and tested successfully with YBCO SQUID. The total heat load could be reduzed to 2 W. These rather low losses indicate that it should be possible to use a miniature cryocooler instead of the GM cooler for the development of a portable system. 4. REFERENCES van den Bosch, P.J., Holland, H. J., ter Brake, H. J. M. and Rogalla, H., Closed-cycle gas flow system for cooling high Tc d. c. SQUID magnetometers, Cry_ogenics (1995) 35 109-116
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Compact Dilution Refrigerator Shigeru Yoshida*, Satomi Moil*, Takahiro Umeno*, Yasuharu Kamioka**, Michio Watanabe*** and Yoichi Ootuka*** *Technical Research Laboratory, Taiyo Toyo Sanso Co., Ltd., Kawasaki, 210, Japan **BMEP & Cryogenics Division, Taiyo Toyo Sanso Co., Ltd., Tokyo 104, Japan ***Cryogenic Center, University of Tokyo, Tokyo 113, Japan A compact dilution refrigerator developed by Ootuka et al. is modified and improve the performances in this study. The modification is made in some parts, which include a condenser and a 3He evacuation tube. Improvements in ultimate temperatures and in handling-easiness have been achieved. The ultimate temperatures is 80 mK in continuous mode, and 36 mK in one-shot mode. Another compact dilution refrigerator having a FRP mixing chamber has also been developed. Its ultimate temperature is 100 mK in continuous mode, and 65 mK in one-shot mode.
INTRODUCTION A dilution refrigerator is a powerful instrument for low temperature experiments, because it produces mK temperature continuously [1][2][3]. Its handling and operation are, however, so complicated that it is hard to complete an experiment in a day. Especially, cyogenic vacuum seals are cumbersome and experimenters are bothered with disassembling/assembling the refrigerators before they start every experiment. Against this background, Ootuka et al. developed "one-day dilution refrigerator", which doesn't have any cryogenic seals.[4] In the development, they traded off refrigeration performance for handlingeasiness. So, while the refrigerator is very easy to handle, its ultimate temperature is as low as 90 mK in oneshot mode and 160 mK in continuous mode, which is not low enough when one compares with conventional dilution refrigerators. As a second stage of development, we intended to improve its performance. We modified some parts of the refrigerator and succeeded in lowering its ultimate temperature. STRUCTURE OF COMPACT DILUTION REFRIGERATOR Basic Structure Figure 1 is a photograph of a compact dilution refrigerator system. It consists of three main parts, e.g. a refrigerator unit, a cryostat and a gas-handling unit. In figure 1, the refrigerator unit is installed in the cryostat and is connected to the gas-handling unit by means of flexible stainless steel tubes. Figure 2 shows a schematic flow diagram of the refrigerator. A conventional dilution refrigerator consists of several parts, that is, a 3He condenser, a still, heat exchangers and a mixing chamber, and they are connected by thin tubes in an adiabatic vacuum chamber. By contrast, the present dilution refrigerator consists of an insert and a jacket. The insert consists of a plunger, a flow impedance, a capillary tube, a condenser, 3He inlet tube and 3He exhaust tube, and is put in the stainless steel jacket. The lower part of the jacket is thermally insulated from surroundings by an adiabatic vacuum. The plunger inserted into the jacket divides it into two parts, e.g. the upper part as a still and the lower part as a mixing chamber. Heat exchanger is formed in a space between the plunger wall and the stainless steel jacket. Samples, lead wires, thermometers and valves are also attached to the insert. The vacuum seal between the insert and the jacket is made by rubber O-ring at the room temperature top end of the jacket. So, one can change samples without breaking cryogenic seals. Circulating 3He gas comes from the top flange into the refrigerator unit. It is liquefied at the condenser and is lead to the mixing chamber through the capillary heat exchanger. After it cools in the mixing chamber, it goes up in the small space between the plunger and the jacket wall, and reached to the still, where it is pumped by 3He circulation pump. Modification of 3He condenser In order to make a structure simple, an independent 1K pot is not installed but main liquid 4He bath is pumped. So, it is necessary to reduce the thermal resistance between the 3He condenser and the 1.5K bath. In the previous model, a part of the stainless steel jacket wall had been replaced by tapped copper tube, into
458
ICEC16/ICMC Proceedings .....~:::::~::.::::::..:;;..:~. :::..:.: :::: ::. ::~,:::~..~:~: ::~:..::.:..:.::: ..:...:......~:
:~:: ::...::~ ...............
:....
.
~:... ..............
:
9 .~. :..~:. ::~......................
9
::~ ...............
:...
.:.
:.
:............
.......... :::::::::::::::::::............. ~ . _ . : ~
.........
. ..........
. ... .......
.
..~::..~,
Figure 1 Photograph of the compact dilution refrigerator
I.=-
1 -,~.:_..:~.:,~- [ ~ 1
"
U
3He Circulati~ Pump
N Still N ~N
~ ~ ~
I ~
~
Liquid He Bath Stainless Steel Jacket
~ 3He Condenser ~ Heat Exchanger Plunger Mixing Chamber
4He Pump LNT 9Liquid Nitrogen Trap MFM 9Mass Flow Meter OT" Oil mist Trap P1-P5"Pressure gauge
Figure 2 Schematic flow diagram of the dilution refrigerator which the 3He condenser had been screwed. In the present model, we use a commercial large-capacity electrical connector. A female connector is a copper tube in which silver-coated beryllium-copper springs is installed and a male one is a copper block which we use as the condenser. When it is installed, the condenser is held firmly by the springs. We find the thermal contact between the condenser and the bath is sufficient. The adoption of the spring contact makes the insert installation easier. Schematics of the modification are shown in figure 3. Prevention from 4He Film Flow As 4He in the still is a superfluid, it creeps up the jacket's inner wall as a film and evaporates from its whole surface. This results in decrement of 3He partial pressure in the circulating gas and degrades the performance
ICEC16/ICMC Proceedings
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of the refrigerator. In order to prevent this contamination, a FRP (Fiber Reinforced Plastics) tube is put inside the jacket as an exhaust tube for 3He. As shown in figure 3, it is connected to the condenser block, and a small tube and a lid made of stainless steel are connected below the condenser so that 4He gas evaporated at higher position should not be evacuated from the refrigerator. FRP mixing chamber Non-magnetic or non-metallic refrigerator is frequently required in several research fields, such as Josephson junction circuit and high magnetic field application. As an optional product, we have developed a mixing chamber made of FRP. Figure 4 shows a photograph of the tail parts of the insert and the jacket. An extended plunger/mixing chamber made of FRP and Teflon tube is attached to the stainless-steel plunger. The extended length is 150 mm. The jacket coveting the plastic plunger and the mixing chamber is also made of FRP.
OPERATION AND PERFORMANCE Ruthenium oxide resistance thermometers are used to measure the temperature of the mixing chamber using an ac bridge circuit. Heaters are put in the still to control thecirculation rate and in the mixing chamber to put a heat load. The following three items are investigated during operation: (1) the ultimate temperature at continuous mode operation, (2) the ultimate temperature at one-shot mode operation, and (3) refrigeration power at continuous mode operation. In the one-shot mode, we evacuate 3He from the refrigerator but stop circulating it to obtain lower temperatures transiently. The ultimate temperatures at each mode are shown in table 1. Experimental results of refrigeration power and theoretical values [3] at 66 gmol/sec of 3He are plotted in figure 5. Comparing with the previous results, we find the performance is improved much. The refrigeration power deviates from the theoretical value below about 0.2K, which means insufficient heat exchange at this temperature range. We also measured the ultimate temperature of the FRP mixing chamber. It is 100 mK in continuous mode and 65 mK in oneshot mode.
3He Gas Outlet
------4
3He Inlet Tube
tag
Stainless Steel Jacket F R P Tube 3He Condens Copper ~ Spri______nngContact Thread Contact Lid Still
Heat E x c h a n g e r Plunger Mixing Chanber
Thermal V a c u u m Insulation a. before modification
b. after modification
Figure 3 S c h e m a of modification
Figure 4 Photograph of a plastic refrigerator. The jacket (left) and the tails of the insert (fight) are shown.
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Table 1 Refrigeration Power . . . . . . .
Operation Mode
3He Circulation Rate l.t mol/sec
Temperature mK
Refrigeration Power ~W
One-shot mode
36
Continuous mode
80
46
0
Continuous mode
120
51
26.6
Continuous mode
158
60
73.5
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CONCLUSION A compact dilution refrigerator has been developed. It has no cryogenic seal so that the setting up of the refrigerator is easy and quick task, and simple experiment can be completed in a day. The construction of the refrigerator is also easy because of its simple structure. The ultimate temperatures are 36 mK at one-shot mode and 80 mK at continuous mode. The refrigeration power at 120 mK is 26.6 I.tW. A dilution refrigerator with a FRP mixing chamber has been also developed. It shows a similar performance, that is, the ultimate temperatures are 65 mK in one-shot mode and 100 mK in continuous mode. The compact dilution refrigerator is marketed as "l.t-Dilution TS-3H100" by Taiyo Toyo Sanso Co., Ltd.. REFERENCES 1 2 3 4
Kobayashi, S. and Ootuka, Y., Cryogenic Techniques ver.2 (in Japanese), University of Tokyo Press, Tokyo (1995) Betts, D.S. An Introduction to Millikelvin Technology, Cambridge University Press, Cambridg (1989) Lounasmaa, O.V., Experimental Principles and Methods Below 1K, Academic Press Inc., London (1974) Ootuka, Y. et al., One-day dilution refrigerator Cryogenics (1993) 33 923-925
Experimental Study of the Dilution Refrigerator without 1K Pot
Minoru Maeda, Toshinobu Shigematsu, Zhongmin Li, Toyoichiro Shigi, Yoshiko Fujii, Minoru Yamaguchi and Masaki Nakamura Dept. of Applied Phys., Okayama Univ. of Science, Ridai-cho 1-1, Okayama 700, Japan
A simple dilution refrigerator without 1K pot was constructed and its basic characteristics were investigated extensively. Several features peculiar to this type of the dilution machine have become clear for the first time and they are discussed in detail. The highest 3He inlet pressure is less than 1.4 atm, which enables one to circulate 3He without a compressor. The lowest temperature of the mixing chamber is 19.8 mK and highest temperature is 570 mK. The 3He circulation rate is limited in the range of 40 ~ 70 # mol/sec. Over these rate, the mixing chamber temperature becomes unstable.
INTRODUCTION The 3He-4He dilution refrigerator without 1K pot liquefies the circulating 3He by Joule-Thomson expansion and by making use of refrigeration capacity of the still and the 3He gas evaporated from it. [ 1,2 ] This scheme makes the installation simpler, economizes liquid 4He consumption, and especially makes the operation much easier. In order to investigate the basic characteristics B (A) 3He pumping tube of this type of refrigerator in detail, a simple (B) Vacuum jacket dilution machine has been constructed, which o pumping tube has only a tube-in-tube heat exchanger .o :0 (C) 4He bath cooler between the still and the mixing chamber. o
~ C o
(D) J - T heat exchmager (E) Aluminum alloy screw (F) Indium seal (G) J-T impedance
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Construction of tile J-T heat exchanger.
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mm, ID 0.3 mm, length 4.5 m) has a shape of sheared double solenoid (height 8 cm, see Figure 2) in the outer tube (ID 15.2 mm). This new type of the heat exchanger was verified to have quite a low flow resistance on the low pressure side and a reasonable heat exchange efficiency of about 90 % from the calculation of enthalpy balance. The J-T impedance (Z=4 • 1012 cm 3) is a Cu/Ni tube (OD 0.4 mm, ID 0.1 mm, length 40 cm) flattened by a roller. The 2nd impedance (Z=5 • 10 ~l cm .-3) is a high temperature part of the inner tube of the tube-in-tube heat exchanger, which contains a fine manganin wire. The 4He film burning heater, which serves also as the still heater, is designed not to warm up the pumped 3He gas. The temperature of the mixing chamber is measured with the Pd-Fe (Fe 35 at. ppm) susceptibility thermometer. The Pd-Fe rod (Dia. 5 mm, length 10 ram) having a metallic thermal conductivity is screwed on to the Cu wall of the mixing chamber in which Cu powder is sintered. This thermometer was calibrated by a ruthenium oxide resistance thermometer graduated beforehand and a nuclear orientation thermometer. Susceptibility of this Pd-Fe sensor was represented by the Curie-Weiss law ( 0 w=9 • 10 .4 K) down to 12 mK with an error less than + 3 %. (see Figure 3) (atm) (mK) Figure 4 shows the gas handling system. The 3He ~-~ l..... i 1 i l i gas is circulated by a diffusion booster pump (D.B.P) (300 g/sec, critical back pressure 0.8 Tort*), a mechanical booster pump (M.B.P) (1500 g/min) and a rotary pump (R.P) (200 Umin) in 1000 series. The rotary pump worked safely at the '~ ..-o--. - . - - 2 highest back pressure of 2 atm** with some reinforcement. OPERATING CHARACTERISTICS
:
0.5
At the beginning of the operation, we tried to liquefy the working gas (~He 7.4 g and 4He 19.6 g) through the J-T impedance. As liquefaction proceeded, the He gas in the storage tank was continuously supplemented to the circuit between 91 Torr=l.333 • 102 Pa,
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ICEC16/ICMC Proceedings 9
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463
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the D.B.P and the M.B.P. Because of the low critical back pressure of the D.B.P and back pressure regulation of the R.P, it took very long time, more than 4 hr, to complete liquefaction. Therefore, we decided to reduce the temperature of the 4He bath to about 2 K by evacuation during the initial liquefaction and to liquefy the working gas from the low pressure side. By this procedure, the initial liquefaction time was reduced to about 40 min (see Figure 5) and also possibility of J-T impedance blockage was greatly reduced. Throughout the experimental run, amount of the working gas was kept constant. The obtained characteristics are given in Figures 6 and 7. In Figure 6, heater power was applied only to the still. Without heat input to the still, both the mixing chamber temperature and the 3He inlet pressure were minimum, 19.8 mK and 0.57 atm respectively. When the still heater power was increased to 1.5 mW, the mixing chamber temperature rose abruptly up to 60.2 mK, although the 3He inlet pressure stayed at 1.35 atm. In this case, enthalpy of the incoming 3He gas at 4.2 K was not sufficiently low to cover the increased heat input. Therefore, the incoming 3He could not completely be liquefied in the still cooler and made the temperature of the mixing chamber rise so high. This effect limits the 3He circulation rate from 40 only to about 70 ~ mol/sec. The curve of the 4He content in the circulating gas shows that the film burning is completed at the still heater power 1 mW. In Figure 7, the cooling power was measured by changing heat input to the mixing chamber with a constant still heater power 1 roW. At about 200 lz W to the mixing chamber, the 3He inlet pressure and the 3He circulation rate began to increase with increasing heat input. At 600 ~z W, the 3He inlet pressure, 3He circulation rate and the mixing chamber temperature reached maximum value, 1.38 atm, 79 ~z moI/sec and 570 mK respectively. Over this heat input, the mixing chamber temperature became unstable. As the mixing chamber temperature approached to 600 mK, large amount of 3He moved from the 3He-rich side to the 3He-dilute side and also to the high pressure part of the 3He circulation circuit. As a result, the phase boundary moved up out of the mixing chamber, which made the mixing chamber temperature unstable. CONCLUSION The basic characteristics of the dilution refiigerator without 1K pot was investigated extensively and following features have been clarified for the first time. 1) The new type of the J-T heat exchanger devised by us works very well with quite a low flow impedance on the low pressure side and a reasonable heat exchange efficiency, ~ 90%. 2) By choosing a low J-T impedance, the highest 3He inlet pressure becomes less than 1.4 atm. This enables one to dispense with a compressor for ~He circulation, but requires some device for initial liquefaction of the working gas. 3) This type of the dilution refrigerator limits the range of the 3I--Ie circulation rate to less than 2 times because of the restricted liquefaction capacity of ~He. 4) The high pressure -~He in the circulation circuit depresses the highest temperature of the mixing chamber owing to move-out of the phase boundary from the mixing chamber. REFERENCES 1 Kraus,J., New condensation stage for a 3t-Ie-4He dilution refrigerator, C wogenics (1977) 17 173-175 2 Uhlig,K., ~He/4He dilution refrigerator without a pumped 4He stage, CIiyogenics (1987) 27 454-457
NUMERICAL SIMULATION OF COUNTERCURRENT HEAT EXCHANGERS IN CRYOGENIC SYSTEMS
M. Kauschke, H. Quack Technische Universit~it Dresden Institut for Energiemaschinen und Maschinenlabor Lehrstuhl for K~ilte- und Kryotechnik 01062 Dresden, Germany
ABSTRACT The knowledge of the non-stationary behaviour of components in a cryogenic system is important for optimum cool-down procedures, design of the control system and stable operation of the overall system. Numerical simulation is a powerful tool to investigate those transient phenomena during start-up. For countercurrent heat exchangers in general and the Joule- Thomson heat exchanger of a cryorefrigerator in particular, the governing partial equations are being assembled.
INTRODUCTION
In the physics community, new ideas occurred for an accelerator which will induce a higher energy onto particles than the ones nowadays in use. One of these ideas is the TESLA project, where superconducting structures are used for the acceleration and the guidance of the beam. For this new machines new cryogenic plants are necessary. For the construction of safe and efficient refrigerators, which meet the requirements of those large accelerator systems, it is necessary to know more about the behaviour of the cryosystem before the actual keyturn. One way of learning about these systems are test facilities. But these test facilities, which for the TESLA project are located at DESY/ Hamburg, can only give indications of the problems which may occur during the scale up. Therefore a simulation of the a cryogenic system based on the data given by TESLA[ 1] is a necessary task. WHAT DO WE NEED? For the numerical simulation of a refrigerator system it is necessary 9 to evaluate suitable software, 9 to find suitable mathematical models of the governing physical laws 9 and to have a full description of the requirements of the helium refrigerator. The simulation software has to fulfil different criteria, one of them being that it has to cope with continuous processes. The software tool has to solve partial (PDE) as well as ordinary differential equations (OPE) together with algebraic equations. That excludes a wide range of commercial software. We were looking for an open source code to have a chance of seeing what is really being calculated and so we are still able to influence the way it is done. And last but not least the software should be affordable. The decision was made in favour of a simulation tool called DSS/2 [2]. This software had been used for the simulation of some aspects of cryogenic system of the SSC project [3]. 465
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For the description of the cryosystem the different kinds of mathematical models for the components have to be found. The governing laws for the components are given in form of partial differential equations as the mass balance, energy balance and the momentum balance. A summary of the components and the necessary information about their characteristics is given in table 1. In the third column the magnitude of modifications necessary to accommodate changes in the actual cryosystem is given. Table 1 Information needed for a full description of a helium refrigerator. Component of the helium Characteristic of the component Variation to the model from refrigerator application to application helium equation of state none transport properties overall geometry low heat exchanger channel distribution fin geometry density, specific heat, thermal conductivity heat transfer and pressure drop characteristics flow characteristic high expanders, compressors efficiency influence of speed insulation heat leak middle throughput characteristics middle valves load characteristics temperature, heat load high control logic high controls For the dynamical numerical simulation the priorities for the handling of the properties of helium are different from the ones for an accurate steady state calculation. The calculation of the properties have to be reversible, consistent and quick. Therefore the decision was made to work with explicit equations only using polynoms of a maximum third order. From the NIST equation of state, coefficients for second order polynomial interpolation equations were calculated on the isobars for selected temperature ranges. During the simulation calculation, all intermediate values can be calculated explicitly with third order polynoms. The thermodynamic properties obtained by this interpolation are reasonably in good agreement with the ones obtained by using the full NIST equation. But the computing time is about 10 times faster. Heat exchangers are important components of the refrigerator. The different stages of sophistication in the simulation of heat exchanger are shown in table 2. Table 2 different stages of sophistication in heat stage of passages gas velocity sofistication basic single pass overall average advanced multiple passes, single pass distribution average effects highly velocity gravity effects advanced profile
exchanger simulation heat transfer pressure drop Fixed ~ or NTU neglected
longitudinal conductance neglected
j - f (Re) rl~ - f (j, T)
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simple factor
transverse temperature profiles
including inertia terms (oscillations)
accurate temperature profiles
The behaviour of a Joule-Thomson heat exchanger during cool down is a good example to test calculation procedures. The way of evaluating the mathematical model is shown.
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The heat exchanger is divided in several subelements and the partial differential equations in time and the one space dimensions for the mass and energy are assembled. From the mass balance the changes in time t of the total mass M of the controlled element i,j in respect to variations of the mass flow rate m in the flow direction x is given. The energy of the controlled element is given by the product of the enthalpy h and the mass. Therefore the course of the energy of the element is the difference between the energy of the mass flow mh and the heat transfer through the boundaries of the element. The amount of heat transferred through the boundaries is given by the heat transfer coefficient k the area of heat transfer A and the temperature difference. mass balance:
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The method of the calculations is to transform spatial derivatives into algebraic expressions at every time step x, thus transforming the PDE into a set of ODEs. For the solution a solver for differential algebraic systems of equations is applied. Special attention must be paid to initial and boundary conditions Initial and boundary conditions are as important as the differential equations. An unsuitable choise of initial and boundary conditions will lead to wrong results or unstable calculations. For this example the momentum equation is not considered, as this increase the stabiltiy of the problem. Several authors such as Janssen[4] have shown that this simplification meets the reality so far, that the increase of calculation time caused by the momentum equation is not justified. EXAMPLE In figures 1 to 4 the temperature profiles in a JT- heat exchanger is shown in the last step of cool down. The calculations start after the exchanger was cooled to 20 K everywhere. At the time x=0, high pressure stream enters the heat exchanger on the left hand side at a pressure of 20 bar and a temperature of 15 K. On the right cold end, the flow is throttled down to 1.19 bar and if liquid is formed, it is separated. No additional heat load on the cold end of the heat exchanger is supposed. Therefore the cool down of the heat exchanger is relatively quick. 20 19 18
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For this calculation the influence of the helium properties was investigated. The calculation time for both methods up to the steady state were compared. The same calculations by using the NIST calculation routines spend 3 times more calculation time on the problem. The difference in the simulation result is not considerable. From this calculations we decided to go further in our simulation with the polynomapproximate for the properties of helium to reach a reasonable calculation time CONCLUSION For the numerical simulation of cryogenic plants the careful selection of suitable software is necessary. Special attention has to be paid to the degree of simplification of the system to find the optimum between correlation with the real system and calculation time. ACKNOWLEDGEMENT For the support of this research project we are pleased to thank DESY (Deutsches ElektronenSynchrotron).
REFERENCES [ 1] G.Horlitz, A Study of Pressures, Temperatures and Liquid Levels in the 2 Kelvin Refrigeration Circuit of a 1830m TESLA Subunit under Different Conditions and System Configurations; TESLA 94-17, DESY(1995) [2] C.A. Silebi, W. E. Schiesser, Dynamic Modelling of Transport Process Systems, Academic Press Inc., San Diego (1992) [3] R.Carcagno, W. Schiesser, A. Yticel, Prototype Helium Flow and Heat Transfer Model and Code: SSCL-N-850 (1993) [4] M.J.P. Janssen, Cycling Losses in Domestic Appliances an Experimental and Theoretical Analysis International Journal of Refrigeration, 15 (1992) 3, 152-158.
The Operation Experience with the Dual Cooling System of the POLO Coil (two phase-supercritical He) W. Herz, M. S(iger, A. Ulbricht, F. W~ichner Forschungszentrum Karlsruhe, Institut f(ir Technische Physik, Postfach 3640, D-76021 Karlsruhe, Germany During 4 test runs in 1994/95, the POLO coil was successfully tested in the TOSKA facility at the Forschungszentrum Karlsruhe, Germany. The NbTi coil has a diameter of 3 m and consists of 4 double pancakes, each with 140 m hydraulic length. The conductor is characterized by a thick walled stainless steel jacket and a dual cooling system (two phase and supercritical helium). The cryogenic flow sheet for supplying the primary and secondary cooling channels and the three force flow cooled current leads are described. Operation experience during the test periods is given. The system was easily to handle. A sensitivity to thermoacoustic oscillations was observed, and some possible reasons were encountered. Experiments for using pressure increase as quench detection signal showed that the pressure increase vanishes within the usual pressure fluctuations.
INTRODUCTION The magnetic field system of a tokamak for the magnetic confinement in fusion consists of two types of magnets: toroidal (TF) and poloidal (PF) field coils. The technology of scalable superconducting TF coils was achieved ~ with the Large Coil Task [1]. The PF coil development was started at several laboratories in the mid of the eighties. In FZK/ITP, the technology development started for an outer PF coil suitable to be used in a running t o k a m a k experiment (e.g. Tore Supra). The high heat load during operation (fast ramp up, plasma control) required special techniques to master this for a superconducting coil. The large conductor length and the relatively high heat load exhaust the temperature margin of NbTi so that forced flow cooling by supercritical helium is not suitable. Therefore, a dual cooling system was used. The coil in Fig. 1 (3 m Q) was constructed and tested in the TOSKA facility at FZK Karlsruhe [2]. The description of the cryogenic supply system and the operation experience during testing is subject of this contribution. DESCRIPTION OF THE CRYOGENIC SUPPLY SYSTEM Fig. 2 shows in principal the simplified flow sheet of the whole system. Fig. 3 shows the cooling circuits of the coil. The particularity of the cryogenic system is the dual cooling system by stagnant supercritical helium for the primary circuits (subcables in the annular space) and forced flow two phase helium for the secondary circuit. The current leads are forced flow cooled by supercritical helium. The two phase helium in the central tube removes the heat at constant temperature while the stagnant supercritical helium in the annular space surrounds the cable to guaranty well-defined heat transfer properties for good stabilization against sudden and short heat pulses. Primary circuit (stagnant supercritical helium) The primary circuits of the double pancake DS1, DS2, DS3, DS4 are supplied across a pressure controlled valve PIC810 to the three conductor terminals LEV920, LEV960, LEV980. The outlets are collected at the conductor joints DSV820, DSV840, DSV860 in a 469
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line with the pressure controlled outlet valve PIC832. Both pressure controllers keep the s t a g n a n t He at a constant level above the critical pressure. Secondary circuit (two phase helium) The two phase forced flow in the secondary circuit was circulated by a piston pump. After a pump breakdown the JT-flow of the refrigerator was directly used for helium circulation. The mass flow of each pancake is controlled by the controller FIC820, FIC840, FIC860, FIC880. A m e a s u r e m e n t system [3] to detect the vapor content of each pancake was installed in each outlet (XI827, XI847, XI867, XI887). The optimized mass flow of each pancake was about 2 g/s. The used pump has a maximum mass flow rate of 150 g/s. In order to adjust the pumping output to the system head curve, a bypass control valve was installed. Current leads The three current leads SZ920, SZ960, SZ980 are also supplied by the JT-flow of the refrigerator. The pressure in the leads are kept under supercritical conditions, by use of the pressure controller PIC900. The helium cools the joints connecting the conductor terminal with the cold ends of the current leads and after that, it runs through the heat exchangers of the current leads. Mass flow controllers at the outlet regulate the helium gas flow. In order to remove the heat losses of the transfer lines and to be sure that the two phase helium was subcooled, the heat exchanger in vessel B400 was installed. The vessels B500 and B600 are mounted to store the expelled helium of the coil after a discharge. Safety principals The design pressure of the system was 2.0 MPa. A programmable logic controller (PLC) Simatic $5-150 was used to control the cryogenic system. Three different protection levels are designed against overpressure, first the PLC, after that a hardware controller, and finally safety valves and at least bursting disks at 1.5 MPa. Since the first two pressure levels keep the helium in the system, the third level released the helium into the atmosphere. The response of the safety valves has usually no impact on continuous operation. The bursting of the disks interrupts the operation for several hours. During six months operation the pressure release was mainly handled by the PLC. OPERATION E X P E R I E N C E P r i m a r y circuit The i n t e r m i t t e n t pressure refilling of the primary circuit initiated oscillations in the secondary circuit and also in the current leads. During refilling periods, the helium in the piping network warmed up and transferred heat into the system. Therefore, refilling was only done by the operators with uncharged coil. The automatic pressure control system proofed to be unsuited under the given boundary conditions. Secondary circuit Especially during using the piston pump the circuits were sensible for oscillations. For example, only small changes of the control valve position brought the system into oscillation. In the worst cases, gaseous slugs in the two phase flow occurred. The recovery to stable flow condition lasts about two hours. The system was generally stable when the flow was circulated by the JT-flow without the pump. Therefore a small vapour content only leads to phase separation in the manifold with uncontrolled cooling conditions and incorrect mass flow measurements. It was necessary to keep the two phase helium subcooled at the coil inlet. Current leads [4] The adjustment of the mass flow of the current lead for different current levels was easily to handle. After increasing the mass flow, oscillations occurred sometimes in the leads also. In order to keep the Nb3Sn at the cold end insert always in superconducting state, the mass flow rate was increased by about 15 % above the nominal rate. The current leads are supplied from one pipe and are not separated by check valves. This coupling could be the reason for oscillations.
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Heat balance The necessary cooling power for the coil was 10 W for each pancake and 20 W for each joint between conductor terminal (LEV) and current lead (SZ). Each lead needs 0.5 g/s without current and 0.08 g/(s-kA) with current. The steady state heat losses of the whole system were about 300 W. In standby operation, the equivalent heat losses were 550 W at 4.5 K~ During the 24 kA coil operation, for example, the average heat load was about 900 W. P R E S S U R E INCREASE AFTER Q U E N C H I N G A transition of the coil from the superconducting to the normal state ("quench") has to be considered as a usual system disturbance. The control system should m a s t e r it without loosing helium to the atmosphere or generating heavy disturbances in the cryogenic system. The system disturbances will be low if the normal zone is detected as early as possible. Up to now, the most sensitive detection is the resistive voltage measurement. Other non electrical methods are of a d v a n t a g e for redundancy reasons, especially for large forced flow cooled magnets. Therefore, the dynamics of the pressure increase during a quench was experimentally investigated. The quench was caused by an inductive heater pulse (Fig. 2) at the inlet of DS1 which is about i m a p a r t from the joint DSV820. In case of a closed valve HCV841, the ventline is disconnected. The conductor and the joint box contain nearly incompressible supercritical helium. The pressure was m e a s u r e d with the cold pressure transducer PTR821 directly at the conductor joint box DSV820. According to the test a r r a n g e m e n t , the quench takes place at the beginning of the pancake. The average energy input due to the quench was about 0.93 kJ. When the valve HCV841 is closed (Icoil = 22.5 kA) which is not for usual operation, the pressure rises 1.7 bar with about 2.0 bar/s (Fig. 4). With the valve HCV841 open (normal operation mode), the pressure change due to the normal zone is within the pressure fluctuation because the pipe connection to the w a r m safety valve contains compressible helium gas which leads to a r e m a r k a b l e damping. Considering the response time when the valve HCV841 is closed, the pressure level changes r e m a r k a b l y i s later t h a n the voltage signal. This means, in our experiment, already about 2 m normal zone. It shows clearly when the valve HCV841 is open the pressure change cannot be used for quench detection at all. Spontaneous quenches started every time at the innermost t u r n which is 75 m far from the pressure sensor. Fig. 4 shows such a quench at 24.3 kA with HCV 841 open where also the pressure increase r e m a i n s within the usual fluctuation. CONCLUSION It was shown t h a t such a dual cooling system can be reliably operated. Avoiding the observed oscillations, some improvements should be considered in the cryogenic circuit to prevent higher t e m p e r a t u r e s in pipe sections (e.g. integration of the safety valves). Also a parallel supply of components (e.g. current leads) should be avoided. All pipes routed to w a r m safety valves have an impact to the dynamic system behaviour. Pressure increase caused by a quench was not observed in the time range of one second after the quench initiation. The pressure increase r e m a i n s within n a t u r a l fluctuation. ACKNOWLEDGMENTS The work has been performed within the framework of the Nuclear Fusion Project of FZK and is supported by the EU within the EU Fusion Technology P r o g r a m m e . The authors t h a n k their colleagues involved in the performance of the experiment. REFERENCES .
.
3. 4.
Komarek, P. H o c h s t r o m a n w e n d u n g der Supraleitung, B.G. Teubner S t u t t g a r t 1995, 362. Darweschsad, M. et al., Proc. MT-14, Tampere, J u n e 11-16, 1995, Finnland. Katheder, H. and S(i~er, M., Cryogenics (1991) 31 327-329. Heller, R. et al., IEEE Transactions on Magnetics. (1994) 30 No. 4
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~ - ~ - ~ .... ,,.o~.
If ~ ~,~ r~, .........~.~~ . . . . _..._.....1 [ ~ B25t]
B3gO
Fig. 2: Principle cryogenic flow sheet of the system 3,8
o
-HCV841 c l o s e d ~
.,,,
..i.. i , ~ . ~ ~ . ~
DSI..
- ~!
9
'
~
-8o
,
~
3@6
PTR821...i
~,4
.
.6o
~
,.
~ ' . , / T
3,2
i HCV841 open ! i i~
-
:: ! "
3,0
2,8
-100
. 0,0
Fig. 1" POLO-coil in its supporting structure ready for installation in TOSKA
. 0,5
J 1,0
1,5
:
2,0
-.--j,.-q
~~t ~
~
.......
r .... : - l g ~ ~
.~.
i
4,0
]o ..11 "~ - " [ ~
........ :-l~-:.~0,~
'~'~ET ~,~
51
3,5
TIME [s ]
! i ~ ~-:-~~
~.~).~.0
~_~.
2,6 3,0
4,5
5,0
Fig. 4: Resistive voltage at DS1 and pressure increase at different operation conditions vs. time
~- ~~.................................... ~ [~-~~z ~o
" 2,5
"~[
~,~:% r'
Fig. 3" The cooling circuit of the coil
~
i
j
I'Y'
O9 O9 ILl n" 13
A VORTEX REFRIGERATOR FOR NMR EXPERIMENTS Muneaki Fujii, Keiichiro Nakamura, Xun Xu, Kengo Kawano and Kunihide Okada Department of Physics, Kumamoto University, Kumamoto 860, Japan
A vortex refrigerator especially designed for NMR measurements around I K has been constructed. The main parts of the refrigerator have been made by non-metal substances to reduce eddy current heating due to applied rf pulses. The sufficient cooling power for continuous measurements has been obtained. The mutual friction between superfluid component and normal fluid component in He II is also discussed.
INTRODUCTION In NMR experiments at very low temperatures, the following two points are important. 1)There must be a good thermal contact between a sample and the cryostat. Since thermal contact resistance between two cold bodies increases rapidly as the temperature falls, it becomes increasingly important how to transfer heat from one place to another. 2)There must be little or no eddy current heating by applied rf pulses. In the presences of some metals, eddy current heating occurs when we apply rf pulses in the NMR experiment. Therefore it is advantageous to make cryostat by non-metal substances and to immerse samples directly into cold fluid. In this way good thermal contact between the samples and the fluid is achieved and unwanted heat flux from the outside does not act directly on the samples. It is necessary to extend the temperature range accessible to experimental investigations. Temperatures between 4.2K and 1.2K can be easily obtained by pumping a bath of liquid 4He, and from 0.01K up to 0.8K by using a 3He-4He dilution refrigerator. But the dilution refrigerator is unstable above 0.8K. A 3He evaporation cryostat can cover the temperature range from 3K down to 0.3K. But it is not suitable for NMR experiment because of the poor thermal conductance of liquid 3He. A vortex refrigerator made by non-metal substances is suitable for NMR measurements of magnetic compounds around 1K. The basic concept of a vortex refrigerator is to obtain lower temperatures by filtering the superfluid component of helium using a "superleak" packed with very fine powder and to transfer heat by the interaction between superfluid vortex and normal fluid. In 1941 Kapitza suggested that a liquid helium entropy filter might be a convenient tool for obtaining a very low temperature. A vortex refrigerator as a cooling device was first invented by Staas and Severijins[ 1] in 1969. A few years before their research, Olijhoek, et al[2] used an older version of this refrigerator in order to study the properties of the supercritical flow of helium in capillaries and succeeded in obtaining temperatures below 0.8K. The experiments of Keller and Hammel[3] on gravitational and forced flow through a superleak showed that the presence of vortex lines is essential for cooling. Microscopically, mutual friction between superfluid and normal fluid is interpreted as a manifestation of the interaction between the elementary excitations in the normal fluid and the cores of the vortex lines. A vortex refrigerator can maintain the temperature from 2K down to 0.8K and remove the heat from 473
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samples by the entropy flow of normal fluid. An additional advantage is that magnetic fields needed in NMR experiments, have little or no effect on their performance. In this note, we describe mainly the characteristic features of the apparatus particularly designed and precisely constructed for NMR measurements of magnetic compounds.
APPARATUS The main parts of the refrigerator have been made by non-metal substances. Another improvement is that we made the apparatus in symmetry as shown in Fig. 1. It consists of two units, each of which is composed of a superleak and two chambers. The two units are connected by a heat exchanger which is immersed in the superfluid He II reservoir. A series connection of chamber A, superleak $2, chamber B, chamber C, superleak SI and chamber D is thermally insulated by a vacuum jacket which is immersed in the bath of liquid He II. A manganin heater of about 30fl is set in the each chamber. Since the basic principles of vortex refrigerators have been well described[l], no detailed discussion will be presented here. i
::::::::::::::::::::::::::::::::::
iiii J1
~
~
J2
~ ~
B
s:~i:i~i~s~i:~:i~i~i:i~i~i~i~:i~s~i~i~i~s~i~i~i~s~i~s~s~i~s~i~i~i~s~s~i~s~i~i~s~i~s~i~s~s~i~s~i~s~i~s~i~!~s~s~i~i~i~i~s~s~i~s~i~i~i~s~s~i Fig. 1 The schematic illustration of vortex refrigerator A,B,C,D:chambers, Ssuperleak, T:thermometer, H: manganin heater, J:capillary
Fig.2 Cooling chamber modified for high frequency NMR (10--200MHz) A:demountable cooling chamber(sample cell), B:variable capacitor, C:rf coil, D:sample, T: thermometer, S:superleak
When B or C is electrically warmed, A-S2-B or D-S1-C works as a fountain effect pump, respectively. Therefore we can make He II flow in both direction. This flow has a higher concentration of the normal fluid component.In the heat exchanger the normal fluid is converted back to superfluid and then enters superleak S1 or $2 sustaining high pressure.If the average superfluid velocity /Is is large enough to create vortices, owing to the interaction between the elementary excitations in the normal fluid and the cores of the vortex lines, the heat from the samples is transferred through capillary J1 or J2. Then C-S1-D or B-S2A works as a cooling device. The chamber D or A is in the lowest temperature. The lowest temperature depends on the inner diameter of J1 or J2 (qb 1 or qb2). The two tubes of superleak and the four chambers are all made of STYCAST1266 to prevent eddy current heating. Four 51~2 1/8W Allen Bradley carbon resistors (Ta, Tb,T~,Td) are used as the thermometers. They are immersed directly into He II flow in the
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chambers A, B, C and D. The resistances are measured by four-wires method. The measuring current is 0.1/x A. The empirical resistance R versus temperature formula of these thermometers has the form: logR
- A + BlogR + C (logR)2 + D (logR)3
(1)
where A, B, C and D are constants. To determine these coefficient, the carbon resistors were calibrated by using 4He vapor pressure thermometry. Then the data were fitted to the empirical formula (1). The superleaks S1 and $2 are made of STYCAST tubes having dimensions 10.0mm o.d., 6.0mm i.d. and length of 35mm, which are packed with aluminum oxide polishing powder particles having a normal diameter of 0.06/x m. The apparatus of second version is shown in Fig.2. This is the one for high frequency NMR experiment (10~200MHz). A variable capacitor operated from room temperature is close to sample cell to prevent the reduction of Q factor in NMR tuning circuit.
RESULTS and ANALYSIS As an example, temperatures of Ta, Tb, Tc and Td are shown in Fig.3(a) for various values of the heater power of the fountain pump unit. Using the symmetrical apparatus, we can get two times information by changing the direction of the He II flow. As the heater power increases, T~ goes up and fountain pump works. When induced He II goes into superleak $2, the temperature of sample cell T a reduces. When the heater power into the chamber C is about 50mW(which is signed by an arrow), the lowest temperature is obtained in the chamber A. At this point, we apply heat into chamber A to measure the cooling power of this refrigerator. The cooling power versus Ta is shown in Fig.4. Using more powerful pumping unit, the temperature of the bath is around 1.2K, then the temperature of the sample cell will be below 1K. 2 3 -DTa
OTb 2 1 ~ATc
AAA
AA
0
A
~oTd A A 9 l-+T_BathA A 0 0 0 0
"0 A v
0 0
A V
~
>
5 1
n n ~ 1 7 6 1 7 6 [] 0 []
1 0
50
100
150
Heater power (roW)
0
200
50
100
150
200
heater power (mW)
Fig.3(a) Temperatures of Ta,Tb,Tc, Td and He bath
Fig.3(b) Normal fluid velocity V, for various
for various values of the heater power of the fountain
values of the heater power of the fountain pump
pump unit ( 4~ 1=3.5mm, qb2=0.3mm)
unit
As mentioned above, mutual friction between superfluid and normal fluid is essential for cooling. Contrary to our experiment, if there is no mutual friction between them, normal fluid comes back into the chamber A, then Ta goes up. The existence of force F~n due to the mutual friction, normal fluid goes out from the chamber A with heat then Ta shows the lowest temperature. The normal fluid velocity Vn through capillary J2 to the reservoir can be estimated as follows. Thermal flow Q to the chamber A is composed of two parts. One of the stainless steel pipe and the other
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of copper lead-wire to the thermometer Ta.
Tba'h A rT'h Q - Asus f 1,:susdT+ cu tCc,,dT Lsus Lc~ r a
in which ~r
(2)
a
and ~r
are thermal conductivities, Asu s and Acu are cross sections, Lsu s and Lcuare
lengths of stainless steel pipe and the copper lead-wire, respectively. Q is removed with the normal component flow induced by the mutual friction Fsn.
O = B sus Crp,, V,,
Tbath + T o
(3)
2
where V is the normal fluid velocity, B st,s is a inner cross section of the stainless steel pipe, o" is entropy of He II and p , is the density of normal component. From the Eq.s (2) and (3), we get
2 V,, =
A sus ~r L sus
A c. ) +~ (Tb,,, h T,~) L c. ~r -
Bsus GPn ( Tbath+ Ta )
(4)
The change of normal fluid velocity Vn for various values of the heater power of the fountain pump unit is shown in Fig.3(b).
1.8
I
'15
'~~ 1.2 0.9 =e~O0.6 ~0.3 0
. . . . . . . .
1.1
1.15
1.2
1.25
Temperature (K)
1.3
.3
1.35
Fig.4 Cooling power of vortex refrigerator versus temperature of chamber A We have been doing NMR experiments of magnetic ions over the wide temperature range and want the cryostat which can maintain the temperatures from 2K down to 0.8K stably under rf pulses. In conclusion we can say that the vortex refrigerator made by non-metal substances is useful for NMR measurements around 1K. REFERENCES 1 Staas, F. A. and Severijns, A.P., Vorticity in He II and its application in a cooling device Cryogenics (1969) 9 422-426 20lijhoek, J. F. et al, Adiabatic superfiuid flow at supercritical velocities through capillaries Physica (1967) 35 483-486 3 Keller, W. E. and Hammel, E. F., Potentiometer for studying the liquid He II film Phys.Rev.Lett. (1966) 17 998-1001
Analytical Model to Calculate the Transient Thermo-Mechanical Behaviour of Long Thin Structures Cooled from a Pipe: Application to the LHC Dipole Thermal Shield
G. Pe6n, G. Riddone, L. R. Williams LHC Division, CERN, CH- 1211 Geneva, Switzerland
The most stringent cooldown for the 15 m LHC cryomagnets will take place on the test bench. The thermal shield has to be cooled from room temperature to around 75 K in 15 hours. A two stage analytical tool developed to calculate the thermomechanical cooldown behaviour of long thin structures cooled from a pipe is presented. The first stage predicts the temperature evolution over time from f'Lxed initial conditions yielding a temperature map of the thermal shield at fixed time intervals. The second stage calculates displacements of the shield structure due to internal temperature differences. Analytical temperature estimates are compared to measurements performed on the Cryostat Thermal Model (CTM).
INTRODUCTION The Large Hadron Collider (LHC) [1] currently under design at CERN incorporates a ring of superconducting accelerator magnets operating below 1.9 K in a bath of superfluid helium. The proportion of radiative heat inleak into the cryostats [2] that finally reaches the magnet cold mass must be minimized and is intercepted by actively cooled thermal shields and associated multilayer insulation systems (MLI) [3]. The thermal shield must be cooled to its operating temperature of 75 K during the same time interval as the cold mass by passing pressurised cold helium gas through a pipe forming an integral part of the shield structure. The asymmetric location of the cooling pipe causes transient asymmetric displacements of the shield structure during cooldown. An analytical tool to simulate the thermo-mechanical behaviour of long thin structures of this type cooled from a pipe has been developed. This tool operates in two stages. The first, in the form of a Fortran, program analyses the temperature/time evolution of the structure from fixed initial conditions and calculates a temperature map of the thermal shield at fixed time intervals. The heat transfer from the thermal shield cooling pipe to the helium gas, the heat conduction within the thermal shield and the temperature rise of the helium gas during its passage through the pipe were considered. The second stage, using the ANSYS [4] finite element program, calculates the thermally induced displacements of the thermal shield structure due to internal temperature gradients as defined in the temperature map for each time step (figure 1). INPUT: - Structuregeometryand material Cooling/ WarmingFluid: Pressure (Average) Inlet Temperature (As a functionof time) Mass Flow Rate
THERMAL
-
I
J
PROGRAM
---) Screen TemperatureMaps __..q after givenTime intervals
STAGE 1
Stresses
F. E. M. ANSYS
INPUTFILEFOR ANSYS
Strains
PROGRAM
Displacements
III I
STAGE2
Figure 1 Analytical tool to calculate the transient thermo-mechanical behaviour of long thin structures 477
I
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CTM T H E R M A L SHIELD COOLDOWN CONDITIONS The measured helium gas inlet temperature and mass flow rate representing typical cooldown conditions of the CTM [5] thermal shield are shown in figure 3 and have been used to validate the mathematical model. The CTM cold mass and thermal shield were cooled simultaneously under an insulation vacuum of 10 -4 Pa. to minimize heat exchange by radiation and residual gas conduction.
MATHEMATICAL MODEL Basic Assumptions, Initial and Boundary Conditions - Heat transfer by radiation and gas conduction to the thermal shield and heat conduction through the support posts is small compared to the heat extracted by the cooling pipe. - The initial temperature of the thermal shield is everywhere constant and fixed at 147 K to allow comparison with experimental results. - The helium mass flow rate is constant (0.5 g/s), the inlet temperature as a function of time is approximated by a 6th order polynomial function and its pressure drop along the cooling pipe is negligible. Equations Governing the Thermal Shield Cooldown Energy equations in the helium gas coolant (1) and in the cooling pipe (2): hgas. 2n:. Rpipe. dz. (Tpipe -Tgas) = mgas. Cpgas. dTgas
(1)
dTshield dTpipe hgas. 2tr. Rpipe 9(Tpipe - Tfluid ) + Kshield. thshield, d z . ~ dx = Ppipe'27r'Rpipe .thpipe.dz.Cpipe. dt
(2)
Heat conduction in the thermal shield structure in the circunferential direction (3) and in the longitudinal direction (4): Kshield thshield dz dT = mshield 9Cshield . dTshield ~ dx dt
(3)
Kshield .thshield .dx. dT .C dTshield dz = mshield shield dt
(4)
9
9
9
Continuity equation in cooling pipe (5):
/9 gas "u gas =
mgas t r . R 2. pipe
(5)
Description of the Solution Method The thermal shield is discretized into 3000 elements within each of which the temperature is defined to be constant. The temperature evolution with time is calculated at time intervals which depend on the helium mass flow rate, the average helium temperature in the pipe and on longitudinal dimension of the element. The energy exchange between elements is calculated at every time interval and from it the temperature map for that step.
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RESULTS The temperature of the CTM thermal shield was measured at 5 points as shown in figure 2" TI1, TI2 and TI3 at the inlet section, TM 1 at the middle section and TR1 at the rear. TI 1 TM1
Thermal shield i
TI1 .............
TM1 \
TR1
..........
........
TI2 .
\
.
.
.
.
.
.
.
.
.
.
.
.
.
1305mm i'", @50rnmj I i 9 _L~/
\ \
~
..........
\:I72. . . . . . . . . . . . . . . . . . . . . . . . . THein
Y! X ~ - - / .... . Coolingpipe
S ;Z
THeout
10m
Figure 2 Locations on the thermal shield where temperatures were measured The measured and calculated temperature evolution with time for points TI1, TM1 and TR1 is shown in figure 4. TI2 and TI3 follow very closely the temperature profile of TI 1.
---A--- TI1
14~I -
i
...• ...tr.. TR1
"1~,
,&
1~1~" "A
"'X,,
9 TCI1 :. l~ He • TCM1 = Tin He = TCR1
2.5
- . o... ~
"O. .... ~ ..... '1~~176176176176 ~ "'!'1 . . . . .
-&- .& . . . . . . .
~. . . . . . . . . . .
1.5
T (K)
m He g/s
6( 1
40 A
A
A
A
,,
.
.
. r .
A
A
A
9
0.5
@
20 01 0
10
20
30 TIME (HOURS)
~
40
t
50
0
60
Figure 3 Evolution with time of thermal shield temperatures. Comparison between measured (TI1, TM1 and TR1) and calculated values (TCI1, TCM 1 and TCR1)
DISCUSSION During the transient stages of cooldown (up to about 20 hours) where heat transfer from the thermal shield to the cooling pipe is most rapid, the analytical predictions of shield temperatures match the experimental results to within + 6 K. Later, under steady state conditions, the effects of heat transfer to the shield (not considered in the thermal model) become significant in comparison to the heat transferred to the pipe. This explains that the measured shield temperatures (as shown in figure 3) appear 12 K warmer than those predicted from calculation which become asymptotic to the helium inlet temperature. It is foreseen to improve the steady state predictions by including radiation effects in the mathematical model.
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CONCLUSIONS The first stage analysis of transient thermal behaviour of long thin structures cooled from pipe has proved to simulate accurately the behaviour of this type of structure in the CTM, and therefore it is to be used for the design of the thermal shields for the next generation of cryostats. Measurements are to be carried out to corroborate the mechanical part of the model.
ACKNOWLEDGEMENTS Guillermo Pe6n, on leave from ICMA, (CSIC-Universidad de Zaragoza), acknowledges a PhD grant from Consejo Superior de Investigaciones Cientfficas (1993).
SYMBOLS ragas Gas mass flow rate Heat transfer coefficient gas-pipe hgas Rpipe Internal pipe radius Tgas Gas temperature Tpipe Pipe temperature m shield Mass of a shield element [3gas Gas density [3pipe Pipe material density
Ugas
Gas velocity
Kshield Shield material thermal conductivity thpipe Pipe thickness
th shield Shield thickness Tshield Shield temperature Cpgas Gas specific heat Cpipe material specific heat Cshield Shield material specific heat
REFERENCES Evans, L.R. The Large Hadron Collider project, presented at this Conference Brunet, J.C. et al. Design of LHC prototype dipole cryostats Cryogenics (1992) 34 ICEC Supplement 191-194 Ferlin, G. et al., Comparison of floating and thermalized multilayer insulation systems at low boundary temperatures, presented at this Conference Ansys, Swanson Analysis Systems Inc., Houston Dufay, L. et al., A full scale thermal model of a prototype dipole cryomagnet for the CERN LHC project Cryogenics (1994) 34 ICEC Supplement 693-696
Characterization
Kunihiko
Daido*,
Of B o n n e t l e s s
Kazuhiro
Cryogenic Valves
Yoshikawa*,
Ryuji M a e k a w a * * ,
and S a d a o
Satoh**
*Fujikin I n c o r p o r a t e d , 9 0 - S i n k e - H i g a s h i m a c h i , H i g a s h i - O s a k a 577, J a p a n * * N a t i o n a l I n s t i t u t e For Fusion S c i e n c e , F u r o - c h o , C h i k u s a - k u , N a g o y a 4 6 4 - 0 1 , J a p a n
C o n t r o l o f liquid helium u s e d for c o o l i n g the L a r g e Helical D e v i c e ( L H D ) can be made m o r e e f f e c t i v e w i t h Bonnetless Cryogenic V a l v e s ( B L C V ) , w h i c h can be l o c a t e d in the p r o p e r p l a c e s for use inside the helical coil by b e i n g c o n s t r u c t e d in such a w a y as to be p l a c e d inside v a c u u m i n s u l a t e d vessels, c o m p a r e d with the c o n v e n t i o n a l l o n g - b o n n e t c r y o g e n i c valves, w h i c h must be i n t e g r a l l y i n s t a l l e d in a valve box. The a u t h o r s c o n d u c t e d p e r f o r m a n c e e v a l u a t i o n t e s t s of 3 d i f f e r e n t kinds of B L C V to study the r e l a t i o n s h i p b e t w e e n the r e s p e c t i v e valve s t r u c t u r e s and their p e r f o r m a n c e and reliability, as well as the e f f e c t i v e n e s s of d e s i g n m e t h o d s .
INTRODUCTION The L a r g e H e l i c a l D e v i c e ( L H D ) is a n u c l e a r fusion test s y s t e m in w h i c h all coils are made s u p e r c o n d u c t i v e and w h i c h r e q u i r e s the g u a r a n t e e of c o n s t a n t c o o l i n g at liquid helium t e m p e r a t u r e to e n s u r e s u p e r c o n d u c t i v e a c t i o n of the coils [1,2]. M o r e o v e r , it is also n e c e s s a r y to a c c u r a t e l y c o n t r o l the helium inside the c o o l i n g piping to u n i f o r m l y cool the object m a t e r i a l d u r i n g c o o l - d o w n o p e r a t i o n . C o n v e n t i o n a l l y , v a l v e s for c o n t r o l l i n g lowt e m p e r a t u r e helium have been v a l v e s of the e x t e n s i o n b o n n e t s t r u c t u r e , in which the a c t u a t o r is i n s t a l l e d o u t s i d e the v a c u u m i n s u l a t e d vessel. H o w e v e r , in c a s e s involving such valves, the i n s t a l l a t i o n area is r e s t r i c t e d . W h e n v a l v e s of this type are u s e d in LHD, it is n e c e s s a r y to e x t e n d the r e s p e c t i v e helium piping to a valve box, b e c a u s e the valves are i n t e g r a l l y i n s t a l l e d in a valve box l o c a t e d o u t s i d e the LHD body. On the o t h e r hand, in cases of B L C V use in w h i c h the a c t u a t o r and the b o d y are i n t e g r a t e d and all the valves are p l a c e d in the v a c u u m i n s u l a t e d vessel t e s t e d at this time, the i n s t a l l a t i o n of the piping and valves can be simplified, b e c a u s e the v a l v e s are d i r e c t l y i n s t a l l e d to the piping inside the LHD, p r o v i d i n g many a d v a n t a g e s , such as r e d u c e d e q u i p m e n t size, p o s s i b i l i t y o f heat leak r e d u c t i o n , etc. R e p a i r s or r e p l a c e m e n t etc. are r a t h e r difficult with B L C V b e c a u s e all of the valves in their entirely are p l a c e d in the v a c u u m i n s u l a t e d vessel. For that r e a s o n , m a i n t e n a n c e - f r e e o p e r a t i o n and high reliability are r e q u i r e d from t h o s e valves. H o w e v e r , t h e r e have not been any of these r e p o r t e d u n d e r a c t u a l w o r k i n g c o n d i t i o n s . It was t h e r e f o r e n e c e s s a r y to make an e v a l u a t i o n for p l a c i n g this type of valve into use, i n c l u d i n g p r a c t i c a l a d o p t i o n for 481
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LHD. Our present report will describe the results of studies made on the relationship between the respective valve s t r u c t u r e s and their p e r f o r m a n c e and reliability based on performance evaluation tests of 3 different kinds of BLCV, as well as the validity of their design methods. STRUCTURE
OF BLCV
Figs. 1, 2 and 3 indicate the s t r u c t u r a l drawings of BLCV, while Fig. 4 shows a c o n v e n t i o n a l l o n g - b o n n e t type cryogenic valve. C h a r a c t e r i s t i c s of BLCV are as follow" 1 . B e c a u s e the entire valve body and the a c t u a t o r are installed in a vacuum insulated vessel and kept at low t e m p e r a t u r e , metallic bellows are used for the dynamic seal of the fluid inside the body and o p e r a t i n g fluid inside the a c t u a t o r . 2.The joints of the respective parts are sealed t h r o u g h welding to completely prevent any leakage into the v a c u u m insulated vessel. 3.The o p e r a t i n g fluid is helium gas, because the o p e r a t i n g fluid is cooled.
r i
Figure 1 B LCV normally open so ft seat
i
Figure 2 B LCV normally open metal seat
Figure 4 long-bonnet Figure 3 BLCV valve normally closed so ft seat
EXPERIMENTS To evaluate the p e r f o r m a n c e of BLCV, tests were c o n d u c t e d with the following flows on the following 3 types of valves (A, B and C): Type A: Normally open, soft seat, o n - o f f ( F i g . l ) Type B: Normally open, metal seat, E% (Fig.2) Type C: Normally closed, soft seat, o n - o f f (Fig.3) (Test flows) Step l : V a l v e seat leak test and external leak test under normal t e m p e r a t u r e . Step 2:Valve seat leak test and external leak test in a state where the valve body is immersed in liquid n i t r o g e n . ( T h i s m e t h o d of cryogenic test is usually practiced by valve m a n u f a c t u r e r s )
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Step 3 : V a l v e seat leak test and e x t e r n a l leak test after 200 times o f o p e n i n g and c l o s i n g the valve in a s t a t e w h e r e the valve body is i m m e r s e d in liquid n i t r o g e n . Step 4 : V a l v e seat leak test and e x t e r n a l leak test u n d e r n o r m a l t e m p e r a t u r e . Step 5 : 1 0 0 0 times o f o p e n i n g and closing a valve filled inside w i t h liquid helium in a v a c u u m i n s u l a t e d vessel.
3
r""
Step 6 : V a l v e seat leak test and e x t e r n a l leak test u n d e r n o r m a l t e m p e r a t u r e .
I
LCV
Step 7 : V a l v e seat leak test and e x t e r n a l leak test in a s t a t e w h e r e the valve body is i m m e r s e d in liquid n i t r o g e n . F i g u r e 5 Test e q u i p m e n t RESULTS
( S t e p 5)
AND DISCUSSION
Fig. 6 i n d i c a t e s c h a n g e s in leak r a t e s from valve seats in their r e s p e c t i v e s t a t e s . While there was s o m e d i f f e r e n c e in the a b s o l u t e value of leak rate due to d i f f e r e n c e s in types, no m a r k e d c h a n g e leak rate w a s o b s e r v e d b e t w e e n the r e s p e c t i v e v a l v e s b e f o r e and after the v a l v e s w e r e o p e n e d and c l o s e d 1200 times. M o r e o v e r , while soft seats i n d i c a t e d a l a r g e r i n c r e a s e of leak rate at low t e m p e r a t u r e than at n o r m a l t e m p e r a t u r e c o m p a r e d with metal seats, this is p r o b a b l y b e c a u s e P C T F E (the soft seat m a t e r i a l ) has l a r g e r i n c r e a s e r a t e s of b o t h s h r i n k a g e and h a r d n e s s at low t e m p e r a t u r e c o m p a r e d with the 3 1 6 L S t a i n l e s s Steel (the metal seat m a t e r i a l ) . As for e x t e r n a l leak rates, the test r e s u l t s w e r e excellent, with v a l u e s no h i g h e r than the m i n i m u m d e t e c t i n g s e n s i t i v i t y o f the m e a s u r i n g i n s t r u m e n t s , in all states. Fig. 7 i n d i c a t e s the r e s u l t s o f c a l c u l a t i o n of heat leak /Cv v a l u e d u r i n g the c o n t i n u o u s o p e n i n g / c l o s i n g of r e s p e c t i v e v a l u e s by the m e t h o d of c a l o r i m e t r y by b o i l - o f f m e a s u r e m e n t m a d e by using a f l o w m e t e r i n s t a l l e d on the d o w n s t r e a m side. To o b s e r v e the heat leak rate in r e l a t i o n to valve s t r u c t u r e , we can see that, in the case of a n o r m a l l y c l o s e d type valve, the liquid inside the valve and the o p e r a t i n g helium inside the a c t u a t o r are p l a c e d side by side, w i t h only one piece of b e l l o w s b e t w e e n them, and t h e r e f o r e , t h e r e is a l a r g e r heat leak d u r i n g the supply of o p e r a t i n g fluid c o m p a r e d with a n o r m a l l y o p e n type valve. To c o n t r o l heat leak at a low level, it is n e c e s s a r y to not only c h a n g e the a c t u a t i n g m o d e but to also r e d u c e the a c t u a t o r c a p a c i t y d e s i g n in a w a y to k e e p the o p e r a t i n g p r e s s u r e at a low level, b e c a u s e a lot of heat leak c o m e s from the o p e r a t i n g fluid. M o r e o v e r , since the a c t u a t o r c a p a c i t y s u d d e n l y i n c r e a s e s as the n o m i n a l d i a m e t e r of the valve b e c o m e s l a r g e r b e c a u s e of the r e l a t i o n b e t w e e n the t h r u s t and the s t r o k e of the valve it is m o r e realistic to limit the n o m i n a l d i a m e t e r of the valve ( m a x . 2 B for e x a m p l e ) . In r e g a r d s to durability, the s e t t i n g of p r o p e r c l e a r a n c e and s e l e c t i o n of the c o m b i n a t i o n of m a t e r i a l s are i m p o r t a n t for p r e v e n t i n g s e i z u r e and galling b e c a u s e the sliding unit is also c o o l e d in B L C V . In the tests p r e s e n t e d at this time, h o w e v e r , all the v a l v e s w o r k e d n o r m a l l y even after being o p e n e d / c l o s e d 1200 times, i n c l u d i n g t h o s e using metals of different k i n d s ( 3 1 6 L S t a i n l e s s Steel and a l u m i n u m b r o n z e ) and t h o s e using the s u r f a c e t r e a t m e n t m e t h o d ( e l e c t r o l e s s nickel p l a t i n g on 3 1 6 L S t a i n l e s s Steel).
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Figure 6 C o m p a r i s o n of seat leak rates
0 TYPE A
1
TYPE B
TYPE C
Figure 7 C o m p a r i s o n of heat leak
CONCLUSION T h r o u g h these tests, it has been c o n f i r m e d that b o n n e t l e s s c r y o g e n i c valves have sufficiently usable p e r f o r m a n c e for c o n t r o l of liquid helium in v a c u u m insulated vessels. 1. C o n f i r m a t i o n has been made that PCTFE, which has many a c h i e v e m e n t s in sealing the m a t e r i a l of valves up to liquid h y d r o g e n t e m p e r a t u r e , has sufficient durable sealing capacity, even at liquid helium t e m p e r a t u r e . 2.It has been c o n f i r m e d that metallic bellows have sufficient durability for the o p e n i n g / c l o s i n g of valves as an a c t u a t o r and for sealing the valve body. 3.It has further been c o n f i r m e d that seizure and galling on the sliding face inside the valve can be p r e v e n t e d by c o m b i n i n g different kinds of metal ( 3 1 6 L stainless steel and a l u m i n u m b r o n z e ) or a d o p t i n g surface t r e a t m e n t (nickel plating on 316L stainless steel). 4 . H e a t leak can be r e d u c e d in normally open type valves. F M E C A (failure mode, effect and criticality analysis) was also c o n d u c t e d for evaluating the reliability of the r e s p e c t i v e s t r u c t u r e s . In the case of the normally closed type, there is a high d e g r e e of risk to e q u i p m e n t , such as c o n t i n u i t y b e t w e e n the inner part of the valve and the a c t u a t o r , l e a k a g e of o p e r a t i n g fluid into the v a c u u m insulated vessel, etc. in cases of b r o k e n bellows, and it p r e s e n t s a p r o b l e m about being a d o p t e d for large e q u i p m e n t , such as LHD. From the a b o v e - m e n t i o n e d results, the most suitable valves for LHD are normally open, soft seat types (Fig. l). 1 . N o r m a l l y open type has little heat leak because the valve is often used in a normally open state and no o p e r a t i n g fluid is r e q u i r e d in the o p e r a t i n g state from the structural viewpoint. 2 . T h e soft seat type has a small seat leak rate. (Usable for c h e c k i n g the respective pipelines by dividing them into blocks in valve units.) We are planning to make further studies for the d e v e l o p m e n t of drive equipment and o p e n - c l o s e p o s i t i o n d e t e c t i n g devices most suitable for this valve for use in vacuum and high m a g n e t i c fields as a goal for the future. REFERENCES 1. Sadao Satoh et al. C o n c e p t u a l design of c r y o g e n i c r e f r i g e r a t i o n for the Large Helical Device, Fusion E n g i n e e r i n g and Design 2 0 ( 1 9 9 3 ) 1 2 9 / 1 3 6 2 . S a d a o Satoh et al. Design and C o n s t r u c t i o n of C r y o g e n i c S y s t e m for LHD , OB7-5 I C E C 1 6 / I CM C ( 1 9 9 6 )
F l e x i b l e C o r r u g a t e d C r y o t r a n s f e r l i n e s , L o n g T e r m E x p e r i e n c e at J E T a n d the E x p e r i e n c e with Supercritical Helium Flow Conditions
W. Obert, C. Mayaux JET Joint Undertaking, Abingdon, OXON, OX 14 3EA, UK
Long flexible corrugated cryotransferlines are widely used at JET for the supply of liquid nitrogen and liquid and supercritical helium. Their low thermal losses allow to use a central distribution system with individual lines to each user with line lengths of up to 90 m. The easy installation of the fully tested cryolines, their flexibility, high pressure stability and compatibilty with the remote handling requirements of JET were important design specifications. Some of the cryolines have been in operation for over 15 years and the paper reports about their excellent reliability records and the main performance data.
INTRODUCTION For the cryosupply of JET - the world's most advanced thermonuclear fusion experiment - the concept of a central distribution system with individual cryolines to a network of load assemblies taking liquid nitrogen and liquid/supercritical helium has been adopted. This concept was chosen in order to keep all control systems (in particular valves and instrumentation) in easily accessible areas, because the loads themselves cannot be accessed in the D-T phase of the project due to the expected radiation level in the vicinity of equipment inside the toms hall. This concept required, however, long transferlines and thus the development of the JET low loss cryolines. The use of low loss long flexible JET cryolines for liquid helium has been first reported at the ICEC12 [1]. These line were a further development of flexible lines used for CERN [2] implementing as a new element a liquid nitrogen radiation shield as an integral part of the cryoline. JET has subsequently commissioned over 30 long (>50m) flexible corrugated cryolines which represent a total length exceeding 2 km. Individual lengths of the cryoline are up to 90 m. Some of the cryolines have been in use for over 15 years and JET gained considerable experience with these lines during this period. JET installed new flexible corrugated cryolines for the 1993-94 experimental campaign for the supply of supercritical helium (in contrast to the two phase liquid helium which is supplied to all other users).
CRYOLINE SPECIFICATION Boundary conditions The operation of the cryogenic loads which are mainly large scale cryopumps at various places at the experiment (for details see [3]) are located in a potential radiation area with high inducted eddy currents. This requires that the cryolines: 9 have low thermal losses 9 have high pressure stability and small thermal mass for fast eooldown 9 are radiation resistant and compatible with remote handling i.e. connection, disconnection and line exchange must be easily performed by a remote handling manipulator 9 have an electrical break to separate the electrical potential of the load from that of the cryoplant 9 are flexible in order to allow for the movement of cryo-loads 9 can be easily r e a r r a n g e d to make space for new equipment 9 can be used as vacuum line for pumping of vacuum spaces of the cryo-loads 9 have minimum time requirement for on-site installation 485
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Thermal losses: The considerable length of the cryolines requires a system with low thermal losses. This_could be achieved owing to the fact that the lines have no intermediate lossy joints and are fully thermally shielded by helium boil-off and liquid nitrogen. Thus the thermal losses of the cryolines are small in the overall thermal balance despite of their length. The specific data are listed in table 1.
Table 1 Thermal loads to JET low loss Helium-cryotransferlines
Liquid helium go-line
<
10 mW/m
Gaseous helium return line
< 100 mW/m
Liquid nitrogen shield
<
2
W/m
Pressure stability An important specification for the cryolines is their pressure stability. The cryosupply for the JET cryo-loads operates at pressures of up to 12 bars for special transient conditions such as cooldown by using the 12 bar high pressure outlet of the cryoplant. The cryolines have also to cope with large pressure excursions of up to 10 bars from unplanned regenerations (quench) of the cryopumps. Due to the special corrugations of the pipes high pressure stability is provided for internal pressure as shown in table 2.
Table 2: Pressure stability of corrugated pipes used for JET cryolines
corrugated line diameter, wall
burst
elongation pressure
buckling 5%
14/18 mm, 0.3mm
>300 bar
>50 bar
> 160 bar
39/44 mm, 0.4mm
> 130 bar
>22 bar
> 80 bar
60/66 mm, 0.5mm
> 130 bar
>17 bar
> 70 bar
Line flexibility Another important issue for the use of the cryolines is their flexibility under operation conditions as they have to cope with the movement of the cryoloads during the experiment which is in some places up to +/100mm with a frequency of 2 Hz. For installation great care has to be taken to avoid torque during the movement as the line has no flexibility for torsion. In this context it has to be noted that KEK/Japan which adopted later the JET cryolines uses them to pull their cryo-load completely out of the operation area without dismantling the cryoline. Vacuum One advantage of the concept of long flexible cryolines is that they can be fully tested cryogenically for thermal performance and vacuum leaks (leakrate < 10-~ mbar l/s) at the manufacturers and installed within a short period of time (typically 1-2 days) by a cable pulling team. The vacuum of the cryolines is also used at JET as a vacuum pumping line for interspaces of the cryoloads. Electrical break/Radiation resistance The cryoplant has to be electrically insulated from the cryoloads for a potential of up to 2 kV and the cryocouplings had to be designed accordingly. The interior shrink coupling is made of a cryogenically and radiation compatible polymer, the coupling itself coated with a radiation compatible glass epoxy sleeve and the connection flanges enamelled or equipped with a polymer spacer. For all the internal spacers of the cryoline and the superinsulation, radiation resistant material had to be used. Material The use of austenitic stainless steels AISI 304 or 316 (equivalent DIN 17440: 1.4541, 1.4301, 1.4435, 1.4306) is a prerequisite to be compatible with the weldability for vacuum leaktighness, ductility at cryogenic temperatures, low thermal conductivity, corrosion and good emmissivity properties.
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JET CRYOLINES Over 30 long cryolines (with lengths between 45 and 90 m) are in use at JET. They originate from the central distribution area and are branched off to the various loads (see Figure 1). The inner liquid helium go lines have diameters between 10-20mm depending on the needs of the specific cryo-loads, the helium return lines have a annular gap of 3-5mm and are rated to minimize the pressure drop on the return in order to guarantee stable operation of the users. The outer dimension of the helium cryolines are --110 mm which gives a reasonable flexibility for installation. Two different types of liquid helium lines are used at JET. There is one type with an integrated go and return nitrogen shield loop and one with a single flow nitrogen loop which is used to supply a nitrogen user at the outlet of the cryoline (see figure 2). Due to the installation requirements at JET the cryolines are installed with height differences of up to 10m and a bending radius of 1.5m for the large 110 mm LHe lines and less for smaller lines and have usually a loop which allows to accommodate a few metres spare length.
He boil off
LN
LHe go
Super insulation
Figure 1: Bundles of JET cryolines leaving the main penetration into the JET basement.
He boil off LHe g o - ~
Vacuum Figure 2 Cryoline with integrated LN loop (Go and return)
LN
Super insulation
~
JG96.317/10
Vacuum Cryoline with single flow LN supply
Couplings/Remote handling A key element for the ease of handling and the performance of the cryolines are the terminations at the end of the lines. The termination of the cryolines provides an entire vacuum enclosure and no vacuum connections are required for installation. Using for the internal go line a simple shrinkage seal and for the external seal a JET standard Vacuum O-ring sealed V-band flange guarantees that the cryoconnection can be opened and closedwith ease even by remote handling with a single-arm manipulator.
Figure 3: Remote handling tool for disconnecting cryoline by opening V-band clamp
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EXPERIENCE AND LESSONS In terms of long term performance the cryolines and the JET cryopumps are practically maintenance free [4]. A spare line which was procured at the beginning of the project is still waiting to be used - see Figure 4 which shows the way of delivery of a 90m long He-line on a simple cable drum.. In particular there were no detectable problems of fatigue, despite considerable thermal cycling, and no deterioration of the thermal performance of the lines over the years could be detected. The flexibility of the lines allowed rearrangement and even complete rerouteing of lines in order to accommodate the installation of new experimental equipment at JET in a short period of time with minimum efforts. No remedial work on the lines has been necessary over the last 15 years. However several design improvements have been made. These concern the details of the inner thermoplastic shrink seal and the electrical breaks. These can crack if overstressed, therefore clearances and tightening has to be followed carefully. Flow of supercritical helium of up to 17 g/s through the corrugated pipes could be established without any problems or flow instabilities. The measured pressure drops were well within the calculated values (< 50 mbar for 2.7 bar/4.5K and a Reynoldsnumber of 3xl 0~). This fits well with a friction factor of 0.098. Whilst there were no vacuum problems with the cryolines themelves problems were found with the feedthroughs for the instrumentation and heaters for the adsorber packages where the glass insulation developed cracks by the force from the sealing clamp. With improved clamps and feedthroughs the problems was solved. During operation with LHelium air leaks cannot be detected due to the large cryopumping of the helium lines and if the systems are operated for a considerable time a stoppage of the helium flow can cause a considerable pressure rise of the insulation vacuum which may cause a problem for the subsequent cooldown. This is in particular important when regular thermal cycling of the loads takes place such as at JET (daily to weekly).
Figure 4 Storage/delivery drum with spare 90m long JET liquid helium cryoline and single pass liquid nitrogen line.
CONCLUSION OUTLOOK Long flexible cryolines at JET allow the easy transport of cryogens over long distances with minimum thermal loads into areas with restricted access. The long term experience with these lines shows that they can be regarded as practically maintenance free. The use of similar lines with integrated superconductors will be an ideal concept for a combined supply of electrical power and cryogens to the superconducting coils for future fusion devices such as ITER. REFERENCES 1. Obert W. et al., Low Loss Flexible Cryogenic Transfer Lines for JET, International Cryogenic Engineering Conference - 9, Butterworth, Guildford, UK (1982), 100-104 2. Blessing H. et al., Flexible Cryogenic Helium Transferlines, Advances in Cryogenic Engineering, Vol. 27 (1982), P761 3. Obert W. et al. ffhe JET Cryosystem, Overview and Experience', Advances in Cryogenic Engineering CEC 9 (1993), Vol 39,pp 493-500, Edit. E Kittel, Plenum Press, NY 1994 4 0 b e r t W., JET Experience with the large Scale Cryogenic System, Proceedings of Symposium NIFS, May 1996, Toki. Japan
Experimental Results with Superinsulated Cryogenic Transfer Line Test Modules in THISTA W. Lehmann, M. Seidler, M. Stamm Forschungszentrum Karlsruhe, Institut ffir Technische Physik, Postfach 3640, D-76021 Karlsruhe, Germany The FZK THISTA Facility is dedicated to tests with regard to thermal insulation in cryogenic apparatuses. It can be used in the temperature range of 4 K < T < 300 K and has a maximum space for testobjects of approx. 1,3 m dia and 2 m in height. Our experience with superinsulated cryogenic apparatuses and transferlines so far resulted in a strong degradation of the MLI quality with decreasing diameters. Good experience with blanket type MLI for larger sized cryostats encouraged us to test this MLI blanket technique also for transferline applications. It is reported about design and equipment of THISTA and first results with transferline test modules concerning MLI blanket quality between ambient temperature and LN2- and LHe-temperature respectively.
INTRODUCTION Although Superinsulation or Multilayer Insulation (MLI) has been applied in cryogenics for more than three decades, its high-tech effect is no self-evident at all. Especially as regards MLI normally applied by winding technique in cryotransfer lines, we were repeatedly faced with unsatisfactory qualities of insulation. These were far from the ideal values in theory and also different from published optimum values for MLI. A major deterioration of the MLI quality with decreasing cylinder and line diameters is attributable to the growing deviation from the superinsulation with the ideal configuration ("floating plane reflection shields in the high vacuum"). Figure 1 shows the qualities of MLI versus cylinder and pipe diameters, respectively, measured at FZK/ITP on various superinsulated systems. The increase in dl resulting from the practical assumption d = dcyl,pipe = dcold instead of the mean diameter of insulation dm,MLI = [(da - di)/en da/di] is comparatively small. For the J E H I E R IR 300.12 MLI discussed below it is < 2% for dcold = 500 mm up to approx. 72% for dcold = 12 mm. The principal reasons of deterioration of the quality of insulation with decreasing diameter in case of MLI achieved by the winding technique are: 9 increased longitudinal thermal conduction through the foils from warm to cold; 9 normally enhanced density of layers with more radial thermal contact points; 9 poorer axial and radial pump cross-sections and conductances with resulting vacuum degradation between the layers, at least in case of manual arrangement; 9 increasing percentage of surfaces with defects due to joints or overlappings. After successful experiments involving an industrial blanket MLI on cylinder shaped [1], curved and plane cryostat test surfaces kept between room temperature (RT) and 80 K it suggested itself to examine whether this type of MLI allows good insulation qualities to be achieved for the transfer lines too. TEST FACILITY AND EXPERIMENTAL PROGRAM The former FZK/ITP test bench TESSI last used to examine MLI for cryostats between RT and 80 K [1,2] was extended to permit a broader spectrum of applications and modernized with regard to measuring data acquisition and processing, respectively. The modified facility is called THISTA (test facility for thermal insulation). It opens up a field of applications in 489
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studies which can be made of the t h e r m a l insulation in cryoapparatuses and cryogenic transfer lines at t e m p e r a t u r e s between 300 K and 4 K (Fig. 2). In THISTA a previously used v a c u u m t a n k (1) is installed upright; the test volume is about 1300 m m dia. x 2000 m m height allowing accommodation of a n u m b e r of test objects. Until now, the boil-off calorimetry method has been used exclusively. For this, the top part of the cryostat has been equipped with the following internals: test (measurement) chamber (2) (LHe, LN2), i n t e r m e d i a t e guard chamber (3) (LHe, LN2), and external LN2-guard (4) with special feed and e x h a u s t lines in t h e r m a l contact with radiation cooling shields. The test modules can be enclosed within a guard copper shield (5), t h e r m a l l y insulated (a), LN2 cooled (b) or electrically heated by a thermocoax heater element. In the investigations described the cryogenic transfer line test modules are filled with LN2 and LHe, respectively. They are linked to the m e a s u r e m e n t chamber by means of a CF multi-interface flange (6) which allows easy replacement of various test modules. The m e a s u r e m e n t chamber and the intermediate guard chamber are suspended at the cryostat lid (7) by means of a support plate (8). To avoid undesired effects such as "stratification" and "superheating" in the test and guard chambers and to ensure an "isothermal surface" independent of the liquid levels in the tanks, the following measures were taken. Vertical copper sheet strips were provided in the stainless steel tanks; the vapor escapes from the test modules into the test (measurement) chamber radially t h r o u g h an upright sintered metal pipe closed on top. This sintered element (9) reduces faulty m e a s u r e m e n t s of the rate of evaporation due to "stratification and s u p e r h e a t i n g effects." It contributes to an uniform b a t h t e m p e r a t u r e and to the formation of relatively small vapor bubbles because larger bubbles might induce oscillations in the exhaust gas flow to be measured. The quality of insulation is determined by m e a s u r e m e n t of the escaping gas flow rate t a k i n g into account the storing effect in the liquid - vapor - exchange volume and the zero losses of the test facility (Hastings mass flow meters and gas meters). In order to provide an adequate insulation surface and m e a s u r i n g effect also in case of small diameters, the test module a r r a n g e m e n t sketched in Fig. 2 (supplementary figure) with a total pipe length of 19,2 m, ND 12, and 0,72 m2 heat exchange surface was used. P a r a m e t e r s of MLI, Transferline Test Modules, Test conditions and Testprogram are visible from Table 1. Table I
Test p a r a m e t e r s
Superinsulation Reflectors
J E H I E R IR 300.12 [1]t12 layer-blankets 6 pm PE foils, 2 x 400 A aluminized 2 mm dia/0,1% perforation Spacers PE tulle, 2 m m mesh size, 5 g/m2 N u m b e r of b l a n k e t s / b l a n k e t layers 0, 1, 2, 3 / 0, 12, 24, 36 Test modules
Vertical LN2-, LHe-filled SS-cylinders of 30, 54, 105~ 219 mm dia First layer on cylinder: 0,1 mm a l u m i n u m foil Cu-tube a r r a n g e m e n t of 12 mm dia with 7 vertical tubes First layer a l u m i n u m adhesive tape 3 M Scotch 850
Guard shield
u n t r e a t e d copper plate cylinder 450 mm dia
Insulation vacuum [mbar] Tcold surface [K] Twarm surface (450 ~ shield) [K] Tvacuum tank [K]
5 x 10-7 ~ 5 x 10-2 4,2 / 77,4 approx. 280 and 303,318 respectively 290
The MLI blankets were tailored and fitted on the test modules by the technician. For the diameters 30, 54, 105,219 mm the blankets were a r r a n g e d with joints and staggered on the perimeter. When only one blanket (12 layers) was used, the a r r a n g e m e n t was such that no gap was visible at the joint. The blankets were cut out from the m a t e r i a l delivered on reels (1.5 m width) such t h a t the module could be insulated without axial joint over the approx.
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2 m length in the axial direction and t h a t the MLI blankets could be 20 mm "turned down" into the bottom and lid zones of the cylinder module. Corresponding to the number of blankets on the perimeter, circular blanket disks overlapping all round were used for further insulation. The positions of the blankets were fixed in the bottom and lid areas with aluminum adhesive tape adhering on one side (3 M Scotch 850). The desired and achieved mean layer density of the whole blanket MLI was 3 to 4 layers/mm. It was more difficult to arrange the blankets on the 12 mm dia. pipework system; due to the small radial dimensions and due to junctions and deflections at the manifold and header, it was arranged in a different way. Here the first MLI blanket a r r a n g e m e n t was a mixture between the mentioned one on the vertical tubes and a helical a r r a n g e m e n t of 200 mm wide 12 layer blankets on the manifold and header. INTERPRETATION OF EXPERIMENTAL RESULTS AND CONCLUSIONS It has been demonstrated with the industrial MLI J E H I E R IR 300.12 t h a t heavy deterioration of the quality of thermal insulation with increasing curvature of the surface to be insulated, i.e. with decreasing diameters of cylinders and lines, can be avoided (Fig. 3). The advantage of blanket type MLI was indicated in THISTA under LN2 and LHe cold surface conditions. Considering convenient handling and the compact design of this blanket MLI, the quality of insulation is very satisfactory. This allows also LHe transfer lines to be manufactured at favorable costs by omission of a cryoshield cooled to 80 K, at least for short sections and laboratory scale applications. In our opinion, the remaining deviation of the measured MLI quality from the optimum values published in [3~5] for d -- 100 mm is generally due to the easy and moderate cost handling of this commercial MLI and, more specifically, to the following reasons: 9 No use of any machines and high technology conditions as in aerospace applications e.g. 9 No ideal plane or cylinder perimeter surface (bottoms, deflections). 9 No "double guarded cylinder technique" as in [3, 4]. 9 No isothermal single overlappings or structured insulation arrangements. 9 No "processing", i.e. no baking out, no pumps integrated in form of carbon paper e.g. [3]. Considering these boundary conditions, a typical factor f relative to the MLI-blanket calculated under "quasi-ideal conditions" (Tw = 293 K, TK = 77 K, ~ = 0.02, Presid.gasN2 -10-5 mbar) of f - ~ I m e a s u r e d / C l q u a s i i d e a l ~ 1,7 for a 3 blankets-MLI on d - 219 mm can be valued quite acceptable. Degradation due to the deterioration of the vacuum is relatively little with the technique of insulation applied, namely -- 50 % at 10-4 mbar (Fig. 4). The influence of the warm boundary surface temperature on the insulation quality Cl is well understandable: approx. ~t ~ Tw 4.s, including radiation and conduction (Fig. 5). The first experimental result with the not yet optimized MLI blanket technique on the 12 mm dia tube a r r a n g e m e n t indicates t h a t also for complex and small dia transfer lines improved MLI qualities are achievable. ACKNOWLEDGMENT We would like to thank Dr. S. Jacob, Central Cryogenic Facility, Indian Institute of Science, Bangalore, for the valuable proposals he made to the concept of the THISTA cryostat test configurations during his stay as a guest scientist at FZK. REFERENCES 9
0
Barth, W. et al., Test results for a high quality industrial superinsulation, Cryogenics (1988) 28 607-610. Barth, W, Lehmann, W., Experimental investigations of superinsulation equipped with carbon paper, Cryogenics (1988) 28 317-320. Scurlock, G. and Saull, B, Development of multilayer insulation with thermal conductivities below 0.1 pW/cmK, Cryogenics (1976) May 303-311. Ohmori, T. et al., Multilayer insulation with aluminized dimpled polyester film, Proc. ICEC 11 (1986)567-571. Nast, T.C. et al., Thermal performance of tank applied multilayer insulation at low boundary temperatures, Proc. ICEC 10 (1984) 586-593.
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40 ~ I empiri )irical values for different transferlines and cryostats H with 20 2 - 50 layers of MLI between RT and 80 K (4 K) 35 [~-(windi ng technique on tubes and cylinders) i ....
9L
't
1
-, ~
l
-q---~ ' ;
200
-t-------t------r, I !
i
300 400 500 600 diameter of tube or cylinder [mini
H
'
-3-I
~
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q [ W / m ' ) l r d ::::....... if-[-------. o9 2 0 L_l. . . . . . ~ . . . . . . -~. . . . . . . ", W/m z ] " =v~-"~ .____~ ___: i.', it ~ I 15 It-;.... .q[ ~/ I - - cl[W/m2]lR 300.12 (3 blankets) _~1 > TESSI, between RT and 80 K
0[-1- ; .....,:i 0 100
'\l
I 81-m
'-~ 5 [
3 F--
J
o
0 700 =
'
1
'
I
'
Twarm = 2 8 0 K ...... 6 p=
-
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'
I
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]
onlyonealuminium foil(LN2) -
~ [W/m'] of Fig.l, MLI winding technique
L~/L--i
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open / closed symbols LHe/ LN2- experiments
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.
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~!
~1 ,
0
100 105
,
150
-
I ~,
200 250 300 350 219 320 TESSI *) *) TESSI results extrapolated from T ~,arr~,= 295 K to 280 K d [ mm ]
Fig. 1 Quality of MLI cl [ W/m z ], Clr [ W/m ] versus diameter d=dcoad [mm] (superinsulated tubes and cylinders, respectively)
\1
l\
4L .....
!
-
Fig.3 Quality of IR 300.12 MLI blankets versus diameter d=dcotd [mm] for different blanket numbers and cryogens
'"
l,o 12
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.
.
.
.
.
--[3- THISTA, d=54, LN2, 1 blanket _ 9 THISTA, d=219, LHe, 2 blankets --X
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9 TESSI, d=320, LN2, 24 layers - - - - Metaplast-reflectors and Interglas-Spacers, ____ unperforated (winding technique) . .
o" 9 8
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~ t~ ...... t
[
',7 . . . . . . . .
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=
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3
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12mm dia rest modute arr'angemenr
0 1( )-7
1
10 6
105
10 4
10 3 10-2 p [mbar]
10-1
Fig. 4 Quality of MLI's versus residual gas pressure p Comparison of IR 300.12 blanket MLI with MLI in winding-technique
1,2 THISTA T 1,0 ..... 219 mm dia test module | 3 x 12 layers / 3 blankets ~, 0,8 ....... T cold = 80 K (LN2 cooling) E p < 10-6 mbar 06 9"-expenmental ' ...... ' data A
=
L
. . . . . . . . . e. . .1.5.0.0 . . . . . . . . . . . . . .
j
'
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1 vacuum tank 2 test (measurement) chamber 3 intermediate guard chamber 4 external LN2 guard 5 guard shield 6 multiinterface flange 7 cryostat lid 8 support plate 9 sintered element a, b, c guard shield thermally insulated, LN2-cooled, heated Fig. 2 THISTA Test facility for thermal insulation
~
0,O50
;
r------
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I / ----/-
. . .-4.... ~ ; ~
"- (T warm ) ~' ~ correlation
+
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--4 ........... --~/-// .......
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Twarm [ K ] Fig. 5 Quality of IR 300.12 MLI versus warm boundary temperature on the 450 mm dia guard shield
=
Design and Construction of Long Cryogenic Piping Lines
Katsumi Kawano, Kazuya Hamada, Takashi Kato, Tadaaki Honda, Kazuhiko Nishida, Kunihiro Matsui, Yadao Hiyama, Kiichi Ohtsu, Shuuichi Sekiguchi, Hiroshi Tsuji, Mamoru Ando*, Tetsuo Hiyama* and Kaoru. Ichige* Japan Atomic Energy Research Institute, Naka Fusion Research Establishment, 801-1, Muko-yama, Nakamachi, Naka-gun, Ibaraki, 311-01, Japan, *Hitachi Oxygen, Co., LTD., 3-1-17, Kokubun-cho, Hitachi-shi, Ibaraki, 316, Japan
An unique heat-leak measurement method has been developed, being useful to determine the heat leak for a long-length cryogenic piping line. The method is only to measure the outer surface temperature of the cryogenic piping line when liquid helium flowing. The method has been verified that the measurement result by the method had a good agreement with the other measurement method such as a liquid helium transfer efficiency measuring method and a enthalpy mass flow rate measurement method. The method provides a simple verification method for cryogenic piping line performance.
INTRODUCTION It is generally complicate and delicate to determine a heat leak of a cryogenic piping line such as liquid helium transfer line and a cold helium transfer line. The heat leak is ordinary evaluated to measure liquid helium transfer coefficient, or mass flow rate and enthalpy difference through the piping line. However, it will be almost impossible to determine the heat leak in the case that there aren't any sensors to measure the inlet and the outlet condition of transferring helium. Japan Atomic Energy Research Institute (JAERI) is the Japanese representative research institute of International Thermonuclear Experimental Reactor (ITER) and has the charge of fabrication of ITER Central Solenoid (CS) model coil[1 ], construction of CS model coil common test facility (CSTF)[2], and implementation of the CS model coil testing. The CSTF has been constructed, involving the development of a large cryogenic system[3]. The construction of the large cryogenic system required some of a new long-length cryogenic piping line to transfer cold helium from the new cryogenic system to the existing cryogenic system which is the test facility for the ITER full size superconducting conductor and its joint. The longest cryogenic line is nearly 90 m and the others are around 50 to 60 m, respectively. Low heat leak is specified to be around 1 W/m including all the heat leak from the joint and support parts of the cryogenic piping line. One line, however, is not available to measure the liquid helium transfer coefficient or the enthalpy difference. Therefore, we has tried to develop a new measurement method that can determined the heat leak of the piping without any sensors at the inlet and the outlet of the piping. The new measurement method is presented in the paper.
DEVELOPED MEASUREMENT METHOD Recent cryogenic piping has a relatively small outer tube diameter due to use of high performance thermal insulator. For instance, around 60-mm diameter outer tube is used for a 20-mm inner tube diameter that is applied for a typical liquid helium transfer piping line. To consider the heat leak from the air to the surface of the outer tube, the heat leak (q) is easily calculated as the following equation, assuming free convection heat transfer coefficient (h); 493
494
ICEC16/ICMC Proceedings Flow Meter
),)
ATM
LN 2
~C-" / 2400
w
O :Measurement Point
gC
\
\1
Figure 1 Heat transfer coefficient measurement in factory
q = h AT S L
(1)
Where A T is the temperature difference between the outer tube surface and the ambient atmosphere, S and L are the outer tube area with unit length and the length of the outer tube, respectively. The free convection heat transfer coefficient is generally in the range of 5- 25 W/m2.K. It will be around 5-6 W/m2"K at static indoor condition. The heat leak is tried to calculate by substituting the 60mm outer-diameter tube and the free convection heat transfer coefficient of 5 W/m2"K. In the case that the heat leak of unit length is assumed to be 1 W/m, the A T is determined to be around 1 K. If the accuracy of the temperature difference measurement is less than 1 K, it can measure the heat leak of around 1 W/m for the cryogenic piping. When the h is evaluated in advance, the heat leak of the piping can be easily determined by measuring the temperature difference between the surface of the outer tube and the ambient atmosphere. Determination of free convection heat transfer coefficient To determine the free convection heat transfer coefficient in advance, an experiment was performed in indoor condition by using the practical cryogenic piping as shown in Fig. 1. Liquid nitrogen was stored in the inner tube of the piping and boil-off rate was measured after cooling the piping down completely. The heat leak was calculated from the obtained boil-off rate. Simultaneously, the surface temperature and the ambient atmosphere temperature were measured at each point of the piping as indicated in the figure. Combining both measurement results, the free convection heat transfer coefficient was determined to be 5.5 W/mZ'K. This value was used to calculate the heat leak of the newly constructed long-length cryogenic piping lines. NEW CRYOGENIC PIPE LINES Three long-length cryogenic piping lines have been constructed and installed to communicate cold helium between the CSTF cryogenic system and the ITER full size conductor test facility as shown in Fig. 2 Their specification is summarized in Table 1. Each piping line was designed and fabricated to achieve the overall heat leak of less than 1 W/K including all the inner tube support and the bayonet joints.
ICEC16/ICMC Proceedings
495
IC~
' Ef Co,d He,ium etum L,no
, ~
~
J,!:
/1
0 34.0 X tl.65 60m
~l~e~,',
]
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ress~
~
'
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i
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Cold Helium Supply Line 0 21.7 •
90m [Cold-Box
Liquid Helium Supply Line ~521.7 • t 1.65 52m
ZX -----W-_ -
20,000L LIQUID HELIUM TANK
ITER Full Size Conductor Test Facility
Figure 2
Table 1
Cryogenic piping between cold box and existing test facility
Specification of new piping line Cold helium supply line
Length (m) Inner Diameter (mm) Outer Diameter Number of Joint Target value(W/m)
86 21.7 60.5 12 0.735
Liquid Helium Supply line 52 21.7 60.5 7 0.97
Cold helium return line 60 34 9 1.03
HEAT LEAK MEASUREMENT RESULTS On the liquid helium supply line, the heat leak was estimated by using two kinds of method such as the liquid helium transfer method and the developed surface temperature measuring method. This line is connected with the 58-m existing line of which heat leak of 3.42 W/m was already determined. When liquid helium was transferred, the liquid helium storage rate of 332 liter/h and the liquid helium withdraw rate of 676 liter/h were measured, respectively. The heat leak was calculated to be 0.98 W/m, considering the heat leak from the existing line. The surface temperature measurement method was performed. The 51 points were measured as shown in Fig. 3, according to one point every meter. The measurement results are shown in Fig. 3. Some of peek are observed in the figure such as the measuring point numbers of 12, 16, 25, 36, and 42, which corresponds to the location of piping joints and flexible tubes. The average temperature difference is calculated to be 0.82 K except the data for the joints and the flexible tubes. On the contrary, the average temperature difference for the joints and the tubes are 1.3 K. The total averaged heat leak is finally determined to be 0.97 W/m that has almost the same as the result of the liquid helium transfer method. Therefore, the developed heat leak measurement method has been verified to be useful to estimate the heat leak. On the cold helium supply line, the surface temperature measurement and the enthalpy deference measurement through the line with mass flow rate measurement were performed. As a result, the heat leak of 0.53 W/m is estimated by the surface temperature measurement method and the heat leak of 0.59 W/m is measured by the enthalpy measurement.
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ICEC16/ICMC Proceedings
The cold helium return line didn't have any sensor to measure the helium inlet and the outlet condition. Accordingly, only the surface temperature measurement method was available. The measured results was 1.026 W/m. All measurement results are listed in Table 2. It is verified that our developed surface temperature measurement method is significantly useful. Table 2
Summary of heat leak measurement
Design (W/m) Surface temperature measurement (W/m) Liquid helium transfer (W/m) Enthalpy method
~L_
High pressure cold helium gas line 0.735 0.527
Liquid helium line 0.97 0.98 0.98 -
-
0.59
Low pressure Recovery line 1.2 1.03 -
302
14.0
300
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298
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i
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296
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294
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Measurement Position (longitudinal)
Figure 3
Temperature profile of the surface temperature measurement
CONCLUSION Heat leak measurement for the long-length cryogenic piping line has been successfully carried out and the performance of cryogenic piping line is satisfied with the design requirement. Using free convection heat transfer coefficient, heat leak can be estimated from the surface temperature of the piping line. The results has good agreements with the both of liquid helium transfer efficiency measurement and enthalpy massflow measurement. New and useful heat leak measurement method for the cryogenic piping line has been developed and established. ACKNOWLEDGMENT The authors would like to thank Drs. M. Ohta, T. Nagashima and S. Matsuda for their continuous encouragement and support on this work. The fabrication work from JECC Corporation, is well acknowledged. REFERENCE
1 2 3
Thome. R., Design & Development of the ITER Magnet System,. Cryogenics (1994) Vol. 34 ICEC Supplement 39-45 Shimamoto, S. and Hamada, K. et al., Construction of ITER Common Test Facility for CS model coil, Proceeding of 14th International Conference of Magnet Technology 1995 Kato, T. et al., Cryogenic system for ITER CS model coil, to be published in Advanced in Cryogenic Engineering (1994) 41
Calculating Radiation Exchange Factors of Radiant Cryocooler by the Monte Carlo Method
Hai Yao,*
Chunhua Gao,**
Chunzheng Chen*
and
Yanzhong Li*
*Institute of Cryogenic Engineering, Xi'an Jiaotong University, Xi'an 710049, China **Institute of Systems Engineering, Xi' an Jiaotong University, Xi' an 710049, China
A computer program, intended for engineering calculation of radiation exchange factors of radiant cryocooler in real enclosure with diffuse emission and specular reflection surfaces, is described. The method is a simple Monte Carlo application, and the flow chart of computer program is provided. The radiant exchange factors of a two-stage radiant cryocooler have been calculated. The results are discussed and compared with the data obtained by the ordinary specular image method. Comparison shows that the Monte Carlo method has equal or better effect in simple cases and strong potential in complex cases. So, it is very useful to the designers of radiant cryocoolers.
INTRODUCTION Spaceborne infrared imaging systems and infrared spectrometers are finding applications in the fields of earth resources, worldwide meteorological data gathering, and air and water pollution monitoring. These space sensors utilize a detector/cooler assembly which is usually designed to be sensitive in the mid-infrared region of the electromagnetic spectrum. Because of its passive operation and inherent long-term reliability, the radiant cooler offers numerous operational advantages over relatively heavy stored-cryogen coolers and power-consuming closed-cycle coolers. To date, the most widely used system for obtaining cryogenic temperature aboard spacecraft have been radiant cooler in China. The passive radiant cooler rejects heat by radiation to the low effective temperature of deep space (less than 4K). So, the radiative heat interchange calculation is very important in the design of radiant cooler. Especially, the calculation of radiation exchange factors (REF) is crucial in the calculation of radiative heat transfer. In the past, the REF of radiant cooler was calculated by specular image method [ 1]. This presentation gives a new method based on statistical theory known as Monte Carlo and illustrates how to use this method to calculate the REF of radiant cooler. There are many advantages of this method compared with the specular image method: (1)The Monte Carlo method can easily calculate the REF of surfaces which have complex geometrical configuration, such as the radiative heat transfer between the radiant cooler and the earth. (2)The real surfaces can be accurately simulated by the Monte Carlo method. (3)The radiation between curved surfaces can be calculated by the Monte Carlo method which is difficult for specular image method. MONTE CARLO METHOD 497
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The Monte Carlo method is in essence a statistical method which can be simply interpreted in terms of physical processes. It is conceptually very simple and particularly well-adopted to deal with the radiative heat interchange problems having complex geometrical configuration and direction dependent radiation characteristics of surfaces. In fact, the method takes the Lagrangian point of view to describe the physical process of radiative heat transfer, while the commonly used integral equation method takes the Eulerian point of view to consider it. The method has been applied to radiative heat interchange problems by a number of investigators[2,3]. The purpose of this paper is to examine radiant heat transfer process of radiant coolers in real enclosure with diffiase emission and specular reflection surfaces by Monte Carlo method. The procedure of Monte Carlo numerical simulation is described below. First, generate emitting point on surface. If the surface has even temperature and radiative characteristics, the weighted mean of area should be taken account of. For instance, area A~ lies in the range between yl(x) and y2(x), where x e[xl,x2]. To generate emitting point on Ai, first pickup two random numbers R~,R2 ~ [0,1 ], and let
(l) Y = Ymi~ + (Yma• - Ymin ) • R:
Then, test the inequation Yl (x) < y < Y2(x)- If satisfied, it means that the point is on A~, else repeat the above procedure. Second, generate emitting energy buddle's direction. The emitting energy buddle's direction (0, ~p) in spherical coordinate is determined by random numbers R3,R4. For the case of d i ~ s e emission, we have sin 0 = ~
(p = 2rcR4
(2)
Third, Trace the energy buddle and determine if it is absorbed or reflected. If the position of emitting point and the emitting direction of energy buddle as well as structure of the enclosure were known, the destination surface Sj and the coordinate of intersecting point can be calculated. Then we can determine randomly whether the energy buddle be absorbed or not. A random number R5 is generated, fiR5 is smaller than the absorptivity of surface Sj, the energy buddle is absorbed, otherwise it is reflected. If the destination surface is specular, the direction of reflection can be calculated by vector product formulation. If the destination surface is not specular, the hybrid specular-diflhse model can be used which will not be discussed ha this paper. At last, keep tracing energy buddles and count the number of absorbed buddles. Let surface i emits Ni energy buddies, and follow the trace of each buddle until it is absorbed. If N~j represents the number of buddies absorbed at surface j, Eij=Nij/Ni is the required REF (the energy fraction that leaves surface i and reaches surface j). CALCULATION OF REFs FOR RADIANT COOLER USING MONTE CARLO METHOD As shown hi figure 1, a two-stage radiant cooler is studied in this paper. Both of the two cones are equivalent to enclosures with five flat surfaces as shown in figure 2. 111 calculation of radiative heat interchange for radiant cooler, the REF of the second stage radiator to cone end and the REF of the first stage radiator to shield end are equivalent to the REF of surface 1 to surface 5. A computer program was developed using the Monte Carlo method and the flow chart is given in figure 3.
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ICEC16/ICMC Proceedings CALCULATING RESULTS AND ANALYSIS
The geometrical parameters and radiation characteristics of surfaces about the second stage are given in table 1. According to these data, the REFs of the second radiator to the cone end have been calculated by
.....
Lm Lc
r ~surface 3
J 2nd stage / t " / ,-'i surface 5
I1 st stage
Two-stage radiant cooler
Figure 2
~.-/ X
Equivalent enclosure
I input parameters of radiator I
~,
I initialize counters: num=0,suc=0
.... ~
] num=num+l I generate emitting point and direction from surface 1
I calculate the arriving surface order k and ._ the intersecting point coordinant 5 ~ five cases of k
I randomly II calculate reflection ]generate direction ]/ direction J ] output calculating results of REF I ~.,.
~, .... end Flow chart of computer program
"~
I
surface 1
~, ,.begin5
Figure 3
//,surface 2
/:5/ ""
surface4
Lp
Figure 1
Z
Y
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ICEC16/ICMC Proceedings
the above computer program. The results with different sample size N are listed in table 2, where E,E0,E~,E2 and E3 represent the overall exchange factor, the exchange factor without specular reflection, the exchange factor with one time specular reflection, the exchange factor with two times specular reflection and the exchange factor with three times specular reflection respectively. In this program, pseudo-random numbers are generated by congmential generator, the cycle of which is 6553 5. IfN exceeds the cycle limit, the random series generated will show a duplicative behaviour and cause the increase of system error. So, we take the calculated exchange factors with N=50,000 as the final result. Table 1
Parameters of the 2nd sta e
...............................................
_g--;--
.....;-_.;.:...
-; ......
_-. -_-;
.:yab!e:2 ...... Ca!.c:u!a:t!n
_ge_gme_t_r!_c__parameter_s.................a_b_so__rptivit_!e_s........................ Wp 32 el 0.98 Lc 98 e2 0.04 Lp 204 e3 0.04 Lm 245 e4 0.04 e5 1.00
N 10000 20000 30000 40000 50000
E 0.9790 0.9803 0.9803 0.9797 0.9795
r__es_u_!ts__of
E0 0.5337 0.5396 0.5432 0.5412 0.5415
E1 0.3924 0.3885 0.3850 0.3870 0.3865
Vs ....
E2 0.0529 0.0522 0.0521 0.0515 0.0515
E3 0.0000 0.0000 0.0000 0.0000 0.0000
In table 3, comparison between our results and that of specular image method [4] is displayed. It can be seen that the results of the two methods have a good agreement. So, it is proved that the program based on Monte Carlo method has high precision and reliability. Table 3
Results comparison between the two methods
.................................. E0 E2 ~
Sp.ecularimag e meth_0__d_................... Mo_n__te_Car!0_method (N-50,000) .... 0.5495 0.5415 0.0530 0.0515
CONCLUSION Compared with the ordinary specular image method, the Monte Carlo method used to calculate the REF of radiant cooler is conceptually simple yet have considerable precision. Furthermore, the application of this method can be extended to solve more complex cases. So, it is very useful to the designers of the radiant cooler. REFERENCES Annable, R.V., Radiant cooling, Appl Optics, (1970) 9 185 Howell, J.R., Application of Monte Carlo to heat transfer problems, Advances in Heat Transfer (1968) 5 1-54 Toor, J.S., A numerical experiment of radiant heat exchange by the Monte Carlo method, Int J Heat Mass Transfer (1968) 11 883-897 Han,J., Integral solution for radiative coupling view factors of space radiant coolers, .Chinese Space Science and Techn..ology (I 988) 2 18-24 (in Chinese)
Development of Loop Heat Pipes for Cryogenic Applications Rohana Chandratilleke, Yasumi Ohtani, Hideo Hatakeyama, Hideki Nakagome Energy and Mechanical Research Laboratories, Research and Development Center Toshiba Corporation, 4-1, Ukishima-cho, Kawasaki-ku, Kawasaki, 210, Japan Since loop heat pipe was first invented in 1987 by a Japanese inventor, some efforts have been made to extend it for cryogenic use. In 1991, a loop heat pipe had been demonstrated to work at liquid nitrogen temperature levels. The present work has taken these efforts still further. This work shows that loop heat pipes can be put into operation at any temperature level fight down to 4 K level. To distinguish it from a capillary-pumped loop, which uses a wick, we call it a loop heat pipe, which is just a pipe coiled in a closed loop configuration.
INTRODUCTION With advancements in cryocooler-based applications such as cryocooler-cooled superconductin~,~, magnets [1], means will have to be developed for effective heat transport between components to be cooled and cryocoolers. Conventional thermal couplings such as copper are being challenged by new developments in heat pipe technology[2, 3]. Lately, two types of room-temperature heat pipes are emerging as cryogenic heat pipes. They are called loop capillary heat pipes[4] and capillary-pumped loops[3]. First invented in 1987 by Akaji [5] for room temperature use, the loop heat pipe had been demonstrated to work with liquid nitrogen in 1991 by Kawai et al.[4]. In 1992, they addressed the issue of choosing the amount of gas required for proper heat pipe functioning[6]. Having connected the cold end of the heat pipe to a cryocooler, Chandratilleke et al.[7] studied heat pipe performance for nitrogen over 60 to 90 K and also effects of heat pipe inclination on the performance. Basically, a loop heat pipe has a configuration as shown in Figure 1, where a single pipe is shown Heating block [-" . . . . . ~
I l
Cooling block
1 .........
U...................
t] ........
Vapor bubble
i
i
I
I
Closed loop of pipe
Figure I Schematic diagram of loop heat pipe; Inset shows a vapor bubble in heated section coiled to form a close loop. Two heat transfer plates are soldered to the turns. When the pipe is filled with a liquid at boiling point, heat will be transported from one plate to the other. Use of a heat pipe with a cryocooler enables a smaller temperature difference between the cryocooler and object to be cooled. This implies that, apart from being able to operate at a higher temperature for better cryocooler performance, the cryocooler can be kept far away from the object to be cooled. In a cryocooler-cooled magnet system, this means that a crycooler carrying a magnetic regenerative material such as E~Ni can be operated at less interferences from the magnet. The present work addresses the problem of developing loop heat pipes for temperatures below 77 K with the final aim of developing a 1 W class loop heat pipe for 4 K use. 501
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A CONDITION FOR HEAT TRANSPORT
Even though no visual observation has been made to clarify heat transport mechanism of the loop, it is postulated here that vapor bubbles generated in heated portion of turns are responsible for liquid circulation and thus heat transport. For a vapor bubble to drive liquid ahead of it, it is suggested here that the pipe diameter be slightly larger than the vapor bubble diameter. Then, as shown in Figure 1, a thin liquid film is formed between the bubble and pipe, and its evaporation will generate the required steady pumping force when a bubble passes the heated zone. Even though an exact bubble size is difficult to determine, it is suggested here using as the vapor bubble size Laplace constant L, defined by the following equation.
L=I, o"
(1)
(P 1 - p v)g where, cy, 9~, 9~ and g are surface tension, liquid and vapor densities and acceleration of gravity respectively. Evidence of a thin film formation can be found in heat transfer improvement as reported by Klimenko[8]. According to Klimenko[8], heat transfer coefficient increases as a result of the thin film evaporation when the pipe diameter is about 1.5 times the Laplace constant for vertical pipe orientation. We can expect that similar conditions would generate a pumping power should it ever to be generated. Laplace constants for several cryogens are shown in Figure 2, which can be used to obtain heat pipe diameters. 2.0
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Temperature (K) Figure 2 Comparison of Laplace constants for several cryogens E X P E R I M E N T A L APPARATUS Figure 4 shows a schematic diagram of the apparatus used for testing heat pipe performance. One heat transfer plate of the heat pipe is connected to a cryocooler and the other to a heater. The cryocooler was a two stage Gifford-McMahon wherever necessary, otherwise it was a single stage one. The heat pipe assembly was placed in a vacuum vessel, which was rotatable. The heat pipe was filled with an appropriate gas using the metering tank so that liquefied gas would occupy a pre-determined volume fraction of the heat pipe. The plate temperatures were measured with carbon-glass resistance temperature sensors. Pressure in the heat pipe is measured with a pressure transducer connected to the room temperature filling line. RESULTS Figure 4 shows limiting heat transport for loop heat pipes using nitrogen, neon and hydrogen as working fluids. Heat pipe diameters were determined by Laplace constant of the relevant fluid. For instance, nitrogen heat pipe had 2 mm inner diameter, while neon's was 1 mm and the hydrogen's 3 ram. Though the nitrogen heat pipe worked at horizontal orientation, the other two did not. This can be attributed to
ICEC16/ICMC Proceedings Cryocooler
Fill line
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Figure 3 Heat pipe performance testing apparatus ready dry-out of heat pipes that use hydrogen and neon, whose liquid-to-gas density ratios are lower than nitrogen' s. However, the dry-out could be avoided by slightly inclining the heat pipe, in which case there will be a reverse liquid flow in the thin film of liquid mentioned earlier. In terms of equivalent solid conductor performance, all three of the heat pipes fared well, transporting at least 10 times more heat. 3O 25 0 ffl
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Mean Temperature (K)
Figure 4 Limiting heat transport for hydrogen, neon and nitrogen A helium loop heat pipe having 0.51 mm internal diameter was designed and tested for performance. See Figure 5 to get an idea of the present helium and hydrogen loop heat pipes. Details of the helium heat pipe are as follows: Half peripheral pipe length between heating and cooling blocks is 516 mm, heated length of a single turn is 30 mm and the number of turns is 10. Figure 6 shows limiting heat transport of the loop heat pipe as a function of inclination of heat pipe plane to the horizontal when the heating blocks are parallel and twisted. When the blocks are parallel, heated lengths of pipes are horizontal and therefore easy for them to go dry. The figure shows that, at small inclinations, twisting can increase the limiting transport by about 70%. In an effort to further increase the heat transport, experiments were performed with the heat pipe plane and heat transfer blocks vertical. In this case, the heat pipe did not work and this can be interpreted as evidence that the heat pipe is not a thermosyphon, but it is bubble driven. At the moment, the limiting heat transport of the helium heat pipe is not large enough for a magnet application, but it will be possible to increase this value to 1 W level by increasing the number of turns. CONCLUSIONS Having selected the loop pipe diameters close to 1.5 to 2 times Laplace constant, it was demonstrated that a loop heat pipe can function even at cryogenic temperatures as low as 30 K, 15 K and 4 K. Further optimizing is necessary for the helium loop heat pipe to have a higher heat transport capacity.
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Kuriyama, T. et al., Cryocooler directly cooled 6 T NbTi superconducting magnet with 180 mm room temperature bore Cryogenics ICEC Supplement (1994) 34 643-646 Rosenfeld, J. H., Wolf, D. A. and Buchko, M. T., Emerging technologies for cryocooler interfaces Cryocoolers - 8 Plenum Press, New York (1995) 743-753 Cullimore, B., Kroliczek, E. and Ku, J., Cryogenic capillarypumped loops" A novel cryocooler integration technology Cryocoolers - 8 Plenum Press, New York (1995) 719-730 Kawai, N., Ono, M., Sanada, Y. and Akaji, H., Development of cryogenic heat pipes Proceedings of 45th Meetin~ on Cryogenics and Superconductivity (1991) 47 (Japanese) Akaji, H., Loop capillary heat pipe Japanese Patent (1994) No. Hei 6-97147 (Japanese) Kawai, N., Ono, M., Sanada, Y. and Akaji, H., Development of cryogenic heat pipes (2) Proceedings of 48th Meeting on Cryogenics and Superconductivity (1992) 275 (Japanese) Chandratilleke, G. R., Ohtani, Y., Hatakeyama, H., and Nakagome, H., Development of looped heat pipes for cryogenic applications Proceeding of International Sessions 73rd JSME Annual Meeting Nihon University, Tokyo (1996) (VI) 140-143 Klimenko, V. V., Fyodorov, M. V., and Fomichoyov, Yu. A., Channel orientation and geometry influence on heat transfer with two-phase forced flow of nitrogen, Cryogenics (1989) 29 31-36
N e w cryosurgical probe suitable for endoscopic application in the minimal invasive therapy Bernhardine Schumann, Armin Binneberg, RalfHerzog Institut far Lull- und K~ltetechnik gGmbH, FB Kryotechnik, Bertholt-Brecht-Allee20, D-01309 Dresden, Germany Object in view of the Minimgl Invasive Therapy is to substitute the traditional open and therefore invasive surgically interventions through fewer invasive surgery. In addition to microsurgery and lasertherapy, which are mainly used in MIT, cryosurgery may also be used. Applications of endoscopic cryotips are possible in several medical disciplines e. g. in the gynecology. This paper presents a liquid-nitrogen-cooled cryotip for use in gynecology. The cryogenic efficiency of the cryotip was proved by laboratory tests. There are presented the cryotechnical parameters, which are important for the medical therapy, such as freezing rate in dependence of the distance, isotherms, cooling power etc.
INTRODUCTION Cryosurgery is understood to be the applications of low temperatures down to 78 K (boiling point of liquid nitrogen) to attain therapy effects in various medical disciplines. Cryosurgery now is significant in the therapy of tumours of definite character, localisations and extend. The basic feature of cryosurgery is to devitalize the tissue by freezing in situ. By a defined regime of freezing and warming the tissue, the formation of extra- and intracellular crystallites destroys the cell membranes of cancerous tissue. In general cryosurgery includes favourable functional and cosmetic results. Cryosurgery started first of all in the field of Ophthalmology, later Dermatology, Oto-Rhino-Laringology, Urology, Proctology, Gynecology, Neurology and others had been opened up. The use of liquidnitrogen-devices is well established in clinical practice. Two modifications exist: - In spray freezing liquid nitrogen is sprayed directly onto the tissue to be treated by the use of an appropriate device, leading to optimum freezing rates and refrigerating capacities due to the intimate contact of the coolant with the irregular tissue surface. - In contact freezing the heat is removed from the tissue by the metallic working area of a closed cryoapplicator. The working area can be put on the tissue either at room temperature before starting the cool down, or the cool down to working temperature can be done before contacting the tissue. Vacuum insulated and LNz-flooded cryoprobes with only 2.6 mm diameter have been constructed at the Institut flit Lull- und K~,ltetechnik permitting percutaneous applications in neurosurgery and dental-facial surgery in case of trigemical neuralgia [ 1]. Cryotips for endoscopic use have been constructed [2]. Cryosurgical probe The cryosurgical probe has the following geometric dimensions: shall length 300 mm, shaft diameter 4.5 mm. The shaft is thermally insulated by vacuum. The cold available for therapy is provided locally at the front end of the tip within an open cycle (Fig. 1). The cryotip is supplied by means of a flexible tube system with liquid nitrogen (working pressure: 4 bar). The therapy time is adjusted by electronic control. 505
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ICEC16/ICMC Proceedings 700
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Thermosiphon Cooler: A Low Microphonic Cooling System for HTC-Devices; Especially for SQUIDs Armin Binneberg, Hanno Buschmann, Ralf Herzog, Johannes Neubert, Gabriele SpOrl Institut ~ r Lufi- und K~iltetechnik Dresden, gGmbH, FB Kryotechnik, Bertolt-Brecht-Allee 20, D-01309 Dresden, Germany Applications of HTC-devices in high-tech areas require a potentially low microphonic (or free) cooling system. Thermal fluctuations, mechanical vibrations and electromagnetic influences are to be minimized (or equalized). A hybrid cooling system was developed consisting of a small split-Stifling cryocooler as refrigeration source and a thermosiphon. The thermosiphon (condenser and evaporator, separated by capillary tubes) is filled with gaseous nitrogen and closed. The condenser is contacted to the cold head of the Stifling machine. The SQUID is mounted on the evaporator and is not affected by disturbances of machine. A cooling power of 0.3 W is reached at 65 K.
INTRODUCTION Use of HTC-devices in high-frequency techniques or other fields of measuring techniques demands special cooling systems adapted to the corresponding measuring tasks. The cooling system has to meet the following requirements: Temperature range for SQUID cooling: < 70 K Nearly no mechanical vibrations Cooling capacity more than 0.1 W
No use of cryogenic liquids Damped electromagnetic noise
All in all the cooling system has to be a compact and portable one. For realization of these points a split Stifling cryocooler from AEG Infrarot-Module GmbH, Germany was selected, having 1 W at 80 K cooling capacity. The thermosiphon was developed for damping and/or separating the influences of mechanical and electromagnetic nature. Furthermore active and passive damping elements were installed. [ 1] EXPERIMENTALS Methods of Compensation of Disturbances Disturbances arising from the mode of operation of the cooler (mechanical motions of parts and eleetromeehanieal signals caused by these motions) must be damped or equalized. So vibrations from the compressor and the cold head (from the split Stirling machine) are diminished by mounting all-metal vibration dampers in the base plate system (see Figure 1). Moreover, the thermosiphon is made as a ,,soft spring", that means the capillary tubes connecting condenser and evaporator are sot~ and formed as an elbow. Bellows are used as the vacuum jacket of the capillary tubes. The evaporator is placed within the sensor dewar by a three point support and the sensor dewar is fastened on the base plate by vibration dampers, additionally. In order to damp electromagnetic noise, sensor place and cold head are separated from each other. The distance between them is up to 300 mm. Condenser and evaporator are at different horizontal levels. Furthermore, electromagnetic shielding materials can be used either as shields or as constructing materials. 509
ICEC16/ICMC Proceedings
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COLD HEADOF SPLIT II COLDHEAD
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Figure 1 Scheme ofthermosiphon cooler Method of Test For testing the thermosiphon cooler temperatures at condenser and evaporator are measured during cooling down. One gets the relation between temperature and cooling time for a set of parameters. The following process variables can be changed as given: Pressure of nitrogen Set point of temperature Volume of the storage tank Diameter of capillary tubes
(0.2... 0.6) MPa. (65,70) K (200, 300, 400) ml (0.8... 1.8) mm.
During the measurements only one parameter was changed whereas the other ones remained still the same. The typical behaviour of the temperatures within the thermosiphon is shown in Figure 2. For valuation of the thermosiphon cooler the amplitude of oscillation was measured at the sensor place from time to time. [2;3] RESULTS AND DISCUSSION The most simple change of parameters was the variation of the volume of the storage tank. But only for a vessel volume of 200 ml the pressure variation doesn't influence the temperature at sensor place. For bigger vessels and pressures of about 0.2 MPa. there is too much nitrogen within the thermosiphon. The circulation of liquid nitrogen is unstable and therefore the final temperature at the sensor place varies with pressure. In Figure 2 the relationship between the cooling down and the diameter of capillary tubes as well as additional masses for simulating a heat load is represented. The diameter of the capillary tubes influences the oscillating properties of the thermosiphon. Enlarging the diameter also means an increase of mass. In that sense the so called ,,soft spring" becomes harder. In dependence of the position of that mass within the thermosiphon the vibrations from the cold head can be damped or increased. Cooling down of the evaporator becomes longer and the final temperature is enlarged. Additional masses simulate thermal heat load at sensor place as well as the cross over behaviour between the variation of diameter. Furthermore, one gets information about the quality of thermal isolation and thermal contacts between all elements of the thermosiphon. This part of investigation shows the complication of laying out such a cooling system for active and passive damping of disturbances from different sources.
ICEC16/ICMC Proceedings
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Figure 2 Relationship between cooling down, diameter variations and additional masses SUMMARY The analysis of the records obtained shows optimal working conditions for the thermosiphon when the diameter of the capillary tubes is about 1.2 mm. Then, the nitrogen storage tank should have a volume of 200 ml filled with nitrogen of a pressure of 0.4 MPa. Subject to the above mentioned qualifications the vibrations of the cold head (without external damping of 16 tam) are reduced to 1 lam in axial direction at sensor place. The net cooling capacity is up to 0.3 W at 65 K and up to 0.6 W at 80 K without changing of temperature. ACKNOWLEDGEMENT The authors thank Mrs. Karin v. Kronhelm from Chalmers Tekmiska H6gskola G6teborg and Mr. Ole Fredrich from Technische Universitat Dresden for their support upon a large scale during their diploma work. This work is supported by the German Bundesministerium ~ r Bildung, Wissenschat~, Forschung und Technologic within 13 N 62975. REFERENCES
.
.
Binneberg, A. and Neubert, J., Vorrichtung zur K0hlung von hochtemperatursupraleitenden mikroelektronischen Bauelementen, vorzugsweise Sensoren. Patentanmeldung, Akz. P 43 12 830.0, Deutsches Patentamt Mtmchen, 09.08.1994 Fredrich, O., StOrarme SQUID-KOhlung. Dip!0marbeit Nr. 380, Technische Universit~tt Dresden, Fakultat ~ r Maschinenwesen, Institut ~ r Luft- und Kaltetechnik Dresden, 04.07.1994 v. Kronhelm, K., Thermodynamische Untersuchungen an einem Mini-Thermosiphon zur KOhlung von SQUID-Sensoren. Diplomarbeit, Chamers Tekniska HOgskola - Schweden, Institut ~ r Lut~und Kaltetechnik Dresden, 16.06.1995
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What Happened to Cryogenic and Superconducting Equipment in the Great Hanshin Earthquake ? Final Report-
Akio Sato, Kohji Fujioka, Tomiyoshi Haruyama, Hiromi Hirabayashi, Takeshi Ishigohka, Atsushi Ishiyama, Yoshio Kawate, Kazu Nishigaki, Shigehiro Nishijima, Takashi Noguchi, Osamu Ogino, Hiroyasu Ogiwara and Osamu Okazaki Special Committee for Cryogenic Association of Japan Palais Dor's Hongo 302, 6-12-8 Hongo, Bunkyo-ku, Tokyo 113, Japan
On January 1995, a disastrous earthquake struck Hyogo and Osaka prefectures in western Japan. It was a type of shallow dislocation earthquake and resulted in huge damage to the area. There are many hospitals, laboratories and universities in this area, and various kind of cryogenic and superconducting equipment including 185 superconducting MRI magnets in operation. The Cryogenic Association of Japan immediately organized a special committee in order to find out exactly what happened to this equipment as a result of the earthquake and to make suggestions on what we have to do for safe operation. This report describes the investigation results on damage to MRI, NMR magnets, Cold evaporator, storage tanks and other cryogenic and superconducting equipment.
INTRODUCTION Early in the moming, 5:46 a.m. on 17 January, a disastrous earthquake hit the Kansai area in western Japan with a magnitude of 7.2 on the Richter scale killing over 6,000 and injuring 415,000. Awaji Island, the epicenter, and Kobe city recorded a 7th degree earthquake on the Japanese seismic intensity scale "Shindo". Shindo 7 corresponds to the 9th- 12th degree on the Marcalli seismic intensity scale (MMI). A lot of cryogenic and superconducting equipment is located in Kobe city and Osaka. A task force of the Cryogenic Association of Japan was organized for obtaining information in order to find out exactly what happened to this equipment as a result of the earthquake and to make suggestions on what we have to do for safe operation. The ad hoc committee made inquiries at hospitals in Hyogo and Osaka prefectures asking about the situation regarding MRI magnets. This inquiries were made under the auspices of Japan Society of Magnet Resonance in Medicine (JMRM). The investigation committee also collected information on cold evaporators in order to find out what happened to them. Damage reports on cryogenic equipment in the universities were offered by two universities in the area. The Kansai Branch, Refrigeration Commission and Applied superconductivity Research Group of the Cryogenic Association of Japan also cooperated with the committee [ 1].
DAMAGE TO CRYOGENIC AND SUPERCONDUCTING EQUIPMENT MRI Body Scanner Magnet The total number of MRI magnets in Hyogo and Osaka prefectures is 240 (185 superconducting magnets, 52 permanent magnets and 3 resistive magnets) as of September 1994 ( See Figure 1 )[2]. The investigation committee received 45 replies from hospitals coveting 46 MRI magnets ( 44 for superconducting and 2 for permanent ). No magnet quench was reported. Nearly half of the reported damages were classified as position shifts. In one case, an MRI magnet shifted about 80 cm (This case is not included in our statistics, because the information was obtained by direct inquiry to a person in a hospital). In some cases refrigerators were broken; cooling pipes for compressors were torn off; computer consoles moved, although the magnets themselves were not damaged. These cases are classified as 'other damage'. Figure 2 shows the relationship between damaged and undamaged magnets from 41 replies in total, classified by the magnet type. 513
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ICEC 16/ICMC Proceedings 140
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1T
1.5T
Permanent
Figure 2 Damaged magnets among the total 46 replies in total, classified by the magnet type. Nearly half of the reported damages were position shifts of the magnet.
The distance the MRI magnets moved varied from less than 1 cm to several tens of 20 cm. The relationship between the number 18 13 no position shift of position-shifted magnets and Japanese 16 m position shifted seismic intensity scale Shindo is shown in 14 Figure 3. In the cases of Shindo 2 to 4, no ~ 12 position shift was reported. In the cases -~ 10 where Shindo was over 4, six MRI magnets z 8 out of 31 shifted their positions. In all cases 6 where the MRI magnet moved anchor bolts 4 were not u s e d . For the other 25 magnets, 2 only six magnets were fixed by anchor bolts, 0 l--! . I-I ,1-1 , _ 2-3 3 3-4 4 4-5 5 5-6 6 6-7 7 Unknown and seven magnets were not fixed. In the Shindo remaining 12 cases, the existence of anchor bolts was not confirmed. It is hard to say from Figure 3 Relation between number ofposition-shitted magthe inquiry result whether fixing magnets to nets and Japanese seismic intensity scale Shindo. the floor by anchor bolts was effective against the earthquake or not. 16 E There is no distinct relationship beo 14 tween the distance the MRI magnets moved -o and Shindo as shown in Figure 4. The dis~ 12 E tance the magnets moved depended mainly ~ 10 i11 on the places where the MRI magnets were = 8 co installed and the method of fixing the magE 6 nets to the floor. It was reported that, in genAZ eral, damage was not severe for MRI mag~ 4 nets installed in basements, or- 2 co There was little damage to the MRI .~ ! magnets themselves, perhaps because these o 0 2-3 3 3-4 4 4-5 5 5-6 6 6-7 7 Unknown machines are designed and fabricated as a Shindo cold ship type, that can be transported in a Figure 4 Relation between the distance the MRI magnets cold state. Because the earthquake struck moved and Japanese seismic intensity scale Shindo. The early in the morning, there were no patients distance the magnets moved depends mainly upon the in MRI full body scanners. The situation places where the MRI magnets were installed and the could have been drastically different if dimethod of fixing the magnets to the floor. agnoses were going on of the time of the earthquake. After the earthquake, many MRI magnets were left in a dangerous energized state, because the number of electric sources for demagnetizing them was not sufficient in the Kobe area. Almost all MRI magnets have an Emergency Shut Down Unit, to be used in case of accidents. In this case of a disastrous earthquake, -
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however, we could not use the Emergency Shut Down Units because they need electric power in order to quench the magnets. Warming up a part of a cold NMR magnet by warm helium gas is a possible method to bring about a magnet quench. However, there is no safety manual which describes this method. Investigation into a safety emergency shut down method for MRI systems is needed. There exists no distinct standard for safe installation and housing of an MRI system including anchor bolts, exhaust line, refrigerator, cooling pipes and so on. In the case where the magnet moved 80 cm, the helium gas recovery pipe was not tom off because a flexible pipe was used. Many users in hospitals pointed out that the safety manuals for emergency cases should be reexamined and arranged for easier usage. NMR Magnet The investigation committee also made inquiries regarding NMR magnets to universities, companies and
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Figure 5 The type of the NMR magnets classified by Proton resonance frequency. The resonance frequency of resistive type of magnet is less than 100 MHz. For the cases of frequencies more than 200 MHz, the magnets are superconducting.
0
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4
5
6
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Shindo
Figure 6 The distribution of the NMR magnets classified by Shindo.
hospitals, and received 17 replies out of 26 inquiries, coveting 36 NMR magnets. The obtained data were analyzed by the committee. The type of the NMR magnets classified by Proton resonance frequency is shown in Figure 5. Figure 6 shows the distribution of the NMR magnets classified by Shindo. Four magnets were fixed by anchor bolts, 19 magnets had no anchor, and for the other 13 cases the existence of anchor bolts was not confirmed. The situation was much more severe than that of MRI magnets. Four magnet quenches were reported in the replies. Three quenches occurred after the NMR cry0stat fell over. Figure 7 shows an example where the NMR magnet leaned against the wall. In all of these cases, the NMR cryostat was not fixed by anchor bolts. The relation between the damages and existence of anchor bolts is shown in Figure 8. In one case, the anchor bolts of the NMR cryostat were torn off by the powerful earthquake tremors, and a magnet quench occurred. The center of gravity was relatively high for NMR magnets in general. Therefore an effective supporting method should be investigated to guard against falling over in addition to anchor bolt fixing. Cold Evaporator 1398 cold evaporators in Hyogo, Osaka and a part of Kyoto prefecture were investigated. The LNG and LPG tanks are not included in this report. No cold evaporator fell over. Also, there are no reports on cryogenic liquid leaks. Slight irregular sinking took place on 66 cold evaporators
Figure 7 An NMR magnet fell over and leaned against the wall. A magnet quench occurred in this case.
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out of 1398.25 cold evaporators sunk irregularly by 0.5 to 1.0 %. The sinking ratio of 41 cold evaporators was more than 1.0 %. The largest sinking ratio was 21.3 % ( Concerning the definition of sinking ratio, see Ref. [ 1].). Generally speaking, irregular sinking could be attributed to ground subsidence. The structures of the cold evaporators had sufficient strength to withstand even Shindo 7 tremors. It was fortunate that many valves were shut off because the day when the earthquake struck the city followed a 3 day-long holiday.
25 20
t-.
I-I Unconfirmed IN No anchorbolts
I-~ I [
15
.IQ
E
Z
10
5 0
quench after
quenchafter positionshifted
no damage
falling down positionshifted without quench Cryogenic Equipment in the Universities Kobe University and Kobe University of Mercantile Marine (KUMM) had various kinds of Figure 8 The relationship between damage and existequipment in their cryogenic laboratories. ence of anchor bolts. Many cryostat and gas manifolds fell over and glass Dewars for liquid helium were broken in the universities. The seismic intensity scale near the universities is estimated to have been Shindo 6. There is a big difference in the ground conditions between the two universities. The KUMM is built on reclaimed land, and liquidization (ground liquefaction) was caused by the earthquake, whereas no liquidization was observed at the Kobe University which is located on hillside. A severe crack was observed near the cryogenic center of the KUMM showing a trace of liquidization and the intensity of the earthquake. At the Kobe University, a purifier fell over and was broken. There was no damage to the dilution refrigerator and NMR magnet. At the KUMM, a liquid nitrogen cold evaporators sank by 3.0 %. One heavy superconducting magnet was set in a cryostat with castors and it was not damaged after moving on the floor many times, perhaps because the center of gravity was sufficiently low. It is hard to say, however, that this method of using castors could be applied in every case. There was no damage to the R&D superconducting ship.
SUMMARY The Damage Investigation Committee investigated the damage to MRI, NMR magnets, Cold evaporator, storage tanks and other cryogenic & superconducting equipment, and reported the findings in this paper. There was little damage to the MRI magnets. This is because these machines are designed as a cold ship type. There were no patients in MRI full body scanners, because the earthquake struck early in the morning. The situation could have been drastically different if diagnoses were going on. The safety manuals for emergency cases should be reexamined and arranged for easier usage. We should also study the safe installation and housing of MRI systems. The damage to NMR magnets was much more severe than that to MRI magnets. Four magnet quenches were reported. More effective fixing methods for equipment with a high center of gravity should be studied. ACKNOWLEDGMENT The members of the damage investigation committee wish to thank Dr. T. Kamei, Chairman of the Technical Committee for the Safety of Equipment for MRI of Japan Society of Magnetic Resonance in Medicine for his great contribution. They wish to thank Mr. T. Kurino, Executive Director of the Society of Nontraditional Technology for helpful financial support. REFERENCES Haruyama, T., Fujioka, K., Hirabayashi, H., Ishigohka, A., Ishiyama, A., Kawate, Y., Nishijima, S., Noguchi, T., Ogino, O., Ogiwara, H., Okazaki, O. and Sato, A., What Happened to Cryogenic and Superconducting Equipment in The Great Hanshin Earthquake Proc. CEC (1995) TH-C1-7 EIZO JOUHO special issue ( in Japanese ) (1994) 26 26
Cryogenic engineering
Superfluid helium
This Page Intentionally Left Blank
Thermohydraulic Behaviour of HeII in Stratified Co-current Two-phase Flow
Rousset Bernard*, Gauthier Alain*, Grimaud Laurent*, Bezaguet Alain**, van Weelderen Rob** *CEA-Grenoble/DRFMC/SBT, 17 rue des Martyrs, 38054 Grenoble Cedex 09, France **CERN, European Organization for Nuclear Research, LHC/ACR, CH-1211 Geneva 23, Switzerland
Experiments were conducted with saturated superfluid helium at CEA-Grenoble to simulate the cooling scheme of the Large Hadron Collider (LHC) superconducting magnets at CERN. Two phase co-current stratified flow was circulated through a 40 mm inner diameter, 92 m long heated tube, with a slope of 1.4%. Mass flow rates and temperatures ranged between 1.5 and 6.3 g/s, 1.8 and 2 K respectively. A full description of the flow was obtained measuring mass flow rate, quality, void fraction, wetted surface and pressure losses by means of specific diagnostics. Results of measurements were clarified through visual observations obtained from a CMOS video camera cooled at liquid nitrogen temperature.
INTRODUCTION The high-field superconducting magnets of the LHC will operate in static pressurized helium II at a maximum of 1.9 K and about 0.1 MPa, in which the static and dynamic heat loads will be transported by counterflows in He 1I to a heat exchanger tube threading its way along the magnet string[ 1]. Inside the tube, a flow of saturated helium II absorbs the heat loads by gradual vaporization of the liquid phase. Provided the tube is made of a good thermal conductor, and its inner surface properly wetted by liquid, the applied heat load can be extracted across a small temperature difference (in the order of a few mK). The length of such a scheme is limited by the pressure drop in the two phase flow, which raises the inlet saturation temperature, by the control delays due to the rather low flow velocities, as well as by the development of possible flow instabilities in the tube. A series of experiments was performed in order to investigate the limit of this cooling scheme and also to develop a predictive model of thermohydraulic two phase helium II behaviour.
EXPERIMENTAL FACILITIES All measurements were performed at Grenoble on our Superfluid Helium Test Facility which has been described previously [2-3]. Following path ABCDEFG, the flow scheme, shown in Figure 1, consists of a J.T. valve, a subcooler between A and B, the test loop between C and D, and finally a saturated bath from which the mass flow rate exits at point E at 100% saturated vapour. The experimental loop contains immersed thermometers (Tx05 and Txl5) at the extremities of the 40 mm I.D. 92 m long line, two preheaters (W1 and W2) to vary the inlet quality, a visual observation section (a 0.2 m long Pyrex tube lightened by LEDs) and a thermal test section. This thermal section serves as LHC magnet heat transfer model, it consists of 0.4 m long copper section surrounded by a heated pressurized helium II chamber. This chamber is maintained at 0.1 MPa by means of a capillary coming from the 4.2 K atmospheric bath which filled the pressurized He II. A variable heater W3 and a thermometer Txl3 are immersed in this helium II pressurized chamber. Total mass flow rate (thto t ) is measured at room temperature thanks to venturi and gas flowmeter. The experimental procedure begins by opening the J-T valve at a constant value in order to fix the desired total mass flowrate. Power W4 in the saturated bath is used to regulate the bath liquid level. The speed of the pumping group is then adjusted to reach the required temperature. A series of runs is performed 519
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increasing gradually the inlet quality. At each step, the same power W3 is injected in the pressurized chamber and the corresponding increase in temperature is recorded.
4.2 K 1 bar
t..___.~ ~ _ _ _ _ _ . _ _ _ ----------
Heat exchanger
,~
Magnetic valve
Video W3
Txl3
Tx02 D
be lUl Ltion
I!
,I!
I,
Copp
w
~:?;~:??i:i:i:~:i:~:?~:?~:~:i:i:!:i:i:i:i:i:i:i:~:i:i:!:i:i:i:i:i:~?i:?~:?i:i:i:!:i:i:i:i:i:i:i:i:i:i:i:i:i:i:i:i:i:!:?i:i:~:i:?i:?i:~:i:~:!:!:!:! 40 mm
T~5
i A
UI
Pressurized
Tx 13
Static He II
o
I
Tw4 Figure 1" Flow scheme HYDRAULIC BEHAVIOUR OF AN ADIABATIC TWO-PHASE FLOW Measurements Due to the high conduction inside superfluid helium, immersed thermometers Tx05 and Txl5 follow the saturation curve. An in situ re-calibration allows an accuracy better than 1 mK for relative measurements between the two thermometers. Using the vapour pressure curve, it's straightforward to calculate the pressure drop along the line. The vapour mass quality ( x ) and thus the superficial vapour velocity (Vg s) is determined by energy
s2 (
)
s2
balance: ~ d rh,o t L~.at x + rh,o t h ! - ~., Q, sl
9
sl
x rhto t Vg s = ~
/9 v being the vapour density, Ato t the
P v Atot
total cross section area, h~ the enthalpy of the subcooled or liquid phase, Lsa t the latent heat of s2
vaporization and ~ Q the power deposited between the absissa sl and s 2 . Applying this equation sl
between point B and point D, and also between point D and point E gives two different ways to estimate x at point D (end of the line).
ICEC16/ICMC Proceedings
521
The relation between void fraction (O~) and quality is inferred from measurement of liquid excess burnt by W4 during a transition between steady states. The change of liquid hold-up (1-O~) is measured by integrating the transient increase of total mass flow rate at the outlet of the cryostat. Analysis No pressure drop measurement has been reported on helium 1I two-phase flow, and very few data exist for helium I at pressures below one atmosphere. Authors compared their He I measurements both with Lockart Martinelli (L.M.) and homogenous model, this latter being found the more appropriate[4]. In our case (with Helium 1I), neither L.M. nor homogenous model was able to fit our pressure drop results (figure 2). A calculation of the flow pattern map[5] applied to the whole range of our experiments indicates that the flow would remain always stratified, owing to the large ratio between saturated liquid and gas densities. For the same reason, even at low qualities, the void fraction (O~) is just a little below one. and the pressure drop can be approximated by a simple vapour flow calculation. On figure 2, all points with qualities, total mass flow rate and temperatures ranging respectively between 0.1 and 0.9, 1.5 and 6.3 g/s and 1.8 and 2 K can be fitted by a linear interpolation, the slope of this latter being equal to the friction factor. Unfortunately, there are at least two drawbacks to this simple model: -the friction factor found is roughly 10% higher than in the case of single gas flow (for which a Blasius law was employed) gas flow assumption doesn't give access to the wetted surface, which is the main heat exchange parameter. To overcome these drawbacks, a complete diabatic two-phase flow model was developed with the assumption of separate flows and horizontal interface. In case of adiabatic flows, the model was simplified and compared to the experiments (figure 2). Computed and experimental void fractions are also in good agreement (figure 3). In addition, for low qualities, video pictures show horizontal interfaces with wetted perimeters in the order of predicted ones (figure 4). For high qualities, the void fraction is so large that the liquid isn't visible anymore.
F i g 2" T o t a l
pressure
91 m l o n g
1000 -
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and
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!
0.5 0.7 Outlet quality
! 0.9
THERMAL BEHAVIOUR OF AN ADIABATIC TWO-PHASE FLOW Thermal measurements were performed with the pressurized chamber. Calibration of the thermal heat exchange between pressurized and saturated bath was obtained in a separate cryostat with the internal tube wetted over its full surface. Thermal exchange is dominated by the Kapitza conductance and was
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measured for the full range of required bath temperatures. Assuming heat flux being transported by liquid alone, the wetted surface fraction can be calculated with the thermal heat exchange calibration: Sw ATc = ~ S w being the wetted perimeter, Sto t the total perimeter, A T e the temperature difference Sto t AT between the pressurized and the saturated bath at calibration and AT thetemperature difference between the pressurized bath and the two-phase flow with the same input power W3 and the same saturated bath temperature as at calibration. Figure 5 indicates a drastic increase of wetted perimeter for superficial gas velocities higher than 5 m/s. This unexpected result, in apparent contradiction with horizontal interface hypothesis was also mentioned previously[6]. Other experiments are planned to discriminate between various possible flow patterns" horizontal stratified and mist two-phase flow, horizontal stratified plus thin films on wall two-phase flow, semi-annular two-phase flow... Fig4: video picture for 6 g/s, 1.9 K and 40 W inlet input power
Exp. 6.3 g/s, 1.81 K - Calc. 6.3 g/s, 1.81 K o Exp. 6.3 g/s, 1.97 K - - - Calc. 6.3 g/s, 1.97 K 9 Exp. 3.5 g/s, 1.97 K Calc. 3.5 g/s, 1.97 K
---
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x
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0.04
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4.0
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Fig5" Numerical and experimental comparison
t
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8.0 10.0 VGS (m/s)
CONCLUSION Over the range of our experiments we have not found any physical limitation in this cooling method for the LHC. We have developed a theoretical model which very accurately describes the two-phase flow. However the enhancement of the heat transfer at high gas velocities as compared with a stratified flow pattern remains to be explained.
REFERENCES Lebrun, P., Superfluid helium cryogenics for the Large Hadron collider project at CERN Cryogenics (1994) 34 1-8 Rousset, B. and al, Operation of a forced-flow superfluid helium test facility and first results Cryogenics (1992) 32 134-137 Rousset, B. and al, Pressure drop and transient heat transport in forced flow of single-phase helium II at high Reynolds numbers Cryogenics (1994) 34 317-320 Huang, X. and Van Sciver, S. W., Pressure drop and void fraction of two-phase helium flowing in horizontal tubes Cryogenics (1995) 35 467-473 Taitel, Y. and Dukler, A. E., A model for predicting flow regime transitions in horizontal and near horizontal gas-liquid flow AIChE Journal (1976) 22 47-55 Casas, J. and al, Design concept and first experimental validation of the superfluid helium system for the Large Hadron Collider (LHC) project at CERN Cryogenics (1992) 32 118-121
Influence Of Fountain Pressure On Heat Transfer To Superfluid Helium Yanzhong Li*, Udo Ruppert**, Ingrid Arend**, Klaus Lt~ders** *College of Energy and Power Engineering, Xi'an Jiaotong University, 710049, P.R. China **Physics Department, Free University Berlin, D-14195 Berlin, F.R. Germany Coating wire using porous materials with appropriate pore size is a simple way to get HelIp inside porous coating of solid immersed in Hells bath. The mechanism results from the thermo-mechanical effect of superfluid, which creates fountain pressure, and allows liquid to become subcooled. The improvement on heat transfer by the method is obviously observed in experiments. A physical model describing fountain pressure inside porous layer is abstracted and the correlation of fountain pressure to the geometric parameters of porous materials (pore size d and porosity e) is established. The fountain pressure is measured experimentally by means of a Siemens pressure transducer.
INTRODUCTION Normal liquid helium (HeI) is widely used to cool superconducting magnet in many practical cases. It is well known that superfluid helium has better cooling properties than HeI. Taking Hell as coolant instead of HeI, the critical current of superconductor and magnetic field intensity of superconducting magnet will be much higher[ 1]. Pressurized superfluid (HelIp) is the best coolant, in which the maximum heat flux and recovery heat flux transmitted are much higher than those in Hells. The cryostabilization of superconducting magnet is also much improved by the outstanding property of HelIp[2]. The different critical performances of a superconductor wire cooled by different coolants could be clearly seen by comparison [3]. From. the viewpoint on cooling properties, HelIp should undoubtedly be recommended. However, the cost of getting HelIp is larger and cooling system must be more complicated, which result in an extra investment and difficulty in control. Nevertheless the excellent property of HelIp is still of strong interest to humanity. It is worthwhile to work out new ways to obtain pressurized (or subcooled) Hell. In present work, Subcooled superfluid is obtained in saturated Hell bath by means of thermomechanical effect. Using porous material with proper pore size to coat wire, a pressure increase is produced inside porous coating and the liquid inside becomes subcooled, by which the heat transfer is improved ultimately. The preliminary results were reported in ICEC 15[4]. THEORETICAL MODEL AND ANALYSIS Thermo-mechanical effect is a basic performance existing in Hell. The effect can transform extemal heat into an internal pressure by means of capillary channel. According to the idea, porous material with appropriate pore size can be taken as the coating of solid sample, which keeps sample from directly contacting with Hells bath. The constructing model of sample and coating is shown in figure 1. The coating 2 of sample 1 (in figure l a) is composed of many connected fine pores, so it can be simplified as capillary tubes as shown in figure lb. When heat transfer occurs, the heat from sample warms up the liquid around it first and causes normal fluid component to increase and superfluid component to decrease, so that a flow happens from one side to another through porous coating under the control of density difference. If pore size is fine enough, the flow of normal fluid will be restricted and only superfluid is allowed to go through the coating, therefore a pressure increase in liquid helium around sample may be built up. The magnitude of the pressure increase depends on heat flux and geometric parameters of porous material. The built fountain pressure results in liquid helium subcooled, as HelIp, which furthermore intensifies heat exchange from solid to helium. 523
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...9
<,a"
9.....
,,[
.I
! I !
I !
T I w --r
III ----
lr
atr
(a)
(b)
Figure l Principle of thermo-mechanical effect in porous coating
1 Sample 2 Porous coating
The higher the fountain pressure is, the more subcooled the superfluid becomes. The maximum of fountain pressure can be calculated theoretically. It is known that the equation of motion for superfluid
Ot
: svv-!ve
(1)
p
Where S and p are the entropy and density of liquid helium, VT and VP are gradient of temperature and pressure. Suppose the Gibbs potentials at both sides of porous coating are equal, and furthermore the fluid will not be accelerating, that is,
Ot
= 0, the equation is simplified into
(2)
aP = p . s
dT
The superfluid under the action of temperature difference will reach an equivalent pressure difference
Ap = p" - P ' =
--Tjf'Tp. SdT
(3)
If temperature changes from T b to TX along the direction of thickness of coating, the fountain pressure reaches its maximum as following APmax =~l,~p,VdT
(4)
When the bath temperature Tb=1.6K, the maximum is as large as 63.3KPa. Differing from pressurized superfluid controlled by external pressure, the helium situated inside porous coating has an inconstant pressure during heat transfer. The situations of pressure under different cooling conditions are illustrated in figure 2. It is analytically found that for the sample with coating (case 2), fountain pressure value during heat transfer goes up along with temperature decreasing rather than constant as in Hells and HelIp, which changes the final point of heat transfer process and enlarges the range of heat transfer in superfluid. It will be favorable to the improvement of heat transfer. As the analysis above, fountain pressure has much influence on heat transfer, furthermore its magnitude relates to geometric parameters of porous coating. The consideration might be done by means of mass flow equation in porous material[5]
ICEC16/ICMC Proceedings l(/t = r 2A 6 p A P 81 r/
(5)
525
Solid
Where r and I are the radium and depth of pore, is porosity, A is the contact area, and r/ is the viscosity of helium. The heat flux from the surface of sample to superfluid is
H
Hel ,,
!
q = --~
C p ( T ) d T = A Cp(TIAT
S
,,
,
,
""I
_
/
|
(6)
Cp is the specific heat of helium. From the equation (2) describing thermo-mechanical effect, the fountain pressure is approximately written as
=
p SdT= p S ( T ) A T
(7)
Where the density p is assumed to be constant due to its little change at different temperatures. Combining the equation (5), (6) and (7), one obtains AP2 = 32 ql q(T)S(T) 1 Cp(T) d2c
y
(~ Hells
(~) Hells+ coatings
(~) Heilp
Figure 2 The change of helium pressure in different bath
(8)
d is the diameter of pore. that is AP oc d - i t -~
(9)
This is the relation of fountain pressure to porous coating geometric parameters obtained by mathematical deduction. It also conforms to experiment. EXPERIMENTAL MEASUREMENT The experimemal sample is made of copper wire Potentialleads (Q40~tm, L~30mm) coated with different materials that are capillary filter, gypsum and Contac~ gypsum-alumina mixture. In order to measure the fountain pressure during heat transfer experiment, the sample is constructed with a narrow gap between two sheets of porous coatings as shown in figure 3. A Siemens Current/iead~~: pressure transducer connecting to the gap reports the pressure increase at any time. The maximum ~ Porousplate fountain pressure (corresponding to peak heat flux) measured at different bath temperatures under different coating conditions are plotted in SiemensKPYPressuretransducer figure 4. A larger value is obtained at lower bath temperature and not much difference can be Figure 3 Principle of the fountain pressure identified for different coating materials. It is measurement in porous coating considered that the change of produced fountain pressure mainly relies on the geometric properties of coatings but not material themselves. The permeating measurement of mass flow shows that the pore size d and porosity ~ are, respectively, d=0.301am, ~ =0.48 for pure gypsum, d=0.331am, ~=0.60 for mixture, and d=0.181am, ~ =0.50 for capillary filter. Although capillary filter has a nominal pore size d-0.021am, its construction of cylinder
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capillary tube is different from the requirement (global pore) of mass flow equation (5). Figure 5 shows the relations of peak and recovery heat flux to maximum fountain pressure, from which an approximate
qp~:ln(APmax) could be found. A conclusion is reached by measurement that the increase
relationship
of fountain pressure is profitable to heat transfer.
8
40
Cu ~p4o~:rnr~
Cu ~o40/zm
~
Hells: 250•
30
,~~
1~ 20
qp ......
qr
te
@ i 1 .B
I
i
i
i
i
i
i
i
I
i
i
i
1 .7
Bath
i
i
i
i
i
i
1
I
I
i
.9
Temperature
i
i
1
1
i
i
i
I
2.1
Tb, K
i
i
i
i
@
I
I
I
I
I
I
1 Pressaxre irtcrease
10 ~Pm.=, K P a
Figure 4 The pressure increase at peak point as Figure 5 Peak and recovery heat flux as the the function of bath temperature function of pressure increase 7]-Capillary filter A: Alumina-gypsum mixture 4Y: Pure gypsum CONCLUSION By means of thermo-mechanical effect, a fountain pressure is set up in HeIIs bath through porous materials. Fountain pressure allows saturated helium II to become subcooled and heat transfer is thus improved. The existence of fountain pressure is verified theoretically and experimentally, moreover the geometric factors affecting fountain pressure are found and the relation between them are obtained. ACKNOWLEDGMENTS The work is supported by the National Education Commission of China. The experiment is completed in TTL, Free university Berlin, Germany, and supported by the Bundesminister for Wirtschaft, ERPSonderverm6gen. REFERENCES Van Sciver S.W., Christianson, O., Stabilization of large superconducting magnets in superfluid helium IEEE Transactions on Magnets (1979) Mag-I 5, 1 Bon Mardin, G., Claudet, G., Seyfert, P., Verdier, J., Helium II in low-temperature and superconductive magnet engineering Adv. Cryo. Eng. (1978) 23 G-5 358 Li, Y.Z., Wu, Y.Z., Wu, Y.Y., Ruppert, U., Arend, I., Liaders, K., Heat transfer from superconductor wire to superfluid helium ICECI 6, Kitakyushu, Japan, May, 1996 Li, Y.Z., Wu, Y.Y., Arend, I., Liaders, K., Ruppert, U., Influence of porous coatings on heat transfer in superfluid helium Cryogenics (1994) 34 supp. 301 Li, Y.Z., Wu, Y.Y., Wu, Y.Z., Ruppert, U., Arend, I., Liaders, K., HelIp achievement in Hells bath by coating wire with porous media Cry_ogenics and Superconductivity_ (1996 Chinese) 24 2
Analysis and Characterization of Saturated Bath He II Heat Exchangers
Steven W. Van Sciver 1,2 and Scott J. Welton 2 1. Mechanical Engineering Department, FAMU-FSU College of Engineering 2. National High Magnetic Field Laboratory, 1800 E. Paul Dirac Dr., Tallahassee, FL 32306-4005, USA
We describe a design and analysis approach for He II heat exchangers that consist of a saturated bath immersed in a pressurized He II reservoir. Analysis is based on analytic and numerical solutions of the appropriately developed He II heat equation. The findings of this study are used to interpret performance data from the He 11 heat exchangers in the NHMFL 45-T Hybrid cryostat. Suggestions for utilization of these findings in the design of similar heat exchangers are described.
INTRODUCTION Many large cryogenic systems provide refrigeration with pressurized He 1I that is cooled by a saturated bath heat exchanger. In most cases, the heat exchanger consists of a closed end vessel or tube immersed in the pressurized bath. A JT expansion circuit supplies saturated helium to the vessel from a large He I reservoir, boiling around 4.2 K. The pressurized He II is in thermal contact with the saturated bath through the wall of the heat exchanger tube. Usually these heat exchangers are made from high conductivity metal to minimize the thermal resistance of the wall itself. The Kapitza conductance coefficient, at both surfaces of the heat exchanger, controls the heat transfer rate. Saturated He II heat exchangers that consist of closed end channels may be limited in their effectiveness by temperature gradients within the bulk fluid. At moderate heat fluxes, He ]I is known to have a finite effective thermal conductivity and a measurable temperature gradient (dT/dx o~ q3). Large temperature gradients, in the saturated He 1I, can cause boiling within the fluid and reduce the effectiveness of the heat exchanger. This condition is undesirable. An analysis is presented of heat transport in a He II containing channel with a surface heat transfer condition along its length and which is surrounded by a bath at constant temperature. This problem has an analytic solution for constant properties, which can be used to assist with design. Temperature dependent properties and more complex geometries require numerical solution.
PROBLEM STATEMENT AND HEAT EXCHANGER DESIGN The relationship between the dimensions of a heat exchanger and its performance must be understood when designing a He II refrigerator similar to what is described in this paper. Primarily the two issues of concern are the surface area and the helium cross-sectional area. We consider a saturated He II heat exchanger consisting of a vertically oriented tube immersed in a pressurized bath of liquid helium. The heat exchanger must have sufficient surface area to meet the temperature difference requirement at the maximum heat flux. In simple terms, the wetted surface area (As) must exceed the value, A s - Q/UAT]ma
(1)
x
where U is the overall heat transfer coefficient. 527
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Additionally, the heat exchanger must be designed such that boiling is avoided within the bulk saturated He II. Boiling will initiate if the local temperature of the helium exceeds the saturation condition. This condition can be qualitatively described by referring to the p-T phase diagram for helium, Fig. 1. The temperature at the surface of the heat exchanger (x=0) is given by To and that at the bottom (x=L) by TL. Assuming that the vapor pressure at the liquid surface is fixed at Po, then the liquid surface will remain at constant temperature, To, independent of the load on the heat exchanger. Due to the heat flux within the saturated bath, the temperature along the heat exchanger will increase with x. ff this temperature exceeds the local saturation condition anywhere within the heat exchanger then boiling will occur. This condition is shown by the temperature profiles in Fig. 1. The critical condition for boiling will occur when the temperature profile crosses the saturated vapor pressure curve or, more specifically, when the temperature gradient at the liquid-vapor interface becomes greater than the slope of the saturated vapor pressure curve,
dT > Pg dx x=0
svP
(2)
where t9 is the liquid density and g is the acceleration of gravity. Equation (2) leads to an approximate relationship for the maximum heat removal rate of a heat exchanger,
Qmax
= Ac( pgRT2~f-l
(3)
where R is the gas constant, f-1 is the He II heat conductivity function, and Ac is the helium cross section. The latent heat is ~ and Po is the saturated vapor pressure at the surface temperature. In addition, we have taken advantage of the vapor being nearly ideal. From Eq. (3), it is evident that the cross section of the He II within the saturated bath determines the maximum cooling capacity (Qm~) of the heat exchanger. This cross section must be large enough to ensure that Qm~ is greater than the refrigeration capacity of the system at normal operating conditions. This refrigeration capacity (Qref), for a typical He II system, can be described by,
Qref - ( 1 - x)p lJ2
(4)
where x is the quality of the two phase helium after the JT expansion and V is the volumetric flow capacity of the vacuum pump. Equations (3) and (4) can be used to generate refrigeration capacity curves. The curves for the NHMFL Hybrid cryogenic system are shown in Figure 2. The maximum heat flux for these heat exchangers is approximately 1.44 W/cm 2 near 1.88 K. At normal operating temperatures (near 1.8 K) boiling is suppressed as Qref < Qmax. On the other hand, at temperatures greater than 1.86 K, the refrigeration power will exceed the capacity of the heat exchanger and boiling will occur.
ANALYSIS Heat transport in a saturated He II heat exchanger can be analyzed in terms of the non-linear turbulent counterflow equation with a surface heat transfer condition [1,2],
d ( -1 dT~ 1~3 -d-~x f
--~x j
PU + ~Ac (Tb - T ) - 0
(5)
where P is the cooled perimeter. We have chosen the convention with the x coordinate measured from the free surface into the bulk liquid.
ICEC16/ICMC Proceedings
529
Analytic Solution Complete solution to Eq. (5) requires numerical techniques since the thermal transfer functions, f -1 and U are both temperature dependent. Such a solution has been previously reported [3]. However, normally heat exchangers are designed to operate at low AT in the fluid allowing a constant property analytic solution. Because of space limitations in the present paper, we discuss only the results of the analytic solution to this problem. The reader is referred to future publications for more detail. Equation (5) can be non-dimensionalized into the form,
d (dO)-m4/3 0
~xx ~ x
-0
where, O - ( T b - T ) / ( T b - T o )
(6) and m 4 / 3 - PU(T b - T o ) 2 / 3 / f - 1 / 3 A c . The parameter m 4/3 scales the
heat transfer to the bath to that carried by counterflow through the bulk fluid. Since Eq. (6) is a second order equation, it requires two boundary conditions. At x =0, 00 =1 by definition. At x - L, the lower end of the heat exchanger transfers heat through its surface such that, (f-ld0/dx)1/3 = UOL(Tb_T0)2/3. Equation (6) can be integrated one time to yield an expression for the heat removal rate of the heat exchanger,
where, M = 1.19[a3f-l(Tb - To)2pu] 1//4 and a - acU3(Tb - TO)2 / 2 P f -1 . Equation (7) can be used to determine the temperature of the He II at the lower end of the heat exchanger for a given set of conditions. To compute the temperature profile along the heat exchanger, one needs to integrate Eq. (6) a second time. Numerical Solution In Eq. (5), the heat conductivity function (f-l) and the heat transfer coefficient (U) are strongly temperature dependent. Therefore, the accuracy of the analytic solution may be limited as the temperature gradient within the heat exchanger increases. At high heat fluxes, TL- To can be quite large because of the cubic dependence of the temperature gradient on the heat flux in He II. For these conditions, a numerical solution to Eq. (5) must be utilized if accurate results are required. An iterative finite-difference technique was used to solve Eq. (5) for the temperature profile in the saturated bath of the heat exchanger. Typical results for various refrigeration loads are shown in Fig. 3. Also plotted, for comparison, is the saturation temperature line. The temperature profile at 30.3 W is tangential to the saturation temperature curve, representing the maximum capacity of the heat exchanger, Qmax. This value agrees nicely with the results from Eq. (3). From Fig. 3, it is clear that the maximum temperature gradient [dT/dx]max as well as the minimum temperature margin [Tsat- T]min occur near the liquid surface. In other words, if boiling can be suppressed just below the liquid surface, then it will not occur anywhere else in the heat exchanger. These results reinforce the usefulness of Eq. (3). Fig. 4 compares the analytic and numerical results for the NHMFL Hybrid heat exchangers. For this design, there is excellent agreement between the two solution methods. Values of ot range from 2.67x10 -~~ to 6.87x10 -7.
CONCLUSIONS Two methods for modeling a saturated He II heat exchanger have been presented and compared. The analytic model assumes constant properties and is good for low heat fluxes. At higher heat fluxes, the numerical method is useful for more accurate results. The numerical results are also useful for understanding the limiting parameters for a heat exchanger. The limiting refrigeration capacity of a saturated bath He II heat exchanger is the point at which two phase boiling occurs below the saturated liquid surface. This point is dependent upon the cross section of the helium within the heat exchanger. Thus, by careful selection of tubing diameters, it is possible to design a heat exchanger in which boiling will never occur under normal operating conditions.
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I C E C 1 6 / I C M C Proceedings H e a t E x c h a n g e r a n d Refrigeration
He Phase Diagram
50
.03
Capacity . . . . . . . . .
. ,,
....
J
m
40
m J a B
,pJ'
.02
r~
3o
Liquid He II
2o
, ,,""
ID
He Vapor
;~
lO ...... i
.01 1.6
1.8 1.9 Temperature (K)
1.7
1
i
i
60
~ 22.9 w
0
A
O O O
ca
O O O
~40
o
A A ZX
A A A
A
[] 30.3 w [] []
[] [] []
[] [] []
[]
~1 20
1
1
1
|
6.
i
._1..........
2.0
2.2
Figure 2. Maximum cooling capacity of the heat exchanger (Qmax) and refrigeration capacity (Qref) as a function of temperature.
T e m p e r a t u r e Profiles ~13.6 w
l
Temperature (K)
Figure 1. He Phase Diagram showing temperature profiles in saturated heat exchanger. The heat load is increasing from left to right. Note that boiling occurs in highest heat load condition.
8O
J
1.8
1.6
2.0
Qref (W) ]
l
T L VS
36.1 W
x
X X X
X X
X • X
Q
1.86 ~d II
|
1.85
......
1.84
.
.
.
Numerical .
.
.
/
.
1.83 1.82
......
E 1.81
[,,.
0
1.80 1.8
1.82 1.84 Temperature (K)
1.86
Figure 3. Temperature profile along the length of a saturated bath heat exchanger for various heat loads.
i__
L---.m~,r
1
I
I
I
l
l
I
I
I
I
I
I
I
I
I
40 10 20 30 Refrigeration, Q (W) Figure 4. Temperature at bottom of saturated heat exchanger. Comparison of analytic and numerical solutions. 0
ACKNOWLEDGMENTS This work is based upon research conducted at the National High Magnetic Field Laboratory (NHMFL), which is supported by the National Science Foundation under Award No. DMR-9527035. The authors wish to thank G. Bon Mardion and F. Viargues for useful discussions during the design phase of the 45 T Hybrid Cryogenic System.
REFERENCES 1. Shajii, A. et. al. Design and Performance of Forced Flow He II Heat Exchangers Adv. Cryog. Engn. (1990) 35 1165-1172. 2. Huang, Y. and Van Sciver, S. W. Forced Flow He II Heat Exchangers Cryogenics (1996) 36 (to be published). 3. Warren, R. P. Heat Flow in a He II Filled Fin Adv. Cryog. Engn. (1984) 29 335-42.
Experimental study of propagation charactaristics of He I -H e II interface
Koji Kamiya, Masahide Murakami and Naoto Yanagise Institute of Engineering Mechanics,University of Tsukuba,Tennoudai 1-1-1,Tsukuba, Ibaraki 305, Japan
In spite of the second order phase transition, He I- He II interface appears by imposing a local temperature difference in a bulk liquid helium system. In the present experiment a kind of interphase layer which divides liquid helium into He I and He II phases and thus includes interface is given rise to by local cooling. We focused our attention on the dynamical aspects of HeI-HelI interface behavior caused by cooling. We investigated the interface behavior by optical visualization and by measuring the temperature with a newly developed superconductive temperature sensor, the working principle of which is based on the variation of the electric resistance of a second kind superconductor. We succeeded in detecting an interphase with a high sensitivity sensor. Some dynamic aspects are discussed in this paper.
INTRODUCTION Since pure liquid helium has a complex order parameter and undergoes a second-order phase transition, interface can not exist without any external forces or flows. If there appears an interface, the transition of liquid helium across the lambda transition line results in a first-order transition. So far, a number of experiments have suggested the existence of an interface between He I and He II under heat flow. Onuki[ 1] found out a solution corresponds to a state with two phases separated by an interface by solving time dependent Ginzburg-Landau (TDGL) equation for one dimensional stationary state. The aim of this experimental research is to produce a non-stationary state by cooling and to detect an interface, where an interface is involved in an interphase layer. The dynamical behavior of a He I-He II phase boundary has been experimentally investigated. An interphase appears when He I under a saturated vapor pressure condition is suddenly cooled from an initial state at slightly higher temperature than T)vby vapor evacuation, and propagates downward from a liquid helium free surface in a cryostat. He II phase is created behind an interphase. We investigated the interphase behavior by optical visualization method and by measuring the temperature in a transition layer with a superconductive thin-film temperature sensor.
EXPERIMENTAL PROCEDURE 531
532
ICEC16/1CMC Proceedings
In the present experiment an electromagnetic valve makes it possible to suddenly start vapor evacuation for sudden cooling of He I. In this way the temperature is reduced abruptly so as to pass through the lambda point along the saturated vapor line. The cooling heat flux across a free surface can be calculated from the evaporating vapor flow rate measured with a gas flow meter. For the measurement of the propagation speed and the temperature profile in a transition layer, we utilize a superconductive temperature sensor made of gold-tin thin-film vacuum deposited on the side wall of a thin quartz rod with a diameter of 40/1 m. The resistance changes continuously in a super-normal transition region, because the element behaves as a second kind superconductor. Any temperature variation can be detected as a resistance change. The sensor has a very high sensitivity and a rapid time response. The temperature range with a high sensitivity can be adjusted so as to include the lambda point within it by changing the thickness ratio between gold and tin layers. The minute trimming of the sensitive range can be made by changing the bias electric current through a sensor.
sensing element ( qb=40gm )
The speed of interphase propagation is
calculated from the time interval between the signals of two sensors separated by about 5ram in the
, , . . / / . / m e t a l needls
direction of propagation. Fig. 1 shows the schelnatic drawing of a temperature sensor. In addition, optical visualization makes the local density variation around an interphase visible in order to investigate the dynamical stability of it during propagation. In the experiment the Schlieren optical system is used. The contrast of a Schlieren image is proportional to the spatial gradient of the
t
/sensor
stem
,' 5mm
!
Fig. 1 The illustration of a superconductive temperature sensor
density.
RESULT An interphase is generated by rapid vapor evacuation from a He I free surface in the saturated vapor pressure state. A transition layer including a phase interphase parallel to a liquid free surface propagates downward through an experimental field, which is observed through the observation windows of the cryostat. Typical examples of Schlieren photographs are shown in Fig.2, where a transition layer propagates downward quite stably. A high-speed video camera is also used to record Schlieren images and to measure the propagation speed. It is seen that the interphase layer propagates with a speed of several centimeters per sec and has a thickness of a few hundred micrometers. Also the dynamic stability of an interphase is expelimentally examined by giving a disturbance to a layer during propagation. Fig.3 shows a series of photographs when the interface passes through the obstacles composed of a seks of horizontal Cu pipes with a diameter of 15mm. First, the interphase is interrupted by the obstacles and diffiaction appears. Then a planar interphase is recovered again. It can be concluded from the pictures that the interphase is dynamically stable.
ICEC16/ICMC Proceedings
533
Fig.4 shows the relation between the cooling heat flux and the speed of propagation of interphase layer measured at two different locations measured from a free surface by the time of flight method with a double temperature probe. It can be pointed out from the figure that the interphase speed strongly depends on the cooling heat flux, nearly proportional to the heat flux. It is also found that the speed of the interphase at the same cooling rate is higher for closer to a free surface.
Fig.2 Photographs of a propagating interphase
Fig.3 A series of photographs when an interface passes through obstacles.
DISCUSSION It is seen from the visualization result a stable interphase layer with a finite thickness can be clearly defined, which includes an interphase. This means that an interphase layer has a definite structure bounding between
534
ICEC16/ICMC Proceedings
It is seen from the visualization result a stable interphase layer with a finite thickness can be clearly defined, which includes an interphase. This means that an interphase layer has a definite structure bounding between He I and He II. It seems that the formation of this layer is not a microscopic phenomenon which can be explained on the basis of quantum mechanics, but such a macroscopic one as a therm',d boundary layer. It is, however, rather difficult to identify the precise location of an interphase within a layer at the present stage of our experiment. Focusing on Fig.3, it is very interesting to note that the interphase layer diffracts around the obstacles in spite that it is just a contact surface which separates two phases. We suppose that the above behavior of an interphase layer might be connected to the propagation of second sound in He II phase. It is suggested from the result of Fig.4 that there exists the minimum cooling heat flux to produce an interphase layer having a finite speed just to compensate the back ground parasitic heat leak of 0.016W/cm 2 in the particular experiment. Unless the cooling rate is higher than the back ground heat leak, no interphase layer appears. It is further supposed that even for a little higher cooling heat flux than the minimum heat flux, an interphase, if it is once generated, may disappear without reaching far below a free surface. The experimental data shows that the propagation speed of interphase is a function of distance from a free surface of liquid helium. The speed of interphase slowly decreases with the increase of the distance from a free surface even at a constant cooling rate.
CONCLUSION It is confirmed that a He I - H e II interphase can
- - e - d i s t a n c e f r o m s e n s o r to free s u r f a c e 917cm - ~ - d i s t a n c e f r o m s e n s o r to f r e e s u r f a c e ' 10cm
be visualized by Schlieren method, propagating
5
downward with a speed of several centimeters per sec and having a thickness of a few hundred micrometers. It is found from the measurement of the temperature and the evaporating vapor
-
a minimum cooling heat flux coinciding with the
"~"
heat
leak
is
needed.
The
propagation speed, nearly proportional to the cooling heat flux, becomes small as the distance from a free surface increases.
, ,~
i i
~
!
!
!
!
,
i
!
'l
!
I
I
...........................
I
I
|ill
,
i
I
I
,::
.....................-
~
.='-~ 3 "='~2
ground
w !
4 _-
.=
flow rate that in order to produce an interphase, back
_l
1
-
I 0 0.03
' . . . . . ~ " . . . . . . . . ! ................. T i = 2 2 1 8 K I
I
I
1
0.035
I
I
I
0.04
I
I
l
I
I
0.045
I
I
I,i
0.05
l
I
I
I-
0.055
cooling heat flux(W/~)
F i g . 4 H e i g h t d e p e n d e n c e o f i n t e r p h a s e s p e e d for two c a s e s o f m e a s u r i n g l o c a t i o n f r o m free s u r f a c e
REFERENCE A. Onuki, Theory of Helium Under Heat Flow Near the 2 point. I. Interface of He I and He II, J. Low. Temp. Phys. 50, Nos. 5/6, 1983
Steady and Unsteady Heat Transfer from a Horizontal D i a m e t e r in a P o o l o f S u b c o o l e d H e I I a t P r e s s u r e s
Wire with a Wide Range
of
M. Shiotsu, K. Hata, Y. Takeuchi, K. H a m a and A. Sakurai* I n s t i t u t e of Atomic Energy, Kyoto University, Uji, Kyoto 611, J a p a n . F u t u r e Energy Research Assoc., Pasteur Building, 103-5 T a n a k a Monzen-cho, Sakyo-ku, Kyoto 606, J a p a n This paper reviews the authors' recent works on steady and transient heat transfer from horizontal wires with diameters of 0.08, 0.2, 0.5 and 0.7 mm in a pool of subcooled He II for liquid temperatures from 1.8 to 2.1 K and system pressures from 5.465 to 101.3 kPa. The followings are described: 1) Steady-state CHF values and a theoretical CHF correlation based on the Gorter-Mellink equations. 2) Life time of Quasi-steady state existing on the extrapolation of a steady-state Kapitza conductance curve for a stepwise heat input. 3) A correlation for the lifetime of the Kapitza conductance state which can generally predicts the lifetime for a stepwise heat input with any waveform in the early stage of heat input up to the stepheight.
INTRODUCTION The knowledge of steady and unsteady heat transfer on a solid surface in He II for various subcooling conditions is necessary for the stability design and safety evaluation of a large-scale superconducting magnet cooled by subcooled He II. The steady-state and transient heat transfer in He II have been investigated experimentally for a onedimensional heat flow system, namely, that for a metal heater positioned at the end of a He II filled constant cross section duct or tube whose sides are insulated. The results of the one-dimensional heat transfer experiments have been analyzed well in terms of two-fluid hydrodynamics and theory of mutual friction. On the other hand, there have been few systematic studies on two-dimensional and three-dimensional heat transfer from various shaped solid surfaces in He II till quite recently, although this would be the case for the heat transfer from windings of superconducting magnets. Recently, the authors investigated systematically the two-dimensional heat transfer (steady-state heat transfer[I-7] and transient heat transfer caused by stepwise heat inputs [1-4,7]) from a horizontal wire in He II: the test wires used were 0.08 mm-diameter Au-Mn(0.9 atomic%) wire, and Pt-Co (0.5 atomic %) wires with diameters of 0.2, 0.5 and 0.7 mm. Experimental conditions included saturated vapor pressures on the liquid surface of the bath with a wide range of liquid temperatures and liquid heads, and subcooled conditions at pressures for various liquid temperatures and cylinder diameters. This paper reviews the outline of the authors' works on steady and unsteady heat transfer from the horizontal wires in subcooled He II at pressures. STEADY-STATE
HEAT TRANSFER
[1-7]
Heat Transfer Process Figure 1 shows the heat transfer processes due to quasi-steadily increasing heat inputs to a 0.08 mm diameter Au-Mn wire in 2.0 K He II for the pressure at the level of the test wire axis, PL, of 6.4 and 101.3 kPa. As shown in the figure, the degree of the surface temperature jump at the critical heat flux(CHF) becomes smaller with the increase in PL and becomes zero (continuous) at around the atmospheric pressure. Critical Heat Flux
The authors [2,3] presented the following correlation for the steady-state CIIF, qst, on a horizontal test wire in subcooled He II (for Ts,,t(PL) > T~).
(1) where f ( T ) -1 = g(Tx)[T~'S(1 - T~'s)] 3 sx = 1559 J/(kg K) , Ax ~_ 1150
g(T),) = p2s~T~/Ax m s/kg . 535
,
TR = T/T),
536
ICEC16/ICMC Proceedings
,00 FF
~9
~
t
-
, %o_ _ , _ _ ~C.duct s ..... .. .
..... //
!
E
/
F
............
: .......
~
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-
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P L = 6 4 k9P a
9
~e~,m~
-
o,,I
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H~eater Dia. 0.08 mm
. ,41e _.o -~ ,,,o : . . . . . . . . . . o ,,, "~.~:..~,.~,:~. '_. . . .
Q~si-$teady State
/t ~ ~ . ~p
I
_
~. . . . . . . . . .
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.
,
. . . . . . .,o. ~
F ~ ~ , . o d , ~,o,. . / k , - ' - ~ ' ~ ~ t e r
- I F-
_/'F
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z
~ oo,o
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-
.
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_
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x ~ ~
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.
i
-
-
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1
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Ill
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P~=~0~.3 ~P~ ~ =2.0 K Dia. 0 . 0 8
mm
loo
-111
--
_ _.. _ ,r~:r~asir~ Heat. _Power :_ Lo ~ . . , - - ~ - - , . ~ . - . :
~T
, ,1
III-k Ill
i
-
-
!-
..... iti
...... 111
-
--
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l-I
III
10
.
noo
......
~ stead y- s i ta ~-~,-----
10
~
Reg;me for Each Pseudo-stepped Heat Input
.~/
0.1
(K)
Fig. 1 Steady-state heat transfer processes.
1
.,,
,
'
10
[ 100
PL
1000
(kPa)
Fig. 2 qst for 0.08 mm-diameter wire versus PL
Tp and T~ are bulk liquid temperature and ,~ temperature, respectively; and r0 is the test wire radius. The modification coefficient K in Eq.(1) was determined to be 0.58 by using the experimental data of q,t for various diameter wires. The exponent of 6.8 in the equation of f ( T ) -1 was changed from the original value of 5.7 for one-dimensional heat transfer in He II; the bulk liquid temperature at which the maximum of f ( T ) -1 occurs was shifted from 1.923 K to 1.96 K with this modification. Figure 2 shows the q,t values for 0.08 mm-diameter wire versus PL, and Fig. 3 shows the q,t at atmospheric pressure versus test wire diameter with bulk liquid temperature as a parameter. As shown in these figures, qst are dependent on Wire diameter and bulk liquid temperature but very weakly dependent on system pressures higher than A pressure. The curve predicted from Eq.(1) for each bulk liquid temperature, Te, is also shown in these figures for comparison. The experimentaJ data are well expressed by the correlation. TRANSIENT
HEAT TRANSFER
CAUSED
BY STEPWISE
HEAT INPUTS
[1-4,7]
Typical time traces of pseudo-step input Q, heat flux q, and surface temperature difference A T ( = T~ - TB) are shown in Fig. 4. Initially, the heat input rapidly increases in time and then it takes a constant value, Qs, after t - tA. The surface temperature difference and the heat flux remain constant at AT, and q,, respectively, for a certain duration ( t B - tA), then they begin to increase and decrease, respectively, at t = tB. The duration tL = tB -- tA is defined as the lifetime of the quasi-steady state heat flux q,. It was first found by the authors that the transient heat flux for the stepwise heat input increases along the 100
'"
PL =102.3 kPa
~.
TB
a
A
1.8K
I0
9 1.9K
S::::::::::~:...
-
TB=I.8K
.o.,,,r,,,.0~r,,
0,
/
~
|
[] 2 . 0 K 3
A2.1K
i ..........~ ....
,-10
X
.~ 2 ~
1"'-
-
5
..... 4..•
qs
-
~
q
z~Ts
/
/.,'11
b
2 o"_ 1 ~
U,e,*mo ~
/
Eq.0) 1
0.03
'
'" .....
9 9I l l l
0.1
1 Wire Diameter
10
Time
,
,
,,,[oo (ms)
(ram)
Fig. 3 q,t at atmospheric pressure versus wire diameter
Fig. 4 Time dependencies of the heat flux a n d surface temperature difference for a pseudo-step heat input
I C E C 1 6 / I C M C Proceedings
537
10 :~1:~
PL =102.3 kPa TB=2.0 K
~ C ~ 0 ~ %
9 A.
E30
OXen,_, "vO
AA
0.01
....
I .... 1
0
0.2 mm 0.5 0.7 mm
l. . . . . 2
PL=102.3 kPa 9
I
A
t ..... 4
0.i
5
I
I,,,I,,
0
( W / m 2)
Heater Dis.
TB=I.8 K
I"
A
, .... 3
(qs " qst) x 10 "s
!" [\
10~ ~ I~ ~ I , 9
Exponential-Step C& 0.08 mm Heat Inputs 0
9
0.001
" ....
mm
PseudoStep Heat Inputs
@
0.1
100 I
Heater Dis. 0.08 mm
(A
0.20 m m 0.50 m m
kO 0.70m m
Exponen-?
tial-Step | 9 0.70mm Heat Inputs k
9~
9
, I... 2
Pseudo I~ Step Heat I v Inputs
4
Estimated Curve for IdeQI Heat Input
VE3
9
Step
I.,,
I.,,
6
8
I,,
, I,..
10
12
IL 14
(qs'qet) x ro 1/3 x 10 -3 (W/m 5/3)
Fig. 5 Lifetime, tL, for varous diameter wires versus excess heat flux, qs - q s t for pseudo-step and exponential-step heat inputs
Fig. 6 Relation between tL/ro and (qs - q ~ t ) x r~13 for various diameter wires
Kapitza conductance curve and the curve's extrapolation, and then reaches a quasi-steady-state (ATs,qs) on the extrapolated curve. The values of tL for the wires with the diameters of 0.08, 0.2, 0.5 and 0.7 mm are shown against excess heat flux beyond steady-state critical heat flux, q~ -qs~, in Fig. 5. The lifetime at an excess heat flux is longer for a larger diameter wire, though the steady-state critical heat flux is lower for a larger diameter as given by Eq.(1). To investigate the effect of the heat input waveform in its early stage, an exponential-step heat input , which increases exponentially in time (Q = Qoeq~) with the period r of 0.5 ms and then keeps a constant value, was used. The data of tL on 0.08 mm wire for the exponentiM-step heat inputs are also shown in the figure in comparison with the data on the same wire for the pseudo-step heat inputs. The lifetime for the exponential-step heat input decreases almost in agreement with that for the pseudo-step heat input with the increase in excess heat flux up to around 5 x l0 s W / m 2, and decreases at a much greater rate for the excess heat flux higher than the value. The effect of the heat input waveform in the early stage was observed significantly. It was already shown in Fig. 3 that the q~t for a certain liquid subcooling is proportional to ro 1/3. Figure 6 shows the relation of tL/ro versus (qs -- qst) X rlo13 on various diameter wires due to the pseudo-step heat inputs for TB of 1.8 K. The values of tL/ro on various diameter wires almost agree with each other and can be expressed as a single curve on this graph, although relatively large difference can be observed especially in the high excess heat flux regime for each wire where the effect of the heat input waveform in the early stage is significant. The integrated values, (1/r~/3) ftt;B(q- qst)dt obtained from each runs for the pseudo-step heat inputs to various diameter wires for the liquid temperatures of 2.0 K are shown on the graph of ..
~
(1/%/3) ftt~(q- qst)dt
1/3
versus (qs - v~t)r0 in Fig. 7. At the time tst, the increasing heat flux reaches the q~t , and at the time tB, the lifetime for the value of q~ ends as shown in Fig. 4. The integrated values for all the experimental runs obtained for each liquid temperature are almost constant and they are independent of the step heights and of the initial waveforms during the rise time to the step heights. It was also clarified by the authors [4] that the integrated values are almost independent of PL from 6.4 to 102.3 kPa. The average of the integrated values for each bulk liquid temperature is shown in Fig. 8. The curve representing the average value can be expressed by the following equation,
fL,'(q-
J / m 8/3
= F(T.)
(2)
where F(TB) = 9.7928 x 105 - 1.8620 x 106TB + 1.1437 x 106T~ - 2.2748 x 105T~ for 1.8 _< TB _< 2.1. It is possible to evaluate the lifetimes due to stepwise heat inputs on various diameter wires from Eqs.(1) 9nd (2). For instance, the lifetimes for the ideal step heat inputs (rise time=0) which cannot be realized experimentally can be estimated by the following equation,
tL = [l/(q~ -- q~t)][Lt~(q- q,t)dt]~,. = [ l / ( q ~ - q~)]r~/3F(TB) ,, 1 / 3
(3)
The curve of (tL/ro) versus (q, -- q~t)r o for the ideal step heat inputs derived from Eq.(3) is shown in Fig. 6 in comparison with the corresponding experimental data for pseudo-step heat inputs. The values of tL
538
ICEC16/ICMC Proceedings ,.--,,
~
loo
14
B
"~ ,-..,' 1 0
C:>
Av~c~ o
X Pt. =102.3 kPa TB=2.0 K
I
PseudoStep Heat Inputs
•
Heater Dia. I~ 0.08 mm 0.2 mm 0.5 mm 0.7 mm
~ I
12 10
8 6
6.4 kPa <. PL <- 102.3 kPa
'~L:'
Exponen- ( : tiaI-Step Heat Inputs
,,....,
0.1
I'
'~
0
1 1 ' ' ~
5
l't'
10
t'
I t ~
15
''
I
20
t t * '
(qs'qst) x ro1/3 x 10"3 (Wm "s/3)
0.08 mm
~
0.7 mm l ' '
25
t L
30
Fig. 7 Integrated values of excess heat flux beyond q+t until the end of lifetime for various diameter wires
~
2
Eq. (2)
0 1.7
1.8
1.9
Te
2
(K)
2.1
2.2
+
Fig. 8 Average of integrated values on various diameter wires for each liquid temperature
on each diameter wire for the pseudo-step heat inputs become lower than the evaluated values for the ideal step heat inputs for larger values of (qs - qst). If one wishes to consider the impact on conductor stability of a transient heat pulse in a superconducting magnet cooled by subcooled He II, the transient heat transfer caused by the ideal pulse heat inputs with the heights Q~ and the width t~ will be of interest; the transient heat transfer is in the quasi-steady state Kapitza conductance regime and a rapid rising of the temperature does not occur as long as the width of the pulsed heat input is equal to or shorter than the lifetime of the quasi-steady heat flux mentioned above. CONCLUSIONS The authors' recent works on steady and transient heat transfer from horizontal wires with diameters of 0.08, 0.2, 0.5 and 0.7 mm in a pool of subcooled He II for liquid temperatures from 1.8 to 2.1 K and system pressures from 5.465 to 101.3 kPa were reviewed. Major contents were as follows" Effects of the test wire diameter, bulk liquid temperature and the pressure at the level of the test wire axis, PL, on steady-state CHF were clarified. Theoretical correlation, Eq. (1), for steady-state CHF based on the Gorter-Mellink equations was presented. This correlation can describe well the experimental results. Heat flux for a stepwise heat input increases along the Kapitza conductance curve and its extrapolation on the q vs. A T graph, attaining a quasi-steady state on the extrapolated curve. This quasi-steady state last a period of time which is defined as the lifetime for the quasi-steady state heat flux. A correlation, Eq. (2), which can generally predicts the lifetime for a stepwise heat input with any initial waveform up to the stepheight at PL from 6.4 to 102.3 kPa was given. REFERENCES 1. Shiotsu, M., Hata, K., and Sakurai, A., Transient heat transfer from a horizontal wire in superfluid helium caused by exponential and step heat inputs, Superfluid helium heat transfer, (1990) ASME HTD-Vol.134. 9-14 2. Sakurai, A., Shiotsu, M., and Hata, K., Transient heat transfer for large stepwise heat inputs to a horizontal wire in saturated He II, Advances in cryogenic eng., (1992) 37A, 25-35 3. Shiotsu, M., Hata, K., and Sakurai, A., Transient heat transfer for large stepwise heat inputs to a Horizontal Wire in Subcooled lie II, Advances in cryogenic eng., (1992) 37A, 37-46 4. Shiotsu, M., Hata, K., and Sakurai, A., Steady and unsteady heat transfer from a horizontal wire in a pool of subcooled He II at pressures from near A-pressure to atmospheric, Heat transfer and superconducting magnetic energy storage, (1992) ASME ttTD-Vol.211, 19-25 5. 5hlotsu, M., tiara, K., and 5akurai, A., Httect of test heater diameter on critical heat flux in He II, Advances in cryogenic eng., (1994) 39, 1797-1804 6. Shlotsu, M., tiata, K., Takeuchi, Y., Hama, K., and Sakurai, A., Critical heat flux on single horizontal wires in superfluid helium at pressures, Heat transfer 1.994, (1994) Taylor and Francis, 5, 141-146 7. Shiotsu, M., Hata, K., and Sakurai, A., Transient heat transfer from a horizontal wire in subcooled He II at atmospheric pressure for a wide range of wire diameter, to appear in Advances in cryogenic eng., (1996) 41
Two Dimensional Heat Transport in Hell Channel Including a Copper Wall
Tetsuji Okamura, Shinji Hamaguchi, Sho Sakuma, Tetsuya Suekane and Shigeharu Kabashima Department of Energy Sciences, Tokyo Institute of Technology, 4259 Nagatsuta, Midori-ku, Yokohama 226, Japan
Temperature distributions in a copper wall and helium in a channel have been measured. The channel is 170 mm in length and 7 mm wide with a gap of 10 mm. Both ends of the channel are open to a superfluid helium (HelI) bath at the atmospheric pressure. One of the side wall of the channel is made of copper and heated. Remarkable temperature rises in the copper wall are also observed when the phase of helium changes in the channel. Temperature distributions along the gap direction were observed in the channel when the maximum temperature in the helium is less than T a as well as when the 2 transition occurs in the channel.
INTRODUCTION From the view points of the stabilization and the quench protection of HelI cooled large superconducting magnets, it is important to understand heat transport characteristics of a superconductor and helium within a cooling channel [ 1]. Experimental studies of a configuration similar to a short section of a cooling channel in a magnet stabilized with copper and atmospherically pressurized HelI have been carried out in the present study. Temperature distributions in a copper wall and helium in a channel were measured for various heat fluxes given at one side wall of a channel, where the channel is filled with not only just HelI but also HelI- HeI coexistence or HelI- HeI - vapor coexistence. EXPERIMENTAL APPARATUS A tested channel is schematically shown in Figure 1(a) and (b), where it should be noted that the scale in the vertical and horizontal directions are not the same. The channel is formed in a fiberglass reinforced plastic (FRP) block and a copper block. The channel is 7 mm wide, 170 mm long, and has a 10 mm gap distance. Three nichrome foil heaters of 0.26 mm thick are stuck on the back side of the copper block. An electrical power can be applied to the respective heaters independently. Both ends of the channel are open to a pressurized HelI bath. The channel is placed vertically in the bath. The temperature distribution of helium within the channel is measured using 7 carbon glass resistors (CGRs) located at points 1-7 as shown in Figure 1. The temperature distribution of the copper block is measured using three germanium thermometers attached in the copper block and they are represented by A, B and C in this figure. The HelI bath temperature is measured by a germanium thermometer and is held 539
540
ICEC16/ICMC Proceedings 10
(a) Side view
f
m----..==,,
f'-I
I
r
'
I I
'
' 1''"1
'I
'
I I I
' 1''"
1 Tb=2.05K
Heater
10 ~ _ v+OFi _
~cD
.
A A
A
+c~
m Channel
~7o
(b) Bottom view
W
Channel
7! Heater -/
10-1 , 10 -2
,
, I,,,,I~
T
i,
10 -1
FRP Heater
L5.J
vA ,~
vm
Cu block
FRP
v~
Cu block
,
Tb~ I ,
, I,,,,I
10 ~
,,,
AT(K)
FRP Channel
FRP ~-- Cu block
Figure 1. Schematic view of the tested channel, (a) side view, (b) bottom view; 1 ~ 7: carbon glass resistors, A, B, C; germanium thermometers.
Figure 2, Steady state heat transfer characteristic for the bath temperature of 2.05 K. The respective marks in this figure represent the results for the positions which are illustrated under the x-axis.
constant during an input of electrical power to the heaters. A standard four-terminal technique is used to measure the value of input power. The temperatures are simultaneously measured for a step-like input heater current. RESULTS A steady state heat transfer characteristic for the bath temperature of 2.05 K is shown in Figure 2. The power is applied to all of the foil heaters at the same time. The x-axis, A T, and the y-axis, q, in this figure represent temperature rise from the bath temperature and heat flux density at the heater surface, respectively.
ICEC16/ICMC Proceedings
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We can see two distinct transitions in this figure when q of 0.5 W / c m 2 and 1.2 W / c m 2 are applied. The temperature distribution in the copper block changes remarkably at these distinct transitions. The input heat is transported by a mechanism of so called HelI internal convection for the heat flux smaller than 0.5 W/cm 2. We can also determine from this figure that HeI and HelI coexist in the channel for the heat flux of 0.5 W/cm 2 < q < 1.2 W/cm 2, and vapor generates for the heat flux larger than 1.2 W/cm 2. Another noticeable fact can be found from this figure that the helium temperature differs along the channel gap (horizontal) direction as well as the channel axis (vertical) direction before the 2 transition occurs in the channel. This fact suggests that the heat is transported by HelI internal convection two dimensionally in the present channel.
' " "
'
'~"
Tb =2"05K ,
g.
q=l'5W/cm'l
I
,' "-,,'
,'
9
,"-..
6
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~
/
\4 "'~
"
-50 0.01
A T(K) J
50E E 0 ".r.~-_0
10 Tb=2.25K ] q= 1.5W/cm2 /
0
/
/
i i ii
8
i
, ....
Middle part heating
. """" / . . / " " " " ~
............
.'
Lower part heating /
/ /
/
/
!
-50 0.01
0.;
,/~ J
I
,
I
,,,,I
0.05 0.1
Helium temperature
5
1 "~;(K)
....'"
,"
........""
Upper part heating
:"
t'
/:
/ /
I" /"
Copper block temperature
--zx--Upper part heating + U p p e r part heating --o--Middle part heating ~ M i d d l e part heating --v--Lower part heating --v--Lower part heating
Figure 3. Temperature distributions in the copper block and helium near the surface of the copper wall along the vertical direction.
0
0.5
1.0
1.5
2.0
q ( W / c m 2)
Figure 4. Variations of the energy stored in the copper block with the heat flux density for the bath temperature of 2.05 K.
The temperature distributions in the copper block and in the helium near the copper wall along the vertical direction for the bath temperature of 2.05 K are shown in Figure 3. The y-axis in this figure represents the distance from the center of the channel. The positive number in the y-axis means the upper part of the channel. The applied heat flux density is 1.5 W/cm 2. HelI and HeI coexist in the channel for this heat flux density. The results for the bath temperature of 2.25 K, i.e. subcooled HeI, are also shown. Solid lines and broken lines in these figures represent the copper and helium temperature distributions, respectively. The marks of A , O and V show the results for the cases of the upper part, the middle part and the lower part heating, respectively. It can be found the fact that the copper block temperature rise at the unheated part is kept considerably lower for the bath temperature of 2.05 K as compared for the bath temperature of 2.25 K. The helium
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ICEC 16/ICMC Proceedings
temperature rise is also lower for the case of the bath temperature of 2.05 K as compared that for the case of the bath temperature of 2.25 K. It can be considered from these facts that most of the applied heat is transported to the bath through HelI because of higher thermal conductivity of HelI than that of copper. In order to discuss about the temperature distributions in the copper wall and helium in universal channels quantitatively, two dimensional numerical analyses combining heat transport in helium (HelI and HeI) [23], heat diffusion in solids and Kapitza conductance are required. The stored energies in the copper block in the steady state are estimated from the measured temperature distributions. Variations of the stored energy with the heat flux density are shown in Figure 4. Three lines in this figure represent the results in the case when the heated parts are different, i.e. the copper block is heated at the upper part, at the middle part or at the lower part. We can see the fact that when the heat is transported by the HelI internal convection (q<1.0 W/cm2), the stored energy in the case of the middle part heating is largest and those in the case of the upper part heating and the lower part heating are almost the same value. It is considered that almost all of the applied heat is transported to the HelI bath through the nearer end of the channel to the heated part when the upper or lower part is heated. The stored energy in the case of the lower part heating becomes larger than that in the case of the upper part heating when the 2 transition occurs in the channel. CONCLUSIONS The temperature distributions in the copper wall and helium in the channel were measured two dimensionally. 1. The temperature distribution in the copper wall changes remarkably when the phase of helium changes in the channel, i.e. at the /l transition and vapor generation. 2. Two dimensional temperature distributions in the helium are observed when the heat flux is less than the /l transition heat flux, q a, is applied as well as the heat flux larger than q a is applied. 3. The value of the stored energy in the copper block in the case of the middle part heating is the largest, and that in the case of the upper part heating is the smallest for the same applied heat flux density. REFERENCES Caravan, E. and Van Sciver, S. W., Axial heat transport in Hell in a thick - walled copper channel Adv. in Cry. Eng. (1981) 27 475-483 Okamura, T., Suzuki, T., Seki, N. and Kabashima, S., Heat transport in HelI channel with phase transition Cryogenics (1994) 34 187-193 Ramadan, N.H. and Witt, R.J., Natural convection in large HelI baths Cryogenics (1994) 34 563577
Numerical Investigation of Evolution of Vortex Line Density in The Case of Transient Heating
Takashi Kanari and Masahide Murakami Institute of Engineering Mechanics, University of Tsukuba, Tennoudai 1-1-1, Tsukuba 305 Japan
Numerical simulation was carried out for the propagation of a planar thermal shock wave, or a second sound shock wave, generated by pulsative heating from a planar heater. The calculation is based on Landau's two fluid equation system supplemented with Vinen's vortex line density (VLD) evolution equation in order to take into account the superfluid turbulent behavior. The source term is introduced to the third term of RHS of Vinen's equation, and the problem of the initial vortex line density is discussed. The numerical result is compared with some experimental data.
INTRODUCTION It is expected that the numerical simulation will replace some parts of experimental studies of superfluid helium (He II) as in the study of ordinary fluid flows. For the numerical simulation of steady thermo-fluid dynamic phenomena in He II, the Gorter-Mellink mutual friction term have been added to the momentum equations to yield fair results. However, a mathematical model for nonsteady superfluid turbulent state has not been completely established yet. Two mathematical models have been considered to treat non-steady superfluid turbulent behaviors. One is an introduction of the modified Gorter-Mellink mutual friction term where the Gorter-Mellink coefficient is assumed to be an a priori function of time. And in the other the two fluid equation system is supplemented with Vinen's vortex line density(VLD) evolution equation. The latter is adopted in the present study because it seems to be superior to the former[l]. The problem of initial VLD in a quiescent state is still one of the open questions. Initial VLD L0 has been usually assumed to be the order of l0 s crn -2 to obtain appreciable effect of quantized vortices, but it is considered to be physically too large. The introduction of the source term to Vinen's VLD evolution equation may give a promising solution to the problem, though it is not widely accepted.
NUMERICAL SIMULATION Landau's two fluid equation system which governs the behavior of He II in one-dimensional situation is written in a vector form as OU
OE
0---~-I- 0X = S
(1)
s U-
v~ pv
,
E-
pn v~ + p~v~ psv~ 1
~
7v~ + # p + p~v~2 + psv~2
,
0 A p ~ p , v ~ , ( v n , - vc) 2 / T Ap~v~(v~ _vc)2 0
B-
Here, p, v, T and p are the total density, the mass velocity, the temperature and the pressure. Subscripts n and s indicate the normal and super components, and vns stands for ( v n - vs). vc is 543
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ICEC16/ICMC Proceedings
the critical velocity and # indicates the chemical potential which is defined by d~ -
+
1A,~ p
1 p~. 2 2p
(21
B is a vector arising from the mutual friction between the normal and super components, and A appearing in vector B is the Gorter-Mellink coefficient defined by 1 h
d - ---3pm(v~
BL L. -v~) 2
(3)
Here, L is the vortex line density (=VLD; total quantum vortex line length per unit volume), and BL is Vinen's mutual friction constant. Vinen's vortex line density evolution equation[2] is written as follows; OL c9(vLL) BL Pn ~ h L2 lot At- OX --- ~1 y - ~ - [Vns[ i ~ - X2--m + ~ [Vns[~ '
(4)
where, vn indicates the drift velocity of vortex tangle and X1, X2 and 7 are functions of temperature. To solve these equations the explicit MacCormack scheme, a kind of finite difference scheme, with the flux corrected transport (FCT) algorithm[3, 4] is applied. The present finite difference scheme is successfully adopted in order to improve the sharpness of a shock front and the suppression of numerical oscillation through the introduction of the FCT scheme compared with the previous attempt[l]. Heat flux given by the heater as an boundary condition is given in a trapezoidal form varying with time. The calculation is carried out for two peak heat fluxes, qm~, 9 and 20 W / c m 2, for the heating time tH fixed at 500 #s except the result given in Fig.1. The thermodynamic quantities are obtained by using a linear interpolation of tabulated values presented by Maynard[5].
RESULTS Figure 1 shows the comparison of non-linear evolution of a second sound shock both between the numerical calculation and the measurement with a superconductive temperature sensor by Shimazaki[6]. Only the peak values are plotted in this figure for the experimental result. The deformation of a wave profile is only resulted from the effect of the non-linear term because of the smallness of heating time, 30 #s compared with vortex development time. The comparison of the simulation result with the experimental one indicates that the present numerical scheme works well. Figure 2 is the result given by numerically solving Vinen's equation to elucidate the decay of VLD from three initial values of L, L0, 102, 104 and 106 cm -2. The vortex convection term of Eq.(4) is neglected in this calculation. It is found from this result that 104 a.nd 106 crn -2 is too large as the initial value of VLD because an experiment run is usually repeated with an interval at least 1 min. It is consequently suggested that L0 should be less than the order of 102 cm -2 as an initial condition for numerical simulations. Numerical result of the time development of L at a location of 5 m m apart from the heater is shown in Fig.3 for the case of q, 20 W / c m 2. The arrival of a heat pulse front is approximately 0.2 m s . In the case with non-zero source term in Vinen's VLD equation, L develops rapidly after the arrival of a pulse front up to l0 s cm -2 at which value appreciable vortex effects appear, and reaches still higher value even if L0 is quite small, as small as 102 crn -2. It is, on the other hand, seen that L0 should be larger than 105 cm -2 to cause equivalent vortex effect to the non-zero source term cases in the case without the source term. However, L0 of 10 s cm -2 is obviously too large as an initial value as discussed above. It may be concluded that it is most reasonable to introduce the source term to Vinen's VLD equation. The existence of t h e term was not accepted in a paper[7]. This may be resulted from their experiment in which the heat flux was too small to cause appreciable effect of the source term. The comparison of thermal shock wave profile is made between the numerical result (Fig.4) and the experimental result (Fig.5), which are profiles at 5 and 30 m m from the heater. The experimental data is obtained by Shimazaki[6]. There can be seen the deformation of second sound pulse due to the interaction with high density quantized vortices in the case of q - 20 W / c m 2, while
ICEC16/ICMC Proceedings
545
there is a little deformation in the case of q = 9 W / c m 2. Numerical result is in good agreemen~ with the experimental data in the case of q = 9 W / c r n 2. However, the manner of deformation is different between both results in the case of q = 20 W / c r n 2 in particular in the tail region behind a thermal pulse, where the profile of the experimental data is getting longer nearly by the amount of the deformation in the main body of a profile. The reason for this difference is not clear yet. This may be the very point of a further study. CONCLUSIONS Numerical calculation of the propagation of second sound shock waves in the case of transient heating with large heat flux is carried out. The initial VLD should be less than the order of 102 c m -2. The source term of Vinen's VLD evolution equation seems to play an important role under the condition of large heat flux. Numerical results agree with the experimental data except in detail. Calculation Experiment
120
9
9
|
100 80
),=4x 102,Lo = 102 7=4xl 02,Lo= 103 7=0 ,Lo= 10 6 . . . . . . . . . 7=0 ,Lo= 105 .......... 7=0 ,Lo= 104 ....................
60 [-.,
40 20 0
0
10
20
.I
I
30
40
109
~i
50
Distance (mm) Figure 1: Non-linear evolution of second sound shock wave. The peak temperatures of profiles are indicated by open circles for experimental data. Bath temperature T = 1.7 K, heating time t u - 30 #s.
106
j ..... j"
/
f ....
........... i ............... .............1 ..i
,o,*
105 >
104 103 10 2
1
2
Figure 3" Time variation of VLD at a distance of 5 m r n from the heater. The abscissa indicates time after the onset of heating. Bath temperature T - 1.7 K, peak heat flux q - 20 W / c m 2, heating time tH = 500 #s. Initial VLD L 0 - 102 and 103 crn -2 for the calculation with non-zero source term.
Lo=106 ......... Lo=104 Lo=102
105 10 4
103 102 101 0.001 0.01
107
Time (ms)
106
r !
108
60
0.1
1
10
100
Time (s) Figure 2" Time decay of VLD from some initial values calculated from Vinen's equation. Bath temperature T = 1.7 K , initial VLD, Lo - 102, 104 and 106 c m -2.
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ICEC16/ICMC Proceedings 1.75
9
1.75
|
5mm 30 mm
1.74
9
~"
|
5 mm 30 mm
1.74
1.73
1.73
1.72
1.72
,
\
!
9
9
1.71 1.70
'
0
1
;
i
~,,
;'
"
9
b. '
1
,
1.71 1.70
3
2
I |
i
I ,
0
i,
1
Time (ms)
2
3
Time (ms) (b) 20 W / ~ ~
(~) 9 w / ~ m ~
Figure 4" Time variation of temperature profiles at 5 and 30 turn from the heater.
1.75
I
9
III
9
9
5ram
1.74
9
,
1.75
9
9
~"
30mm ........
|
|
9
1.74
1.73
1.73
1.72
1.72
1.71
1 |'
1.70 1
2
Time (ms) (~) 9 W / c m ~
J
\ ,
%.
%
1.71
I | I
1.70
0
5mm
a
30mm
a
9
|
2
I
0
I
;"
1
9
'
2
3
Time (ms) (b) 20 W / c m ~
Figure 5" Experimental data of time variation of temperature profiles at 5 and 30 rnrn from the heater. REFERENCES 1 Murak~mi, M. and Iwashita, K., Numerical computation of a thermal shock wave in He II, Computers & Fluids (1991) 19 443-451. 2 Vinen, W., Mutual friction in a heat current in liquid helium II III. Theory of the mutual friction, Proc. R. Soc. London (1957) A242 493-515. 3
Fletcher, C., Computational Techniques for Fluid Dynamics, Springer-Verlag (1988), 143-170.
4
Book, D., Boris, J. and Hain, K., Flux-Corrected Transport II: Generalizations of the Method, Jour. Comp. Phys. (1975) 18 248-283.
5
Maynard, J., Determination of thermodynamics of He II from sound-velocity data, Phys. Rev. B (1976) 14 3868-3891.
6
Shimazaki, T. and Muraka.mi, M., Measurement of Characteristic Time of Quantized Vortex Development Using a Thermal Shock Wave, to be presented at ICEC 16 (1996).
7
Childers, R. and Tough, J., Critical Velocities as ~ Test of the Vinen Theory, phys. rev. lett. (1973) 31 911-914.
Heat Transfer From Superconductor Wire To Superfluid Helium Yanzhong Li*, Yezheng Wu*, Yuyuan Wu*, Udo Ruppert**, Ingrid Arend**, Klaus Liaders** *College of Energy and Power Engineering, Xi'an Jiaotong University, 710049, P.R. China **Physics Department, Free University Berlin, D-14195 Berlin, F.R. Germany A systematic research on NbTi/Cu wires (single core and multi-filaments) was conducted in HeI, Hells, and HelIp, separately. The results measured in three baths show a large difference in current sharing zone of a same sample, but a similar wire with insulation layer in Hell baths gives the results like those of bare wire in HeI. It means the insulation of superconductor is the main obstacle of heat transfer. In order to,<~domplete the measurement in a small cryostat avoiding destroying sample, a prediction based on the pre-test at room temperature is performed to provide the maximum of resistance and voltage of sample, which makes the measurement high efficient and more reliable.
INTRODUCTION Superconductor magnet is usually cooled by normal liquid helium, however higher cryostability could be obtained in superfluid (Hell). Superconductor cooled by HelI shows a larger critical current and recovery current than those in HeI. Furthermore, pressurized superfluid helium has better cooling properties than saturated one has. Some research work on heat transfer between superconductor magnet and different coolant was separately performed by former researchers[l]. In the present work, a systematic research is conducted to find and compare the heat transfer patterns of superconductor in different coolant baths. Meanwhile the influence of the insulation layer of superconductor on heat transfer is also taken as an important content to study. EXPERIMENTAL SETUP The HelIp cryostat is cooled down by pumping, and controlled by the opening of magnet valve through initial setting of temperature (for Hell bath) or pressure value (for HeI bath). The HelIp chamber is cooled by the surrounding Hells, and reaches, at last, same temperature as Hells and same pressure as HeI/GHe chamber (figure 1). The system can be adjusted from 1.6K to T% and from saturated vapor pressure to an atmosphere. The measurement circuit of sample is described as figure 2. The sample is heated by voltage supply with a triangular waveform, whose amplitude and frequency can be freely adjusted. The changing speed of supplied voltage is 4-~7mV/s. Current signal is read from normal resistor R N (1 mfZ ). Current and voltage drop of sample are recorded by X-Y recorder instantly. The temperature range of measurement is 1.65-~2.15K for Hells and HelIp and 2.2-~4.2K for HeI at saturated state. Pressure for HelIp bath is kept at 0.1MPa. 547
electricalI ,,
connection;
, I pump
Thermal barrier
J
Filling ~~ve
GH~
Hellp
:~=:!~:i;!i;.!~!~!i~.i~i:.~!i~!i~;i~=~i~~ells i!i!=.i!~:ii!i~;!.i==i!~ J
FeedI ml through I
Figure 1 Principle of experimental setup
LN2
548
ICEC16/ICMC Proceedings i
Ur
Waveform generator
D
Power supply
RN
t j / _
X
"T
sample NbTi foil
I
Sample
Figure 3 Schematic drawing of sample and its supports
Y
X-Y recorder
Figure 2 Diagram of measurement circuit Samples are made of NbTi alloy wires as listed in table 1. Figure 3 shows the sample support construction made of NbTi foil (0.3mm thick) and with elastic construction for the sample contraction. Potential leads and two extra potential lines are lead from different points to determine whether heat might be produced in solder points. It is repeatedly confirmed in many experiments that the reference Ur has same value of U. Therefore no effect from solder points needs to be considered. Table 1 Technical data of NbTi/Cu wires Model S 1-1.5/0.07
F54-1.35/0.05
, No. of Cores
Diameter (~tm)
Ratio of Cu to SC
1
70
1.5
54
50
1.35
Type Straight Twist
PRE-TEST OF SUPERCONDUCTOR AND CIRCUIT DESIGN To get and study the performance during quench, one must supply an enough large current to overpower sample to become normal state. It is clearly known that after quench, sample resistance rapidly rises and the heat produced in conductor greatly increases, which may result in the sample burning out at the point where the highest temperature or worst cooling condition is. In order to avoid the destruction of sample, in principle, one should limit the temperature increase or the resistance increase of the sample. However, the maximum of sample resistance or voltage drop must be known before measurement. Pre-test at Room Temperature Several previous tests for superconductor wire have been done at room temperature to get some ideas about the maximum of resistance and voltage. Because after quench the sample is surrounded by only helium gas, and the heat conductivities of helium gas and air are not much different, the maximum current and the highest resistivity are easier to get at room temperature than at low temperature. So, the results at room temperature could be used to predict measurement at low temperature. A NbTi wire (S1-1.5/0.07) is first used as sample at room temperature, which is gradually powered by supplying a current. X-Y recorder records the curve of voltage-current of sample at same time. The point, where the voltage and current signals become unstable and the color of wire surface begins to change, is taken as the limitation, in which the voltage per unit length Umax/L=25.71 V/m, the resistance per unit length Rmax/L=19.26 fEm. The temperature of superconducting wire at the point is about 400~ which is calculated according to the resistance-temperature correlation of pure copper. Circuit Design To prevent sample from being destroyed after quench, reducing the resistance of circuit is also an essential way besides controlling the sample resistance increase. That is useful for a constant voltage supply system. The voltage drop on the circuit wire will overload on the sample when its resistance
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changes from zero to a larger value after quench. Therefore, the maximum of circuit resistance for a certain sample could be worked out by calculation. An ohms law for the supply circuit at critical state before quench can be established as E] = ImaxR, where E1 is the potential of power supply, Ima x is maximum current, and R is total circuit resistance except sample. After quench, the sample resistance jumps from 0 to rmax, so that the relation is E 2 = vmax (R + rmax)- Assume a quench happens rapidly and no time elapses, then E 1 = E 2, that is, rmax /max R = Umax (R + rmax). Due to rmax >> R, then R = Umax Suppose the sample is 30mm long, the rmax /max maximum current for NbTi wire (S1-1.5/0.07) at Hell is about 25A, and the maximum voltage Umax is equal to 0.771V according to the results measured at room temperature, then the circuit R must be less than 30m~. To reduce the resistance, the circuit is divided into three parts depending on the local temperatures. The circuit wire and feed-through in liquid helium are all made of NbTi wire (Q0.5), the parts in gas are constructed with high pure copper wire (purity 99.999%, Q0.5x4), and the parts at room temperature are built with copper cable (Q50). After the arrangement, the total resistance R is lower to 27mf~ (at 4.2K), which is satisfactory to requirement. MEASUREMENT ON SUPERCONDUCTOR Figure 4 and 5 show the recording curves of U-I for superconducting wire (S1-1.5/0.07) in HeI and HelI. When point A is reached, sample quenches and quickly changes to point B, where superconductor completely becomes normal and surrounding Hell becomes HeI. Sample at point B bears a larger voltage due to the changed ratio of sample resistance to circuit resistance. The purpose of reducing circuit resistance is to lower the voltage at point B and protect sample against burning out. By improving cooling condition, the current at point A is higher, which also results in an increase of the voltage at point B. Therefore the possibility to destroy sample is higher. The difference can be obviously seen by comparison of Figure 4 and 5. In superfluid there is a voltage step before quench in U-I diagram, which means that the copper matrix has delivered the current, i.e., current sharing. Current sharing happens also in HeI, but in which the status is unstable so that subsequent quench happens immediately. In fact, the cryostability of superconductor mainly depends on the balance of ohm heat generation G and surface cooling Q in current sharing zone[2]. In HeI, G>Q, the generated heat cannot be cooled completely. The rest of heat is stored in conductor and raises the temperature to promote quench. In Hell, G
0.6
0.6 !
>
0.4"
1~4Ti.12S;(71..ielOlm) > ::5 >~0.2
>~0.2 0.0
l
0.4
B
D
CurrentI, A Figure 4 U-I curve recorded in HeI
A
0.0.
Nbri Sl (70~m) at Z06K 94KPa,Hellp
B
. . . . . .C.
.........................,,,,,,,,,,~
D
CurrentI, A Figure 5 U-I curve recorded in Hell
A
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ICEC16/ICMC Proceedings
Quench current Ima x and recovery current I r ( from point A and D in figure 4 or 5) as the function of bath temperature Tb are separately plotted in Figure 6. It is seen that the higher current is obtained in better cooling property bath. In HelIp, current increases monotonously along with Tb decreasing. When temperature goes down to 1.9K, the applied current is over the carrying ability of circuit and the sample is burnt out after quench, so the measuring system cannot be used for the measurement at Tb< 1.9K in HelIp. The wire F54-1.35/0.05 was measured in Hells and HelIp. The recorded curves for bare wire are similar to figure 5, but the curves for lacquered wire are similar to figure 4 like bare wire at HeI. The Ima x and I r for F54-1.35/0.05 wire as the functions of T b are plotted in figure 7, in which the results of lacquered wire are also illustrated. It is seen by comparison, the Ima x is reduced by 15% or so in HelIs, and by 30% (Imax) and 60% (Ir) of reduction in HelIp due to the existence of insulation. The Ima x and I r of lacquered wire have no clear difference in Hells and in HelIp, which means the insulation layer breaks down the direct cooling of liquid to solid. In this case, the cryostability of superconductor cannot be enhanced by only improving the liquid cooling property. More consideration needs to be paid to the heat transfer performance of insulation layer. The related research was reported earlier[3]. 35
20
Nb5 S1-1.5/0.07
30
NbTl FS4-- 1.a5/O.O~
at Hel, Hells and I-lellp 00"
<[
-
2 75+ 105rn.'m.
15
25OU"~
20.
!
.~
J.
J,.
p..
,,w
.w~
a.w
.w ,,r
Jw
JLj.~
15"
o-O.e
10-
oe 0'
1.~
,
1'.7
,
1'.~
!
2'.1
//,
j_e_e-e--~-o-o--o-e !
zs
.
.
.
.
3'.0 ~s
4'.0 4.s
|
|
|
|
|
|
i
|
Ba.th
Bath Terrprature Tb, K
i | |
| | = i , , | i | i
|
Telm.~e~'o.hu.'re
|i
i
I l l | i l l
~'b,
i
K
Figure 6 Im~ c and lr of S1-1.5/0.07 vs. T b ~ "
]m(TJc
......
:]r
0 - Q " HelIp
Figure 7 lma x and I r ofF54-1.35/0.05 vs. T b OO" Bare wire ~ O : lacquered wire -(~)-O: Hells
G" HeI
CONCLUSIONS (l)Cryostability of superconductor depends on the cooling properties of coolant. However, it is more important to improve the heat transfer performance of insulation layer for a technical superconductor with varnish layer when it works in Hell bath. (2) A prediction based on the pre-test of superconducting wires at room temperature is an efficient way to provide useful data for preventing samples from being destroyed after quench. ACKNOWLEDGMENTS The work is supported by the National Education Commission of China. The experiment was completed in TTL, Free university Berlin, Germany, and supported by the Bundesminister ffir Wirtschaft, ERPSondervermogen. REFERENCES Li, Y.Z., Heat transfer on superconductor wires with porous coatings in superfluid helium Dissertation ofXi'an Jiaotong University (1995) Collings, E.M., Applied superconductivity, metallurgy and physics of titanium alloys Plenum Press 1986 Li, Y.Z., Wu, Y.Y., Wu, Y.Z., Arend, I., Ruppert, U., L0ders, K., Heat Transfer Enhancement in superfluid helium by means of porous coatings J. of Xi'an Jiaotong University (1996) 30 5
The Effect of Spacer Arrangement on the Heat Transfer in He I and He II Channels
Hisayasu Kobayashi and Kouji Kawakami Atomic Energy Research Institute, Nihon University, 1-8 Kanda, Surugadai, Chiyoda-ku, Tokyo, Japan
Even when the heated copper surface is partially covered with a spacer, the high heat transfer rate in a He II channel does not decrease significantly above the critical heat flux at which normal-fluid sets in on the copper surface. This phenomenon should be of a great interest for superconducting magnets cooled with He II. On the other hand, the nucleate boiling region is reduced by the present of such spacers in He I.
INTRODUCTION Although the stabilization, including film boiling heat transfer for superconducting magnets cooled with He II, has been studied so far [1,2], the heat transfer characteristics above the critical heat flux has not been made the explicit purpose of an experimental investigation. We have made a preliminary measurement of the change in the heat transfer rate for a copper surface embedded in one side of a rectangular channel when spacers partially cover this surface. The characteristics are evaluated in terms of the critical heat flux Q ~t/Acu at which normal fluid appears on the copper surface (area Acu), the peak heat flux Qp/Acu at which film boiling starts, and the film boiling heat transfer rate. Since the boiling state above Q ,t in a channel may switch from film boiling to a turbulence of a mixture of liquid and gas [3, 4], the state above Q ,t is not necessarily the same as the steady film boiling state in an open bath. It is important to point out a few facts concerning the heat transfer from the side wall of channels at a given bath temperature Tb. 1) The critical heat flux Q ~/Ach per channel cross-section Ach is determined not by the surface condition but the channel geometry, dominantly by the length L (through the Gorter-Mellink conduction) [5]. 2) Therefore, Q x/Aex per exposed area Aex can become infinite at Aex - 0, at least in theory, as long as Q ~/Ach is not changed by the existence of spacers. This phenomenon parallels the fact that a reduction of the size of heated surface in an open bath results in a larger Q ,t [6]. 3) Furthermore, Q x/Aex increases because the effective L is shortened by the existence of the spacers, and a smaller effective length L increases Q ~/Ach. 4) Heat transfer in He I behaves differently since it is less influenced by the channel geometry, except for the indirect effect on the convective flow [7,8]. EXPERIMENTAL 551
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The test channel with a rectangular cross-section of 10 x 2 mm 2 and a length of 150 mm was fabricated from two fiberglass epoxy blocks (A and B in Figure 1) and a heat transfer cell. The cell consists of a copper block, germanium thermometers and a heater of manganine wire (101.8 f2) wound around the bottom of the copper block. Figure 1 shows the rear half of the channel with its opened at the both ends at the top and bottom. The copper block, the thermometer and the heater are fixed with vacuum-impregnated epoxy resin. The cell is attached to the center of the channel in such a way a square copper surface is exposed to the liquid inside the channel. The copper surface is partly covered with a fiberglass epoxy spacer of thickness 2 mm, the same size as the channel gap. The spacer can be glued to blocks A and B with silicon grease in two directions, either the normal (.1_) or parallel ( [[ ) to the long axis of the vertical channel, see Figure 1. Let 77 be the ratio of the exposed area Aex to the full square area Aeu (10 x 10 ram2). Data with 77 from 0.35 to 0.92 have been compared with 77 = 1 ( no spacer N). The channel is sufficiently long as that Q a. is small enough to cover a wide range of the 'film boiling state'. Under such conditions, the refrigeration power of the pressurized He II cryostat is sufficient to keep the bath temperature Tb constant within a few mK during the sweep of the heater current. Although the major part of the Joule heat produced in the heater is delivered to the liquid through the exposed copper surface, some heat diffuses to the surroundings including the spacer. Thus, the dependence of the temperature rise of T~ on the sweep rate of the heater current I (mA/s) was measured for the range from 0.18 to 3.6 (mA/s).
Block A
Block B
Heater
Exposed
Surface
Spacer Cu Block
Channel
Figure 1 Schematic arrangement of the test channel. Only the rear half of the channel is shown. T1,
T2, T3
and T4 represent the positions of the thermometers. RESULTS AND DISCUSSION Figure 2 shows traces of T~ as a function of heat flux Q/Acu for normal position of the spacer in He I and He II with I as parameter. Both Q ,~ and Qp are defined by the turning points of the traces. Although T~ is completely independent from I below Q a. and
Qp for the range of I measured, a strong dependence appears
in the 'film boiling' region. Similar jumps in T2 of about 0.5 K at Q a. behind the spacers have been observed (not shown). The relatively small increment in thermal conductivity/c of the insulators should not dominates the I dependence in these temperature ranges,/c (20 K)//c (4.2 K) = 1.5. It seems that a relatively long time is needed till a steady turbulent flow of the mixture of He II, He I and the gas is established [4].
ICEC16/ICMC Proceedings 4
'
!
'
,
!
i
Tb- 1.9K Posi ti on _L
Tb=4.2 K P0si ti on / 3
ab
L
E r
553
c d
a b c
d
e
z= 2
.~.2
o .==E
L) .==: CI
1
1
0
10
20
30
40
T1 [K]
50
0
10
20
T1 [K]
(a)
30
40
50
(b)
Figure 2 Traces of temperature T~ as a function of the heat flux Q/Acu with the sweep rate of heater current I as parameter for normal spacer arrangement (_L, 77 = 0.35) at bath temperature Tb of (a) 4.2 K and (b) 1.9 K. I (mA/s): a ) 3.6, b) 2.6, c) 1.8, d) 0.90, e) 0.51, f) 0.26. The open circles show the extrapolated equilibrium temperatures Teq at I = 0, see text.
'
t ,
I-
Tb = 4,2 K
'
'
'
I
'
Tb- 1.9K
'
'
'
'
'
'
t
'l
.,0 -!
I
'
I
'
I
'
~-
r~n
o
2 r
2
o
1
1
0
10
20
Teq
30
[K]
40
50
0
10
(a)
20
Teq
30
[K]
40
(b)
Figure 3 Equilibrium temperature Teq as a function of the heat flux Q/Acu at bath temperatures Tb of (a) 4.2 K and (b) 1.9 K. O" N (no spacer, 77 = 1), O" spacer_L (normal to the channel, 77 = 0.35), II- spacer_l_ (77 = 0.92), A" spacer_l_ (77 = 0.46 ),D" spacer
II
(parallel to the channel, 77 -- 0.5).
50
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ICEC16/ICMC Proceedings
The equilibrium temperature rises Teq shown in Figure 2 (open circles) were determined by extrapolating the T~ vs. i curves (not shown) to their intersections at I = 0 with Q as parameter. Teq as a function of Q/Acu is shown in Figure 3 for no spacer, normal and parallel spacer arrangements. To present the over-all cooling effect, the data is plotted not for Q/Aex but for Q/Acu. The reduction in heat transfer rate by introduction of the spacer is not significant in He II, unlike in He I. In other words, putting a few spacers on a superconducting conductor cooled with He II should not largely harm its stability. Heat transfer from the spacerless surface (N) in all regions in He I is enhanced by the convective flow passing through the channel without interruption by the spacer (see Figure 3a; Qp/Acu for N is larger than 0.8 W/cm 2 which is usually obtained in an open bath). The convective flow in He II also raises the nucleate boiling heat transfer to Qp [5, 7, 8] (see the N data of Figure 3b) as well as the enhanced 'film boiling' heat transfer region. The expanded state from Q a. to Qp is, however, unstable against a heat input with high I, i.e., the film boiling heat transfer may occur at Q 2. before reaching Qp. Thus, induced nucleate boiling heat transfer by the convective flow of the bulk liquid can not be relied upon to stabilize against transient heat input for both He I and He II. The time constant to establish induced flow of bulk liquid seems to be quite large [9]. Thus, when comparing the heat transfer rate as a function of 77, the bulk liquid flow through the channel is blocked, and then, the extrapolated data to 77 = 1 should be taken as the standard. Although the effect of the spacer arrangement has not yet been clarified sufficiently well in the present preliminary studies, the heat transfer rate in the parallel (II) spacer arrangement approaches that of the spacerless state (N) due to the induced flow of bulk liquid. This work was supported partially by the National Institute for Fusion Science, Japan. The authors wish to thank Dr. L. Boesten for kind reviewing for the manuscript and to Mr. Y. Fujimura and Mr. T. Murata for their technical assistance. REFERENCES 1
Meuris, C., Influence of an uncooled region on the stability of superconducting conductors International Institute of Refrigeration, Saclay (France) (1981) Commission A1/2 161-168
2
Meuris, C., Turck, B., Seyfert, P. and Claudet, G., Transient stability of superconductor cooled by superfluid helium at atmospheric pressure International Institute of Refrigeration, Saclay (France) (1981) Commission A1/2 215-223
3
Gentile, D. and Fransoirs, M.X., Heat transfer properties in a vertical channel filled with saturated and pressurized helium II Cryogenics (1981) 2_! 234-237
4
Breon, S.R. and Van Sciver, S.W., Boiling phenomena in HeII confined to a channel Cryogenics (1980) 26 682-691
5
Kobayashi, H., Uchida, M., Kajiura, Y., Konagaya, T. and Akedo, Y., Heat transfer in branched channel filled with He II Cryogenics (1996) 36 145-148
6
Kobayashi, H. and Yasukochi, K., A sample configuration effect on the heat transfer from metal surfaces to pressurized He II (1980) ICEC-8 171-174
7
Lehangre, S., Boissin, J.C., Johann,,s, C. and De La Harpe, A., Critical nucleate boiling of liquid helium in narrow tubes and annuli (1980) ICEC-8 (1980) 274 - 275
8
Sydriack. S.G., Cryostatic stability equation IEEE Trans. on Mag. (1977) MAG-13 682-685
9
Kobayashi, H., Uchida, M. and Akama, K., Transients in the heat transfer to He I and He II confined in narrow space. Fusion Engineering and Design (1993) 20 505-509
Pressure effect on the heat transfer in bath of superfluid helium
Ruzhu Wang
Peng Zhang
Jingyi Wu
Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai 200030, P. R. China
Considering the basic heat transfer rules of superfluid helium and the phase diagram of helium, pressure effects on heat transfer to HeII are especially studied. If the bath pressure is less than ~,-Pressure (P~), special peak heat flux density relations are shown to correlate bath pressure, hydrostatic head and modified pressure items (Van-der-waals pressure and fountain pressure). If the bath pressure is greater than )~-pressure(P>P~), a generalized formula of peak flux density of HeIIp bath is shown.
INTRODUCTION Engineering applications of superfluid helium (HeII) has been focused on by cryophysical and cryogenic researcher after the superconducting magnet was successfully cooled by superfluid helium, several kinds of subcooled superfluid helium cryostat were successively developed. A lot of papers about the heat transfer in superfluid helium bath were published[ 1], however the influences of pressure on the heat transfer in HeII bath have not yet systematically studied. In this paper the influences of various types of pressures (bath pressure, hydrostatic head, Van-derwaals pressure and fountain pressure) on the heat transfer in a bath of HeII are studied. PEAK HEAT FLUX DENSITY qp W H E N P < Pz Saturated superfluid helium (a) Noiseless film boiling In the research of saturated superfluid helium, the peak heat flux density qp is reached by overcoming the subcooled degree ATm~x. By Clausius-Clapeyron equation and assuming 9v, 9 and L are the densities of gas helium, HeII and the latent heat of vaporization, and 9gh is the hydrostatic pressure, we get T ATmax pgh (1) Pv L The relation of heat flux q and temperature gradient VT in HeII is 3 1 q VT (2)
f (T)
For a cylindrical heat transfer problem (wire radius r0, q0 is the heat flux at the radius r0) , eqn. (2) be written as dT = f (T)qo3(ro)3 dr r So if eqn.(3) is integrated from r =oo to the boundary of gas helium and by assuming the temperature of the boundary is ATmax if film boiling occurs, the peak heat flux density at r0 is got like bellow: qp3= 2tI~(r0_______)). T 1 r0 p---~,f(----~ 9pgh
can (3) rise
(4)
where tI)(r0) is the size effect parameter introduced for modification. From eqn.(4), we see that the 3 varies linearly with the hydrostatic pressure 9gh; but if the hydrostatic pressure tends to zero, quantity qp the heat flux density qp will tend to zero, too. It is obvious that this kind of deduction deviates from the reality. Figure 1 demonstrates the test results from reference[2] that shows a big value of qp at a zero 555
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hydrostatic pressure. We can deduce further: assuming the experiment is done in space, that is, there is no gravity, if eqn.(4) is applied, because g=0, thus qp=0. In fact, the two kinds of situations above are impossible. The critical heat flux density of superfluid helium is of some value even if it is in the state of zero gravity or the pressure head h tends to zero. The reason is that a special pressure - Van-der-waals pressure exists when film boiling onsets, this pressure is caused by Van-der-waals force. Gradt tried to use this pressure term to explain his experimental results qualitatively, however non quantitative explanation was shown[2]. If the physical concept of Van-der-waals pressure is applied, the problem above can be successfully solved by modifying the pressure item. The Van-der-waals pressure is APA=~/v 2 (5) where a is Van-der-waals constant and v is the specific volume of the formed gas film. So the maximum pressure difference between gas and saturated liquid in the state of film boiling is APmaX : pgh + AP~ (6) and equation (4) can be modified to 3 2~(~) 7' 1 q r -- . . . 9 . . ( p g h + AP.~ ) (7)
~,I, f(7')
ro
25
I0 9 8
20
15 x"
o-el,
5
0
20
40
60
80
100
120
h,mm
Figure 1 The peak heat flux density as a function of the immersion depth h, HeIIs bath with a temperature 2.0K (.noiseless film boiling, PtRhl0 wire with a diameter 401am), i:experimental result,-:theoretical result.
-
200
500
80{) AP = ~ h
1100
1400
1700
2000
+ A/' .t + A t ' F , P a
Figure 2 Relations between peak heat flux density and pressure increase AP, experiment was pertbnned in HeIIs at 2.0K using a RhFe-wire with a diameter of 381am), i:experimental result, -:theoretical result.
The above theory can be verified by the experimental results shown in Fig. 1, in which a PtRhl0 wire with a diameter of 40~tm is used as the sample, and the Hells bath is kept at 2.0K. The correlated size parameter ~(r0)=0.011, and the calculated Van-der-waals pressure APA=113-130Pa. The determined qp value at a hydrostatic head of zero is qp=5.8 W/cm 2, the experimental result shows a value of 5.5W/cm'[ the relative error between them is only 5.5%. More details are shown in Fig. 1, which indicates that the Van-der-waals pressure term can be combined with hydrostatic head successfully to describe the peak heat flux behavior. In order to explain the pressure effect on the heat transfer further, we use the measured data from reference[3], in which a wire (RhFe-wire with a diameter of 38jam) is wrapped by porous coatings, so there exists a fountain pressure outside the wire surface. Detailed measuring apparatus and the interesting results have been shown by Li et al.[4]. Considering the combined effects of hydrostatic pressure, Vander-waals pressure and fountain pressure, the pressure difference between the gas fihn and saturated liquid, AP, is written as pgh + AI'~ + AI'~;, eqn.(7) becomes 3 24~(Ib) it' 1 ql,
= .
.
r0
.
.
.
A,L .f(7)
AI'
(8)
The calculated results are shown in Fig.3 by using eqn.(8), the data points ( i ) are the experimental results from reference[3], in which AP~: are 0, 948Pa, 1673Pa respectively. In the case of APv=1673 Pa, there exists A t ' + t~. > I~, where Ps is the saturated vapor pressure, but the peak heat flux densities are
ICEC16/ICMC Proceedings
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almost the same when P > t~. So we can simply substitute AP with I~ - t~s,. At the hydrostatic head, the estimated ~(r0)=0.02, detailed comparisons between the theory and experiments are shown in table 1. The above discussion is true for noiseless film boiling which often happens when the hydrostatic head in Hells is small (typical example is a hydrostatic head less than 12cm[5]). From the photos of the boundary of GHe/HeII, it is seen that the film is very stable and there is no disturbance in the film, the film's thickness is unique along the cylinder. These factors demonstrate that the total mass transfer in the boundary of GHe/HelI is zero[6]. 25
20
Table 1 The effects of combined various pressure terms (hydrostatic pressure, Van-derwaals pressure and fountain pressure) on the peak heat flux density. p,gh At> t AI>t,,
qve
qt, lh
A
364
151
0
11.2
11.44
2.1
364
151
948
16.2
16.2
0.0
364
151 1673
18.0
17.5
2.8
(r.) 6".) (r.)
•
e,i
•
~" 10
2) (Wcm-') (%)
* qp~- experimental result, qpth-theoretical result, A - relative error between %~ and qpth
5
J
10 ~
,
J,.~.
,
i
10 4
,-~
lO s
P,Pa
Figure 3 Pressure effect on the peak heat flux density of HeII bath at 2.0K when P>P)~(x: experimental results with 511am diameter RhFe wire).
(b) Noisy fihn boiling Contrary to the noiseless fihn boiling state, there exists severe vaporization - condensation phenomena at the boundary, and the sound of boiling can be heard during experiment. A typical example is that for noisy film boiling phenomena, it can be found in heat transfer experiment in HelIs with a hydrostatic head more than 12 cm. The noisy phenomena will possibly disappear when the hydrostatic head goes much higher. For HeIIs at 2.0K, when HeII is slightly subcooled (for example, with a bath pressure of 0.004 Mpa), the noise from film boiling heat transfer measurement will not be heard. This pressure range, though not clearly defined, is just the transition region for heat transfer in HeII, however a quantitative equation of pressure effect on peak heat flux is difficult to show. Subcooled superfluid helium If HelI is slightly subcooled, and when the pressure P is less than or equal to P)., and P is greater than t~.(7}~)(the saturated vapor pressure at T~0, the total subcooled degree is t ' - t ' s + pgh when the wire is submerged in the depth of h in HeII. So the Clausius-Clapeyron equation gives 7' ATi~a~ . ( 1 ' - 1's + p g h ) (9) A,L
If P - l~s,>> pgh, then the peak heat flux density can be written in the equation as following .~ 20(r~,) 7' 1 qp
= . . . . . . r,, p,,L
(lO)
Even if 9gh tends to zero, APA becomes relatively important, the APA item can still be neglected in comparison to ( P - t~,). The applicability of the heat transfer equation in HelI can be tested at small subcooled degree under the condition of t's < P < P;, by eqn.(10), such as the experiments at 2.0K[7]. The experimental result is 15 W / c m 2, and the theoretical one is 15. I W / c m 2 when P is 0.0045MPa. So the relative error is only about 2 % , at this time O(r 0) is 0.029. Furthermore, the calculated peak heat flux q,, is 16.2 W / c m 2 at P=P)=0.005MPa, the experimental one is 16.6W/cm 2 the relative error is only about 2.5% PEAK HEAT FLUX DENSITY qp IN HEll WHEN P > t'~
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It is seen from the phase diagram of helium, that there are GHe, HeI and Hell layers when a wire is heated when the film boiling is initiated at a pressure of P>Px. When heated, the thin film of HeI will be formed around the wire if the heat flux is big enough. In another words, there will appear T=T~ at the radius of r=r0, however HeI will be vaporized in millisecond time because the heat flux is big enough and HeI around wire will change to GHe, GHe will be wrapped by HeI, HeII respectively. The author has fully discussed the calculation of the critical heat flux about the complex phase boundary in reference [ 1]. The fundamental condition needed to initiate film boiling of HeIIp is that there appear HeIIx (T~ = Tt ) at the position of wire/HeII boundary, and a temperature gradient (r = r0, T = T~;r = 0% T - T8 ) exists. When TB deviates from Tx greatly (T8 - T,t > 0.06K) in general, by integrating the Gorter-Mellink equation from TB to T~., we get
, 2~,'o_______~)IT~ dT
qp =
ro
r, f ( T )
(11)
Because the X-line in the phase diagram is a line nearly vertical to the T axis [when P=P~, T~,=2.172K, when P=0.1MPa, T~=2.163K], the effective heat transfer coefficient f(T) ~ (f(T,P) 1 in reality) varies little from the saturated pressure to 0.1MPa. It can be concluded from eqn.(11) that peak heat flux q~ varies little with bath pressure P from 0.005MPa to 0.1MPa. The experimental results of the peak heat flux q~ at 2.0K using RhFe wire with a diameter of 51~.tm when the bath pressure are 0.005MPa, 0.05MPa, 0.1MPa respectively have proved the deduction (shown in Fig.3). The modified factor ~(r 0) is of big difference to the one in eqn.(4) or (10) because of the influences of the boundary phenomena when the film forms and the transient process. From experimental results, we get q~Qb)=0.0709 and the theoretical line is displayed in Fig.3. The symbol 'x' represents experimental results. Thus we can see that the theory demonstrates accurately the relation between P and q~. CONCLUSION The above studies show that Van-der-waals pressure plays an important role in the heat transfer in HeIIs for a small hydrostatic head; Fountain pressure can be a very interesting pressure term for special experiment, which will influence a lot on the critical heat flux in HeII; Sometimes hydrostatic head, Vander-waals pressure and fountain pressure will combine together to influence heat transfer in HeII. In HeII with a small subcooled degree (P
Kapitza Conductance of Niobium for S. R. F. Cavities A. Boucheffa, M. X. Francois and J. Amrit L. I. M. S. I.- C. N. R. S., B.P. 133, Orsay, 91403, France Cedex The Kapitza conductance is measured at the niobium-helium II interface for temperatures ranging from 1.5 K to 2.1 K, using a new experimental method in which the heat flux is directed from the liquid to the solid. These experiments clearly show the effects of chemically polishing rough surfaces and the influence of surface oxides on the heat transfer process. A comparison with existing measurements and theories are also made.
INTRODUCTION The interest in the thermal boundary conductance (Kapitza conductance) between niobium and helium II arises from the fact that niobium has revealed to be an appropriate material in the construction of superconducting cavities for particle accelerators. Due to the Skin effect, the electromagnetic waves (1-3 GHz) present in the cavity penetrate into the wall of the cavity and dissipate energy by Joule heating. This leads to numerous undesirable effects like loss in the Q factor of the cavity, loss in the beam definition and possibly quenching. It is therefore necessary to understand the heat transfer mechanism between the Nb wall and its surrounding He II coolant if optimum functioning of these cavities are to be maintained. In the experiment, which is based on a novel technique, we have considered not only surfaces of Nb that were prepared according to the standard procedure for the S.R.F. cavities, but also rough and chemically treated samples of different bulk purity content given by the RRR values. EXPERIMENTAL Experimental Cell Configuration : a performant test facility In an attempt to overcome problems, like determining the temperature of the solid near the interface, encountered in the standard technique for measuring the Kapitza resistance, we developed a new experimental configuration. Some of the advantages of this latter are : (a). The heat flux in the Nb is nearly one dimensional since the ratio of its diameter to thickness is relatively high. (b). The absence of thermometers on the sample implies an undisturbed uniform temperature field. (c). An easy and non-destructive sample mounting and dismounting procedure. The experimental set-up and cell are shown in figure 1. The experimental cell is composed of a stainless steel cylindrical support having an external diameter of 80mm and an internal diameter of 40mm. Two niobium plates, each 50mm in diameter and 2mm in thickness, are mounted on either ends of the support by means of two stainless steel flanges which are bolted together, The niobium plates are identical in that they were made from the same bulk material and their surfaces were prepared under the same conditions. An indium seal between the flange and the sample assures a superleak tight cell. This assembly forms a cavity which is filled with He II (called internal bath) through a stainless steel capillary tube of length l m and an internal diameter of 0.2ram. The experimental cell is immersed in a 4He cryostat ; helium surrounding the cell shall be referred to as the external bath. Manganin wire wound on a crosspiece made of epoxy constitutes an electrical heater resistor Rheaterl which controls the internal bath temperature. The latter is measured with an Allen Bradley 100W carbon resistor (AB 1). The temperature of the external bath, also measured with an Allen Bradley 100W carbon resistor (AB2), is controlled to within +0. l i n k with the aid of a heater resistance Rheater2. The Allen Bradley resistance thermometers have a high sensitivity in the experimental temperature range, that is, dR/dT -- 10 4 W.K -1 at 2K. 559
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ICEC16/ICMC Proceedings
Experimental Technique This is based on the fact that He II has a very small thermal diffusion time constant and that the Nb sample is at the same temperature as that of the external bath (To). The experimental procedure consists in measuring, under equilibrium conditions, the difference in temperatures between the intemal (Ti) and the external baths for different heat fluxes Q dissipated by Rheaterl. By symmetry, the heat flux through each Nb sample is Q', where Q ' = (Q - Qloss)/2. The difference in temperature AT = Ti- To and Q' are simply related by : (SAT/Q') = RG = RK + RNb
(1)
where RG is the global measured thermal resistance having two contributions 9RK the Kapitza resistance (or conductance hK) at the Nb-He II interface and the thermal impedance RNb = e/K, where e and K are the Nb sample thickness and thermal conductivity. The Kapitza resistance thus takes the form"
(2)
RK = (SAT/Q') - (e/K) = 1/hK
Now, Qloss is due to heat losses via the stainless steel support, the capillary tube and the electrical heater and thermometer wires. However, with the help of a numerical simulation these heat losses are quantified, thereby reducing the relative error in Q' (dQ'/Q') to < 0.2%. The thermal conductivity of each sample was determined by F. Koechlin [ 1] prior to measurements. Sample preparation Firstly, a series of measurements were done on Nb samples having rough surfaces (that is, as received from the supplier) with different bulk purity content. Sample 1 (RRR-~ 40), which was supplied by Wah Chang, was laminated from a sheet to a thickness of 2 mm. Sample 2 (RRR ,~ 180), which was supplied by Hereaus, was prepared as sample 1. Sample 3 (RRR --- 270) was machined from a rod supplied by Wah Chang. Secondly, to study surface effects on the Kapitza conductance, sample 2 was chemically polished according to the standard procedure for S.R.F. cavities. Finally, the influence of purification by titanium was investigated (samples 5 and 6 with RRR --- 370). RESULTS .
.
.
.
.
.
..------
J-nn io , Ua,
- -
--~
. . . . . .
11
r
Capillary ~ Experimental cell Internal bath Heating resistance
I
i i i i 1
] :~2.- -~', I [ ~-'~;--r.-',,i ,, i ! "~K,.._._..) u- / \r~ ._~~
(b)
i i i i
i i i 1
i i I i
i i i i i I i i i~
1
~ _-,,*---~---
eq |
R.e~malationthermomete _ / (AB2)
r !
~~L
T ~--~" ; .~J
Stainless steel flange :_... Niobium plate
=r--
:: - " ~ " . " : " ' : ' - ' - " - : " L ' T ~ T M
~
~ ! ~~-----~---
--
...~ .
He filling capillary Cylindrical support
l ' ~ ~ ' - ~.~:."-i'~"~i.?LLLd:: ~ - '.'?L~..: ' ' -..~-"'.'. '. -. .-. . . .
Manganin heater
~----~-~--~'~..,,..l.
Niobium plate
i
!,!
..
l ~-~
; -"Stainless steel flange
r
9
o
0,1
~
"~
E-..........]..............! I t
"----
J
i i .,i
+ II
9
O
sample 6,7 sample 5 sample 3 ........ sample 1 sample 4 sample 2
~ 1 7 61
1,5 i
,
! i i
<>
.....
1,6
1,7
1,8
1,9
2
2
T(K)
I
Fig. 1 9Experimental set-up and cell
Fig. 2 9Kapitza Conductance of different Nb samples
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561
The experimental results of the Kapitza conductance between Nb and He II are shown in fig.2 for seven samples. The experiments were conducted at T < 2.1K and at saturated vapour pressure. The Kapitza conductance follows a T 3 behaviour for all samples as shown by the fits (solid lines), independent of their surface state. But, the magnitude of the conductance is between 25 to 50 times larger than the Khalatnikov's Acoustic Mismatch theory [2], shown by the solid line (a). The dotted line (b) represents the upper limit for hK as predicted by the phonon radiation model [3] for a perfect transmission across the interface. Effect of bulk purity. As seen in fig. 2 the bulk sample purity, indicated by the RRR value, does not have an apparent influence in the heat transfer process. For example, sample 3, which has a R R R value of 270, has a lower thermal conductance than sample 1, which has a RRR value of 40. This suggests that the discrepancy in the different sets of data arises mainly from the factors which influence the state of the solid surface, that is, its chemical and physical properties, and eventually the nature of the He close to the boundary. Influence of surface chemical polishing The effects of a chemical treatment were studied on samples 2 and 4 which are represented by open and full circles in figure 3. These two samples are identical, except that sample 4 was polished chemically as described in ref.[4]. Figure 3 shows that the heat transfer across a polished surface increases by a factor 2. Observations made in experiments by Mittag [5], whose results are also indicated in this figure (by crosses), show that not only heat transfer is enhanced, but a change in the temperature dependency may occur as well with polishing. The actuel heat transfer mechanism which is improved with polishing remains unknown. However, our results suggest a decrease in the reflection of the solid phonons at the boundary. Effects of Titanification Figure 4 shows the effect of surface impurities in heat transfer. Sample 5 (RRR -- 370) was coated with a layer of titanium. The Kapitza conductance of sample 5 (open diamonds) is close to that of rough surfaces. Now, approximately 3 mm from the surface of sample 5 was removed by chemical polishing (sample 6). We recall that surface oxides combine with Ti, a group 5 metal like Nb. Removing the Ti layer therefore
1,2 L .... 1
o
s
0,6
r'"
......
I''""
......
""l
....
_1
................................................................... ]........ ~ .............
i
!
i I - ........... i........~ !
~
i
i
.......i............. i.............. i. ~ . ~ " ~ i
i
.,X"
......
. . .........
"~
~
= 0
~ N
0,2
hd
1,4
f.t
6 0,4 r,.) "~
"
...... !iT............ i ..... -....... ;.~"~x--'~. - ' ' x 7 ' - ' i .............
' - ~ - X"- ~K- -~- -X- -.,~- XI~o~gh Nb s~mple (rcf.6) 0
....
1,5
i ....
1,6
!!
1,6
i ....
1,7
i ....
1,8
i ....
1,9
T(K)
i ....
2
00,6 ,6 4 0,4
0,2,2
!!!iiiii
1,5
i ....
2,1
00,8 ,8
2,2
Fig. 3 9Kapitza Conductance after chemical treatment
1,6
1,7
1,8
1,9
T(K)
2
2,1
2,2
Fig. 4 9Influence of purification by titanium
eliminates the presence of oxides on the surface of sample 6. The Kapitza conductance of sample 6 (solid diamonds) is a factor 2.2 larger than that of sample 5. Also, this conductance is about 25% larger than a sample which is only chemically polished. Another 45 mm (sample 7) was removed from the sample 6 surface and the Kapitza conductance remained unchanged in comparison with sample 6. This suggests once again the influence of surface boundary on the Kapitza conductance. Further, from a technological point of view, removal of a few mm of the surface after titanification suffices to improve the heat transfer.
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ICEC16/ICMC Proceedings
CONCLUSION The experiment shows that heat is evacuated more easily across a Nb-He II interface than expected on the basis of the Khalatnikov theory. The results suggest that this is due to the nature of the boundary rather than bulk properties of each medium. The Kapitza conductance for rough surfaces is of the order of O.057xT3 (W.cm-2K-1). The conductance is a factor --2 larger for surfaces that underwent a chemical treatement. Finally, for samples that were treated for impurities by a titanification process, the Kapitza conductance is of the order of O.15xT3 (W.cm-2K-1), that is, heat is removed at a rate of 0.86W.cm-2.K -1 at T = 1.8K, the optimum temperature of a S.R.F. cavity. REFERENCES
,
,
4. 5.
Koechlin, F., Proceeding of the International Conference on Superconductors, ICMAS-90, Grenoble, France (1990) Khalatnikov, I. M." Introduction to Superfluidity, W. A. Benjamin, Inc., New York, New York (1965) Snyder, N. S., Heat transport through helium II : Kapitza conductance Cryogenics (1970) 10 89-95 Boucheffa, A., Th~se de l'UniversitE Paris 6 (1994) Mittag, K., Kapitza conductance and thermal conductivity of copper niobium and aluminium in the range from 1.3K to 2.1K. Cryogenics (1973) 10 94-99
Thermal Behaviour of Electrical Multilayer Insulation Permeable to Superfiuid Helium B. Baudouy t , A. Boucheffa*, M.X. Francois* and C. Meuris CEA/Saclay, DAPNIA/STCM, F-91191 Gif-sur-Yvette Cedex *CNRS/LIMSI, B~tt. 502 ter, Campus universitaire, F-91405 Orsay
Electrical multilayer insulations made of Kapton | tapes and prepreg or adhesive Kapton ~ tapes, used in dipole magnets, may offer a complicated arrangement of thin helium channels which cannot be easily predicted and modelled. Several insulation systems have been tested in order to characterize their helium channels pattern. Heat transfer data analysis shows clearly the contribution of supeffluid channels inside the insulation. Appearance of the vortex-free regime for very small temperature differences (10 .5 to 10.3 K) and of the Gorter-Mellink regime for higher temperature differences allows to establish the mean value of the channels diameter. We present in this paper the thermal behaviour of several combinations of insulating materials with different geometrical arrangements and porosities.
INTRODUCTION The research and development program for the Large Hadron Collider dipoles developed at CERN includes stability studies which are carried out in collaboration between CERN and CEA/Saclay. For NbTi magnets cooled by superfluid helium the most severe heat barrier comes from the electrical insulation of the cables. This paper reports a work which is part of the thermal study program. It deals with the intrinsic qualification of different insulation systems. Their global thermal performance in the surroundings of the winding is studied with a different experimemal model [ 1]. Classically, an insulation is a composite made up of a first wrap, a tape wound around the cable for electrical insulation, and a second outer wrap protecting mechanically the inner wrap, creating helium channels and gluing to the next conductor to keep the coil in shape. In the tests described, wraps are reproduced as plane layers. One (samples B22 and B23) or two (sample B25) sublayers of 11 mm wide Kapton | tapes with an overlap of 50 % are used for the first layer. For the second layer, adhesive Kapton ~ tapes, 12 mm wide with a spacing of 2 mm (B23 and B25) or 4 mm (B22), are employed.
THERMAL QUALIFICATION OF INSULATION Two 50.3 cm 2 wide insulation samples are clamped between two isothermal baths of He iI, each of them being able to reach different temperature. A detailed description of the experimental set-up is given in [2] and summarized in figure 1. Temperature measurement is made after attaining a stationary temperature for both baths, the outer one being regulated and held constant for the whole test over the range of power dissipation values in the inner bath. The difference in temperature between the two baths is plotted as a function of the generated power. This curve, of which an example is given in figure 2a, characterizes the overall thermal resistance between the heated bath and the cold bath. The measured resistance includes the Kapitza resistance, the resistance of the insulating material and the resistance of the helium paths through the insulation samples. Tests have been performed at Saclay in a pressurized He II cryostat. Preliminary results reported in [2] have shown that heat transfer is influenced by He II heat transfer even for small volume of helium inside insulation. At low heat flux, heat transfer is purely governed by He II. Fits in this range of heat flux Doctoral fellow CEA-Jeumont lndustrie-CERN 563
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ICEC16/ICMC Proceedings
have allowed to determine an equivalent geometrical factor A/L1/3 which characterizes a permeability of the medium related to He II heat transfer in a net of narrow channels (equivalent cross-sectional area A, length L) opening on either side the insulation. For higher heat flux, a model which considers insulation material and He II thermal paths in parallel agrees with measurement within 10 %.
Figure 1 Schematic diagram of insulation and experimental setting. Dimensions are in mm
EXTENSION OF MEASUREMENTS TO VERY LOW TEMPERATURE DIFFERENCES In another test facility at LIMSI, with a very precise temperature controlled saturated He II bath, using a resistance a.c. bridge equipped with a lock-in null detector, it has been possible to measure temperature difference as small as 10 laK across the same insulation samples. Figure 2b shows an typical example of the inner bath temperature rise AT, for different temperatures of the outer cold bath Tb.
Figure 2 Temperature difference between the two He II baths as a ftmction of the power dissipated m the inner bath for various outer bath temperatures. Sample B22. (a) : range [1 mK, AT~ = T~ -Tb]. (b) : small AT
MEASUREMENT ANALYSIS Heat is transported in He II according to the movement of the normal fluid from the hot source towards the cold source. The motion of this viscous fluid undergoes a resistance whose magnitude varies according to the associated Reynolds number and the possible presence of vortices in the superfluid component.
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Very often the heat flux densities q = Q/A are such that the double convection velocities v = v, v~ exceed the thresholds for the formation of vortices. The mutual friction generated becomes thus dominant and leads to a transport law known as the Gorter-Mellink regime [3,4], IqlZTq--f(T)VT, where f(T)-
(9~3s4Ta)/(A(T)9,) with the notations used by the authors stated, that leads to"
Q3_ ~
~f(T)dT L
(1)
In media of small dimensions, the critical velocities for the formation of vortices become sufficiently high such that the corresponding critical flux densities are non-negligible [5]. This gives a small range in which heat is transferred by a normal fluid without any interaction with the superfiuid component. The Landau two-fluid model without interaction [6], leads to the transfer law V p - p s V T , with ~'p - -(12g./d2)'~ for narrow slits of thickness d in the Poiseuille regime, that is" Q - (Ad2)(ps)ET 12~t~ AT
(2)
The analysis of the experimental results, of which an example is reproduced in figure 3, indicates : 9 the importance of the He II contribution to heat transfer by an estimation of the purely conductive part through the solid structure, Q~ol, from measurements of the conductivity and of the Kapitza resistance, which were done in the same experiment cell on samples of plain Kapton | foil 9 the possible existence of a laminar Landau regime 9 in the affirmative case, the determination from equation (2) of the geometric factor Ad2/L - X~' Aidi2/L, equivalent to the totality of n micro-channels 9 the verification of the hypothesis by the temperature independence of this geometric factor 9 an estimated value of the critical heat flow Qr and the corresponding superfluid velocity v~ 9 the Gorter-Mellink regime by a research of a zone (in which Q~o~is ever lower than 0.1 Q) having a Q3 dependency and, within the precision of fiT), a value of the geometric factor A/L v3 - E~' A~/L~ v3, after equation (1), and finally, the verification of its non-dependence with temperature.
Figure 3 Different heat transfer regimes. Sample B22. The solid curves are the best fits (including measurement errors) to the data using equations (1) and (2). The dashed line is the estimated purely conductive part Those different results, extracted from the AT(Q) curve of sample B22, are reported in table 1. We note that the geometrical coefficients corresponding to the two regimes stated above are almost constant in a large temperature range. It is possible to deduce values of A and d, that is A = 5.74 mm 2 and d = 2 2 ~tm,
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ICEC16/ICMC Proceedings
by making the hypothesis that the He II slits are of length 5.5 mm corresponding to the overlapping of the tape of first insulation layer, and by supposing identical dimensions for all the parallel slits. The critical flux densities QdA vary from 0.32 W/cm z to 0.68 W/cm 2 when the temperature of the bath varies from 1.7 K to 2.0 K. The critical velocities (superfluid values) vary from 1.0 cm/s to 3.1 cm/s and the associated Reynolds numbers from 156 to 100. We thus verify that the normal fluid flow is laminar in the Landau regime. Table 1 Equivalent geometric factors of the insulation sample B22 Temperature Ad2/L 10"14m 3 A/L 1/3 10"5m"~/3
1.7 K 53 + 51 3.29
1.8 K 53 + 7 3.20
1.9 K 49 _+6 3.25
2.0 K 41 _+9 3.17
2.05 K
3.25 1 The precision includes the precision in the measurementsand the precision in the fits of Landau and Gorter-Mellink zones. The preceding results obtained for different insulations are summarized in table 2. For sample. B25, the slits are very narrow, the wall friction of the normal fluid induces a non-negligible temperature gradient which does not allow a good identification of the Gorter-Mellink regime. We observe that the results of B22 and B23, for which L is a priori identical, brings to front a factor 2.6 on A which is coherent with the ratio 2 on the spacing of the second layer tape and the fact that the thickness d of the slits also decreases when the spacing decreases. The doubling of the first layer (B25) can be understood by a strong reduction of the effective permeability, which we must attribute not only to the increase of the channel lengths but equally to a decrease of d.
Table 2 Characteristic geometric parameters of the He II slits of different insulation systems Sample first layer Kapton| spacing of the number tape (50 % overlap) second layer tape 2 B22 150 HN (38 lam) 4 mm B23 150 HN (38 lam) 2 mm B25 2 x 100 HN (25 lam) 2 mm 2 Kapton| 270 LCI tape (adhesive, 68 lam)
Ad2/L
10"14m3 49 7.1 0.18
A/L1/3
10"-~m~/3 3.25 1.19
(A/L 1/3)/mtot
10-3m"1/3 3.23 1.17
CONCLUSION The results and the method of analysis presented allow a detailed qualification of insulations by determining characteristic geometrical magnitudes of thermal properties and, in particular, that of parameter d, the thickness of slits which can only be attained experimentally.
ACKNOWLEDGEMENTS The authors are grateful to Dr D. Leroy and Dr B. Szeless from CERN for helpful discussions. Special thanks to Mrs A.M. Puech and Mr Gaubert for the sample preparation and measurements.
REFERENCES 1 Burnod L. et al., Thermal modelling of the LHC dipoles functioning in superfluid helium, Proc. EPAC C one (1994) 2 Baudouy B. et al., Steady-state heat transfer in He II through porous superconducting cable insulation, Proc. CEC-ICMC Conf. (1995) 3 Vinen W.F., Mutual friction in a heated current in liquid helium II, Proc. Roy. Soc. A240-A243 (1957) 4 Gorter C.J. and Mellink J.H., On the irreversible process in liquid helium II, Physica 15 (1949) 5 Arp V., Heat transport through helium II, Cryogenics 10 (1970) 6 Landau L.D. and Lifshits E.M., Statistical Physics, Pergamon, Oxford, UK (1958)
Pressure Gradient Caused
by
Quantized
Vortex
in
Superfluid
Helium
Minoru Yamaguchi, Yoshiko Fujii, Masaki Nakamura, Toshinobu Shigematsu and Toyoichiro Shigi Dept.of Applied Physics, Okayama Univ.of Science Ridai-cho,
Okayama
700,
Japan.
The temperature difference and the pressure difference through He II in a capillary glass tube have been measured very precisely.
The quantized vortex
line density was calculated from the temperature difference data using the numerical scaling coefficients.The pressure difference data give the information of interaction between the vortex line and the tube wall.
The pressure
dissipation caused by the vortex line was only one tenth of mutual friction between the vortex line and the normalfluid.
In the terms of eddy viscosity for
superfluid velocity field, we have calculated the coefficient of the interaction between the vortex line and the superfluid. INTRODUCTION A great deal of experimental and theoretical studies on superfluid turbulence have been performed [1],[2],[3],[4],[5], however, the influence of the vortex line on pressure dissipation is still in question. To study this problem, we have measured the temperature andpressure difference across the glass capillary tube in which the thermal counterflow was built up. EXPERIMENT Fig. l shows schematically the arrangement of the experimental cell to carry out the temperature and pressure difference measurements. Heat supplied in lower part of the cell produces thermal counterflow in a capillary glass tube.
During a series of measurement,
the temperature of upper part of the cell was always maintained constant within
I x 10SK.
The temperature difference across the tube was measured with the accuracy of 1 p K using a thermocouple (Au[0.03at.%Fe]-NboTi) connected to a SQUID detector.
Tow membrane type pressure
transducers were used to measure the pressure. The 8 ~ m-thick phospor bronze membrane and the transducer body form a electric capacitor.
Fig. 1 Schematic arrangement of this experiment. 567
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ICEC16/ICMC Proceedings
The value of pressure is obtained from the resonant frequency(f) of a tank circuit connected to a tunnel diode BD-5.
The stability of
able to be detected.
NO. 1 2 3 4 5
was about 106 , so that the pressure change of
1 • 10-2 Pa was
The size of the tubes used in this experiment is listed in Table 1.
Table tube
f
1
Table
d (x 10-2cm) 1. 3 0 I. 2 6 i. 8 0 1. O0 2. 6 6
1 (cm) 9. 6 9. 6 9. 5 9. 7 9. 5
T (K) 1. 3 i. 4 i. 5 1. 6 1. 7 1.8 1.9
2
(Ref.[7])
CL 0. 080 O. I 0 0 O. i18 O. 1 3 3 O. 1 5 4 0.174 0.195
(I ,I--CLI1) 0.710 0. 710 0. 713 0.719 0.731 0. 742 O. 7 5 5
RESULTS AND DISCUSSIONS The temperature differences (AT) as a function of the heat flux (W) for tube NO.2
are shown in Fig.2.
If the heat flux is less than W~, AT is expressed by a following relation,
128r/.1
AT
w
nd
(])
7'
where r/,, is the viscosity coefficient of the liquid helium, 1 is the tube length, d is the inner diameter of the tube, p is the density of liquid helium, S is the entropy density and T is the temperature of the liquid helium respectively.
Hence, in this state the normal component of liquid helium is considered
to be the laminar flow. In the turbulent state, the temperature gradient across the tube V T is represented as follows[6],
VT-
VI"t = ~ P.fi
where
,
(2)
t~[,, = p.~xct(Ill - c t I , ) L V
VT~ is caused by the contribution of the normalfluid in the laminar state, F,, is the mutual
friction force between the vortex line and normalfluid, p.~ is the superfluid density, K is the quantum of circulation, a is the interaction coefficient, L is the vortex line length density, V is the relative velocity of two fluids. (I It - c t I~) is the parameter of the dynamical scaling method. L was calculated from f l L v2
VT
data using the numerical value for (Ill-crib), [7].
Fig.4 shows the relations of V and
for tube NO. 1, where fl is represented as follows [6],
13 = (
) In(aLl/2 )
( where,
a is the radius of a vortex filament ).
In the well developed turbulent state f l L ~/2 can be represented as follows, pL !/2 - -
CL2
V
The coefficient CL2 obtained
(3) from the experimental data is shown
The value of ci. 2 is independent on the tube diameter. vortex line length.
for tube NO. 1,2,3,4,5
in Fig.5.
Hence, we should consider that L is the net
ICEC16/ICMC ICEC 16/ICMC Proceedings Proceedings tube N0.2
0
I
I
569 569
I
~..*. .*
A
100
0
W
200
( p watt)
Fig2 Temperature Werence A7' aB a
function of the heat flux W for tube N0.Z.
n 1
0.2 N A.
I
0
v
t
2Q
10
l c m - s-€1
Fig4 The vortex Pine density
3
for tube NO.1.
an a function of Y
t
n
0.1
"
1.2
1.6
T m
1
3
(K)
F i g 5 Coefficient cLZ as a function
c
of the temperature for the tube NO.1,2,3,4,5.
v
(cm-
h
s-1)
Fig.6 The i n d u d premure gradient 8s a function
.
Icr
.
p PPr
of Y for tube N0.1.
.
i
o
a
e
o
,
Q.001
E .6 20 T IK3 Fig.8 T h e parameter & as a function of f .2
the temperaturefor tube NO.1,2.3,4.6.
as
CX103
* $">
Fig.7 V f ; aa a function of PI4 for tube N0.1.
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ICEC16/ICMC Proceedings
The pressure differences (AP) are shown in Fig.3 for tube NO.2. In the turbulent state, the pressure gradient across the tube may be represented as follows,
(4)
V P - Vl"L oc C~psM~V
where VPL is the contribution of the normal fluid in the laminar state and c~z is any coefficient of the interaction between the vortex line and superfluid. For the experimental data., the coefficient c a does In this experiment, flVP r increases in proportion to V 2 as shown in
not remain in a constant value. Fig.6(where VPr = V I " - VP s ).
Then, we assume that the pressure gradient is given as follows, 32r/, VI~. = - ( d 2 )zV,, where y is a characteritic parameter.
(5)
The form of r/, may be assumed as follows[2],
r/~ = p~x~(d)L ';2
(6)
where ~ is a fitting parameter.
Then,
by means of (6), we have the following results.
VP r = -( 16ps2tc~Y )vL';2
(7)
pa
and when we substitute eq.(3) for V of eq.(7) [V/r [ = (16p.,2tc~y fl) pd )'( L CL2 VPr as a function of pL is shown in Fig.7. viscosity derived from VPr combined parameter ~y is Then
16p,~:
(pdLi, )
L~/2=2000 (cm-~),
(8) The combined parameter ~:y
data is shown in Fig.8. 3.5 • 103.
In this ~y,
For the representative resulut (in if
1.7K), the
we take Y' =0.5, then we get ~:=7.0 • 10-3 .
may be compared to a (I II --CLI;)we can calculate as follows,
of superfluid eddy
When
y =0.5,
~ =7.0 • 10-3
and
16p~ (pdLiJ2)=0.002.
This value is about one tenth of the value of a (I)l -cs REFERENCES [ 1] Vinen,W.F.,Proc.R.Soc.London ser.A242(1957). [2] Brewer,D.F.and Edwards,D.O.,Proc.R.Soc.London,ser.A251 (1959)247;Phios.Mag.6(1961)775;6(1961) 1173;7(1962)721. [3] Childers,R.K.and Tough,J.T., Phys.Rev.B 13(1976) 1040-1055. [4] Martin,K.P.and Tough,J.T., Phys.Rev.B27(1983)2788-2799. [5] Tough,J.W.,Superfluid Turbulence,Progress in Low Temp.Phys.,edited by D.F.Brewer(North-HoUand), (1982)8,Chap.3. [6] Donnelly,R.J., Quantized Vortices in Helium II, Chap.7 (p.215-254), (Cambridge Univ. Press.) [7] Schwarz,K.W., Phys.Rev.B38(1988),2398.
Heat and Mass Transfer between Two Saturated He II Baths X. Huang, J. Panek, and S.W. Van Sciver National High Magnetic Field Laboratory, 1800 E Paul Dirac Dr., Tallahassee, FL 32306, USA The present paper examines the heat and mass transfer processes between two saturated He II baths. The two baths, formed from two stainless steel cans of 50 mm ID and 660 mm length, are connected at the top and the bottom with two 100 mm long tubes. The bottom tube, with a 5 mm ID, is filled with liquid He II; the top tube, with a 1.3 mm ID, is filled with saturated He vapor. Heat transfer between the two baths is then governed by the counterflow process in the bottom tube and vapor mass transfer process in the top tube. Steady state and dynamic models, based on energy and momentum equations, are presented and agree very well with experimental results. Our study also demonstrates that the mass transfer process is a far more efficient heat transport mechanism than the counterflow process in vapor/He II two-phase systems.
INTRODUCTION A new cooling scheme using two-phase vapor/He II has been proposed for the superconducting dipole magnet strings in the Large Hadron Collider (LHC) [ 1]. The objective is to absorb the heat generated at 1.9 K in the LHC with minimum temperature rise by utilizing the latent heat of liquid He II. Such a cooling scheme, however, suggests two-phase flow and heat transfer in saturated He II over the whole range of vapor qualities. Unfortunately, despite preliminary experimental tests performed at CERN demonstrated the feasibility of such a cooling scheme[2], little has been know about the basic heat transport mechanism in two-phase vapor/He II system. To better understand the basic heat transport mechanism in a two-phase He II system, we conducted analytical and experimental studies on the heat and mass transfer between two saturated He II baths. Because the counterflow heat transfer in the liquid and mass transfer in the vapor are decoupled in this problem, we were able to evaluate the two transport processes individually and thus identify the dominate mechanism in the system. In the present paper, we present the analytical solution to the steady state process as well as the numerical modeling of the transient process of the system. Experimental results are shown to be consistent with the predictions based on the theoretical models. EXPERIMENTAL APPARATUS A schematic of the experimental apparatus is shown in Figure 1. Two stainless steel cans, 5.0 cm in ID and 66 cm in length, are connected at the top and the bottom with two 10 cm long ss tubes. The top tube has a 0.13 cm ID and the bottom tube has a 0.5 cm ID. The two cans are installed inside a vacuum can which isolates the two cans from the He I bath. Liquid helium is supplied to the cans through a JT expansion valve from the He I bath. A 56 cm long coaxial capacitance type liquid level meter is used to monitor the He II level inside the cans. Two germanium resistance thermometers, calibrated in situ against the helium vapor pressure, are placed inside each can. Analysis of the calibration suggests the error associated with the temperature measurements is less than 1.0 mK. A 170 ~ resistive heater, located at the bottom of the left side bath (the warm bath), is used to supply either steady state or transient heating to the warm bath. The right side bath (the cold bath) has a 25.4 mm ID pump-out line at the top connected to a high capacity vacuum pump through a control valve. Differential pressure across the two He II baths is measured with a variable reluctance differential pressure transducer. The experiment procedure begins with filling both baths with liquid helium to about half the height of the cans. With the JT fill valve closed, the vapor pressure inside the He II baths is then regulated to a prescribed value. In a steady state experiment, a constant power is supplied to the warm bath resulting in temperature and pressure increases in the warm bath. The pressure increase in the warm bath then forces liquid to flow to the cold bath until the difference in the hydrostatic head between the two baths offsets the vapor pressure difference between them. Data are taken when the temperatures in both baths reach steady state. For the transient experiments, because of difficulties in maintaining a constant vapor pressure in the cold bath during a transient heating, the control valve is 571
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closed and constant volume is maintained. The dynamics of this closed volume system is then studied by depositing a square heat pulse of 1.0 s duration to the warm bath and recording the temperature, pressure, and liquid level traces. ANALYTICAL MODELING Steady State Process Because of the extremely high effective thermal conductivity of He II, it is assumed that the temperature of each He II bath is uniform and equals to the saturated temperature at the corresponding vapor pressure. The steady state heat transfer between the two baths is governed by the counterflow in the bottom tube and the mass transfer in the top tube. For turbulent vapor flow in the top tube, the equations are given as
Qc =
f l (T) dT
/'
(1)
and (Ps(Th)- ps (Tc))pgdg 1"25/ 4/7 0.158L//0.25
Q - Qc = Aghfg
(2)
where Q is the total heat deposited in the warm bath, Qc is the heat conducted away by the counterflow process through the bottom tube, ps(T) is the saturated vapor pressure at temperature T, Af is the cross sectional area of the bottom tube, Ag is the cross sectional area of the top tube, L is the length of the tube, hfg is the. latent . heat .of evaporation, pg is the average vapor density, dg is the top tube inner diameter, and # is the vlscosxty of the vapor. Transient Process During a transient process, the changing temperature and pressure in each bath result in liquid level changes in each bath. Consequently, the energy transfer between the two baths not only involves transient Gorter-Mellink conduction and vapor mass transfer but also mass transfer by the liquid in the bottom tube. The energy balance for each bath during a transient process can be written as
]
~- pfVfiefi + P g i ( V - Vfi)egi = Qi -+ riaghg _+rhfhf _+Qc
(3)
where the plus sign is for the cold bath and the minus sign is for the warm bath. The subscripts f and g designate liquid and vapor, respectively. V is the volume of the can, Vfi is liquid helium volume in the bath, e is the specific internal energy. Qi is the external heat deposited in the bath, Qc is the energy transfer by the transient Gorter-Mellink conduction. The vapor flow in the top tube and the liquid flow in the bottom tube are governed by the one dimensional continuity and momentum equations cgp t- cg(pu) _ 0
&
anu
(4)
0x
~u
2Cf PU2 _ K p u 2 _ ~
P'&- +PU~x = -
d
(5)
where Cf is the frictional factor for the fluid, d is the tube diameter, and K is the coefficient for the entrance effect. The transient counterflow heat transfer for the warm bath is
~]1/3 (6)
Qc = A f [ f l ( T ) ~
x=0 For the cold bath, the derivative should be evaluated at x=L. To evaluate Eq.(6), the temperature profile along the bottom tube is needed. This temperature profile can be obtained by solving the He II energy equation,
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o,-tr rhfCp onrr Cp oqt _--4
A f p f c3x
3 i rl(T) _~]1/3 =0
oax
(7)
where9 C P is the specific heat of liquid He II. The final equation needed is the mass conservation equataon for each bath,
]
~- pfV fi + Pgi (V - V fi ) __.rhg 4- rhf = 0
(8)
where the plus sign is for the cold bath and the minus sign is for the warm bath. The initial conditions for these equations are
Tc=Th=Tinit, rilg -- rilf = 0, Wfi -- Vinit , at t=0. where Tinit is the initial He II bath temperature and Vinit is the initial He II volume. RESULTS Figure 2 displays the steady state relationship between the heat input and the temperature difference between the two baths at 1.9 K along with calculations based on Equations (1) and (2). Two important conclusions can be drawn from this figure. First, although the upper tube has a flow area about 15 times smaller than that of the bottom tube, the heat transferred by the mass-transfer process through the top tube is still greater than that by the counterflow process over majority of the heat input range. This clearly suggests that the mass transfer process is a far more efficient heat transfer mechanism than the counterflow process. Second, the experimental data agree very well with the model prediction especially at power input below 0.9 W. Figure 3 plots the temperature traces for the warm and cold baths after a 10 W square heat pulse of 1.0 second duration was sent to the warm bath. The initial liquid helium level is 37.5 cm and the initial bath temperature is 1.790 K. The curves are numerical solutions to Equations (3) through (8) with helium propertiesevaluated by HEPAK. During the initial heating, the temperature in the warm bath increases rapidly and reaches the maximum at t=l.0 s when the heater is switched off. Subsequently, the warm bath starts to cool down while the cold bath continues to warm up until both baths reach the same temperature at t = 15 s. Figure 4 shows the measured liquid level in the warm bath as a function of time along with the numerical prediction. When the transient heating is initiated, the rising pressure inside the warm bath forces the liquid to flow to the cold bath and thus the liquid level drops. At about t=5 s, the difference in the hydrostatic head equals to the vapor pressure difference between the two baths and the liquid flow reverses. When the two baths finally reach the same temperature at t -- 15 s, the liquid flow also stops and the liquid level in the warm bath almost returns to its initial value (the amount of liquid evaporated by the transient heating is negligible). CONCLUSIONS We have demonstrated the behavior of two-phase He II/vapor heat and mass transfer in an idealized experimental configuration. The steady state result clearly suggests that the vapor mass transfer is the dominate heat transfer mechanism in two-phase He 13Jvapor systems. The dynamic model presented in the paper agrees very well with the experimental results. Our work continues with a study of the fully coupled horizontal two-phase He II flows. ACKNOWLEDGMENT The authors wish to thank Reda Daher and Vincent Cochran for their valuable technical support. This work is supported by the U.S. Department of Energy - Division of High Energy Physics under grant DOE-FG02-96ER40952 REFERENCES 1. 2.
P. Lebrun, Superfluid helium cryogenics for the Large Hadron Collider Project at CERN, Crvozenics(1994) 34 1-8 A. I3ezaguet, et al., The superfluid helium model cryoloop for the CERN Large Hadron Collider, Adv. Cryo. Engr. (1994) 39, pp 649
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Figure 1. Schematic of experimental apparatus
Figure 2. Steady state heat input versus bath temperature difference at 1.90 K
Figure 3. Temperature traces of the warm and cold baths after a square heat pulse. power=-10.0 W, duration=l.0 s.
Figure 4. Liquid level in the warm bath after a square heat pulse to the warm bath. power=-10.0 W, duration=l.0 s.
Measurement of Characteristic Time of Quantized Vortex Development Using a Thermal Shock Wave
Takeshi Shimazaki and Masahide Murakami Institute of Engineering Mechanics, University of Tsukuba Tennoudai 1-1-1, Tsukuba, Ibaraki 305, Japan
The characteristic time of quantized vortex development is experimentally measured by using a propagating thermal shock wave. A thermal shock wave generated by a pulsed heating from a planar heater is deformed through the interaction with quantized vortices. The deformation is measured by means of a superconductive temperature sensor. It is found tha, t the characteristic time is inversely proportional to the square of applied heat flux. It is also found that appreciable effect appears on transient heat transfer if the vortex line density exceeds approximately 105cm/cm a.
INTRODUCTION He II is regarded as an excellent coolant for such as superconducting magnets. However, there are still several unsolved open questions concerning highly transient heat transport in it. It is indispensable to make those questions clear in order to understand the heat transport properties for the practical applications of He II. One of the most important and complicated questions is how the quantized vortex lines behave. When the relative velocity between the superfluid and normal fluid components exceeds certain critical value, quantized vortices are induced and the dissipative effect due to the interaction with quantized vortices get to be appreciable on heat transport phenomena. For steady or quasi-steady cases, the approach introduced by Gorter and Mellink[1] is usually used to take into account the effect of quantized vortices on heat transfer and is believed to give fair results both qualitatively and quantitatively. It is, however, known that the approach loses the validity for highly transient cases. In the cases that the time scale of thermo-fluid dynamic phenomena becomes comparable to or shorter than that of the characteristic time of evolution of quantized vortex lines in such as highly transient cases, their development and decay should be taken into account. The development and decay of quantized vortex lines were first successfully formulated by means of phenomenogical approach by Vinen[2,3,4]. The Vinen vortex line density equation has been widely used in many investigations[5,6]. In this study the characteristic time of quantized vortex development, defined from the point of view of a transient heat transfer, is derived by analyzing the thermal shock wave profiles measured with a superconductive temperature sensor. The result is also compared with the numerical result based on the vortex line density equation.
EXPERIMENTAL SETUP AND PROCEDURE Whole measurements are carried out in He II under the saturated vapor pressure condition. Figure 1 shows the main experimental apparatus itnmersed in He II. It consists of three ma.in parts, a superconductive temperature sensor, a. planar Ni/Cr thin film heater and a cylindrical 575
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side wall or a shock tube[7]. The distance from the hater to the sensor can be varied between 0.1 to 150 m m with 0.1 turn step. The sensing element of the sensor is superconductive thin metal film consisting of gold and tin fabricated on to a side wall of quartz fiber (1.5 m m in length and 40 #m in diameter). Constant current is applied to the element and the temperature variation is measured as an variation of voltage drop across the element due to the superconductive transition. Trapezoidal current pulse is applied to the Ni/Cr thin film heater which is 27 m m x 27 mrn and 400 ~ in thickness to generate a thermal shock wave. The cylindrical side wall, 150 rnm in length and 25 turn in inner diameter, maintains the one dimensional character of a traveling thermal shock wave inside the wall. The characteristic time of quantized vortex development tv~ is obtained by analyzing thermal shock wave deformation due to the interaction with quantized vortices. If the quantized vortex line density is not sufficiently high, a thermal shock wave profile changes only as a result of hydrodynamic nonlinear effect, that is to say the formation of a shock wave and the deformation into a right-angled triangular form. The applied heat from a heater is wholly transported by a thermal shock wave[7]. On the other hand, if the vortex line density becomes high enough, dense quantized vortices bring about dissipative effects on the heat transport, and a partial declination appears in the late portion of the plateau of a thermal pulse as illustrated in Figure 2. The time interval, t~, from the wave front to the point at which the declination of plateau becomes appreciable is defined as the characteristic time of quantized vortex development from the point of view of a transient heat transfer. The point of onset of declination is practically defined as the instant at which the deviation exceeds 3 cr of the data fluctuation.
RESULTS AND DISCUSSION Figure 3 shows typical measured thermal shock wave profiles for four cases of the heat flux %. It is seen that the point of declination onset gets closer to the wave front and the deformation becomes larger as qp increases. Figure 4 shows the measured t ~ as a function of % in double logarithmic plot for two cases of temperatures, 1.70 K and 1.90 K. The solid and broken lines, of which inclinations are -1.9 and -2.0, respectively, represent the linear regression results. It may be concluded that tv~ can be approximately related with qp in the following equation.
t~-
C ( T s ) q ; 2,
(1)
where C ( T s ) is the temperature dependent coefficient given experimentally. The results are also compared with the numerical result based on the Vinen vortex line density equation given by
dL L3/2 d--t- = a Iv,~[ - /3L2 + 7
Iv. l 5/2 ,
(2)
where v~ is the relative velocity between the superfluid and normal fluid components, L is the vortex line density defined as the total vortex line length per unit volume, a and /3 are the growth and decay coefficients given by Vinen, 7 is also given by Vinen[4]. The third term of the right hand side is the source term which is usually neglected. However, Kanari et al. [8] reported that the term is indispensable especially in the case of strong heating. Figure 5 shows the time evolution of quantized vortex line density obtained numerically by solving the equation without neglecting the third term for three cases of the applied heat fluxes, where v~, is given as a function of applied heat flux %. It can be seen that the final density reaches higher value and the development becomes faster for larger %. It is easily seen that the time duration which is required for the vortex density to reach certain value L is a function of heat flux. The time duration in which L reaches a number of values, from 103 to 107cm/cm 3, is plotted in Figure 6, where experimental data. a,re also plotted for comparison. It is seen that the inclina,tion of the regression result va.ries fl'om - 5 / 2 for the vortex line density of 103cm/cm :3
ICEC16/ICMC Proceedings to - 3 / 2 for 107cm/cm 3. The experimental data are found to reasonably follow the line of L = 5 x 1 0 4 c m / c m 3 of which declination is - 2 . It can be concluded from the figure that if the vortex density exceeds approximately lOScm/cm 3, transient heat transfer will be appreciably affected by quantized vortex lines.
CONCLUSIONS The characteristic time of quantized vortex line development is experimentally measured frolla the deformation of thermal shock wave profiles. It is found that the characteristic time t~, can be related with the applied heat flux qp as t~, - C(TB)q~ 2. It is also found that when the vortex line density exceeds the order of lOScm/cm 3, the transient heat transfer get to be affected by quantized vortex lines.
ACKNOWLEDGEMENTS This research was partly supported by JSPS (Japan Society for the Promotion of Science)
REFERENCES
1 Gorter, C. J. and Mellink, J. H. , On the irreversible process in liquid helium II. Physica (1949) 15 285-305 2 Vinen. W. F. , Mutual friction in heat current in liquid helium II. I. Experiments on steady heat currents. Proc. R. Soc. London A (1957) 240 114-127 3 Vinen. W. F. , Mutual friction in heat current in liquid helium II. II. Experiments on transient effects. Proc. R. Soc. London A (1957) 240 128-143 4 Vinen. W. F. , Mutual friction in heat current in liquid helium II. III. Theory of the mutual friction, Proc. R. Soc. London A (1957) 242 493-515 5 Murakami, M. and Iwashita, K. , Numerical computation of a thermal shock wave in He II. Comp. & Fluids (1991) 19 443-451 Fizdon, W., Schwerdtner, M. v., Stamm, G. and Poppe, W., Temperature overshoot due to quantum turbulence during the evolution of moderate heat pulse in He II. J. Fluid Mech. (1990) 212 663-684 Shimazaki, T. , Mura.kami, M. and Iida, T. , Second sound wave heat transfer, thermal boundary formation and boiling: highly transient heat transport phenomena in He II. Cryogenics (1996) 35 645-652 Kanari, T. and Murakami, M. , Numerical investigation of evolution of vortex line density in the case of transient heating, to be presented at ICEC 16 (1996) 9 Nemirovskii, S. K. and Tsoi, A. N . , Transient thermal and hydrodynamic processes in superfluid helium. Cryogenics 29 985-994
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Figure 2 Schematical drawing of the definition of Figure 1 Schematical drawing of the main the characteristic time of quantized experimental apparatus immersed in He II. vortex line development tve.
Figure 3 Superposed thermal shock wave profiles generated by the trapezoidal heat pulses of various heat flux value qp. The temperature T B is 1.90 K. The heating time tH is fixed at 500 Its. The distance from a heater to a sensor z is fixed at 10 mm.
Figure 5 Time evolution of quantized vortex line density obtained numerically by solving the vortex line density equation without neglecting the third tenn. The heat flux is supposed to be applied during the calculation. The initial VLD is assumed ,.) to be 10-cm/cm 3 for every case. (The calculation is done by Kanari et al. [8])
Figure 4 Measured characteristic time of quantized vortex line development tve as a function of applied heat flux qp for two temperatures. Linear regression results are also shown in the figure.
Figure 6 Comparison of the characteristic time tve and the time duration required for the quantized vortex line density L to reach certain values.
Cryogenic engineering
Heat transfer
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Non-dimensional Correlation for Boiling Heat Transfer from Sintered Porous Layer Surface
Rongshun Wang, Anzhong Gu, Zhen Li, Jianhai Huang Institute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, Shanghai 200030, P.R.China
Boiling heat transfer characteristics and data on sintered porous layer surface are summarized. Non-dimensional criteria which affect the pool boiling and channel boiling are analyzed. Then non-dimensional criteria correlation of pool boiling and channel boiling which offer designing foundations for industrial application of this technique are obtained
INTRODUCTION There are two methods to increase boiling heat transfer, 1)enlarge the heat transfer area, 2)enhance the boiling heat transfer coefficient. The latter doesn't involve consumption of material, which is an active and effective way. It's effective to improve nucleate boiling heat transfer coefficient by means of porous layer surfaces. Boiling heat transfer coefficient can be raised up to approximately 10 times as large as that of a smooth surface. The critical heat flux can be raised up to 80%[2] more than that of a smooth surface. It's because of the foregoing merits that its manufacturing technology and technique develop quickly. However, boiling heat transfer is a liquid-gas two phase process, whose heat transfer and mass transfer are extremely complex, then, there is two-phase flow of capillary in the porous layers , which make the boiling heat transfer mechanism even more complicate. Much research have been carried out by over a semi-century. Mechanisms were proposed according to the range of research respectively, which fail to agree with each other. Currently main research methods can be classified into two groups:(1) set up a mechanism model on the basis of porous surface structures combined with experimental investigation. (2) find out the dimensionless criteria that affect the heat transfer based on experiment and propose dimensionless heat transfer correlation, which largely aim at respective research range. This article is intended to summarize dimensionless criteria correlation applicable to boiling heat transfer on porous layer surfaces on the basis of a series of experiments. Boiling Heat Transfer Characteristics of Sintered Porous Surface Summarizing past views and our results of experiments, we consider that the enhancement of boiling heat transfer on sintered porous surface lies in the magnanimous interconnected large-sized hollow pores within the porous surface. These pores can effectively hold back the vapor to form nucleate centers, which enables them to sustain nucleate boiling under the condition of relatively small superheat. The foregoing ways are not possessed by smooth surface. Therefore, heat transfer on porous layer surface is enhanced 5 to 10 times that of smooth surface. Effect of Main Factors on Nucleate Boiling on Sintered Porous Layer Surface Heat transfer characteristics of porous layer surface are decided mainly by structural characteristics, physical properties of materials and boiling medium. The main reason that makes heat transfer on porous layer different from that on smooth surface is from structural characteristics. Structural characteristics of porous layer are expressed by particles diameter dp, thickness of the porous layer 6 and the porosity e. There exists an optimum particle diameter with which boiling heat transfer coefficient reaches maximum when thickness of the porous layer and the porosity are constant. This is found to agree with our analysis and experiments[4]. As the particle diameter decreases, the overall surface area in porous layer increases, 581
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which leads to the enhancement of boiling heat transfer, while the liquid-gas flow resistance in the capillary increases at the same time, which is not beneficial to boiling heat transfer. Likewise, there exists an optimum thickness of the porous layer with which boiling heat transfer coefficient reaches maximum when the porosity and particle diameter are set. It is the same reason as that of the foregoing. Literature[3]'s conclusion is aimed at optimum particle diameter of a definite boiling media matched with thickness of porous layer. This is in line with the above analysis. Control of the porosity is relatively difficult, and there's few experimental investigation on it. There's an optimum porosity based on the analysis of heat transfer area and flow resistance inside capillary pores. Currently, the porosity between 40% and 60% is considered proper. C O R R E L A T I O N OF POOL BOILING HEAT TRANSFER
There are many interconnected factors that affect the nucleate boiling of sintered porous layer surface. How to properly select criteria affect the accuracy and creditability of criteria correlation directly. Though there are many public reports on dimensionless criteria correlation, few can undergo further tests. Some indexes produce essential change when data are reduced or conflict with qualitative mechanism model. Therefore, we should find a dimensionless correlation in accordance with it's physical significance and correlative index don't produce great and essential changes when the numbers of correlative data reduce and increase. Dimensionless heat transfer correlation of cryogenic fluid nucleate boiling on smooth surface are summarized as below[7].
Nu - A,(r, e)(Gr pr)m[E,g(t')] n E 1=
Hs~pg A T
there, A,(r,O) is a coefficient related to surface condition and physical properties of fluid surface, g(P) is a coefficient related to system pressure while p=lbar, ~:(P)-1.
m,n obtained through correlating
experimental data. The equation expresses briefly the criterion GrPr(movement and physical properties of liquid) and the vapor criterion El(liquid disturbance and the latent heat transfer caused by vaporization on heating surface). With reference to above equation, the criteria which affect nucleate boiling on sintered porous layer surface are as follows: 1 Re and Pr which express movement and physical properties of fluid should be exhibited separately, because of the higher frequency of vapor bubbles escaping from porous layer surface, smaller departure diameter and greater agitation than those of a smooth surface. 2 A~(r,O) is used to express heating surface characteristic of a smooth surface. While the porosity, the thickness of the porous layer and particle diameter are major characteristic parameters of porous surface, but, best correlating result is obtained by adopting ~-,6/ dj, as criterion. 3 (1- C)2p /(a',;[,) is expressed as criterion for taking the peculiarity of porous layer structure into account and introducing the effect of thermal conduction of sintered particles. 4 p~ / p~ is used to express the effect of pressure. Thus dimensionless criteria correlation applicable to nucleate boiling on sintered porous layer surface is as below
Nu - "f( Re" pr" ~'6 (l - e')2 v P' I
<-7'
'/,
O-
] 0.5
1
there, N u -
ql
AT2.'
I-
6 o d l,
g(p, - p~)
it - c 2 ~ + ( 1 - c ) 2
' "'
P'
Re-
pght-g
,Pr
-
Ai
Correlating 75 data points of nucleate boiling heat transfer of LN 2 and R-113 on sintered porous surface[3][6]leads to following dimensionless criteria correlation:
Nu_89ReO.325prO.34
(c6-0.376ll
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583
(l-~')2p-0.114 P,
The measured Nusselt numbers are compared with the calculated ones in Figure l with a mean squared deviation of 15%, 84% of the data fall in the range of 25% and the overall data fall in the range of 35%. Correlative index and constant vary a little with the decrease of data points, therefore the above equat ion is stable, in the meantime, all of indexes are in agreement with our analyses[4]. CORRELATION
OF BOILING
HEAT TRANSFER
IN C H A N N E L
Boiling heat transfer in channels is confined by the space and affected by flow velocity and flow pattern, thus exhibits many characteristics which is different from that of the pool boiling. General dimensionless equation of smooth tube boiling with natural circulation is considered in [ 1] as follows:
Aft,-f(M, Pr, W,h,~,~) W is proposed as a new Re correlation to express dynamic properties of fluid boiling in tube,
w=q(1/d)~176g(pl --pg) Apparent height h expresses the difference between heat transfer rate inside the tube when boiling at low liquid level and that when the maximum liquid is filled. It also expresses the effect caused by the increase of two-phase vapor quality during the boiling process. Thus W and h express the difference between channel boiling and pool boiling. When sintered parameters are set, the criteria correlation of sintered porous layer channel is similar to that of smooth tube, but the index of I/d in W is certainly not 0.5 for porous tube. Thus adopt following form, N u - ")r Re' Pr' deq 'h' ~Tcr there,
4F d~q- -~,
F:flux area, L:heating perimeter.
The structural parameters of sintered porous layer are~= equation correlating 95 data points[2,5]is obtained. N I t - 004135Re~
50%,6=0.5mm,dp=
125/ml.A following
pr3SlT(de,7)':'471"~'l'~7~.,..J <"ls98h'-~41k.
the measured Nusselt numbers are compared with the calculated ones in Figure 2 with a mean squared deviation of 2.3 5%, 95% of the data fall in the range of 10% and the overall data fall in the range of 12%. Correlative index vary a little with the decrease of data points, therefore the equation is stable. We studied all the parameters which affect the boiling heat transfer process to obtain an correlation by dimensional analysis and correlated these data with a mean squared deviation of 2.32%. But it exerts a great influence on some index of dimensionless criteria when correlated data are decreased. Thus failed to express the actual process. Criteria used to express the physical process are overlapped. CONCLUSIONS 1 Boiling heat transfer characteristic and data on sintered porous layer surface are summarized. Nondimensional criteria which affect the pool boiling and channel boiling are analyzed. 2 The pool boiling heat transfer correlation is obtained with a mean squared deviation of 15%,
Nu-
89 Re ~
Pr ~
,f:b"
(1 - ~)2, --,:,,,4
P,
-
~/]'l 3 The channel boiling heat transfer correlation is obtained with a mean squared deviation of 2.3 5%, N i t - 0.041 35
Re" 937~pr 35,7(__)'"47'3(7L,18'~'~---~h~:,41 kd~,,,/ \ I~.,.J
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Fig. 1 Comparison between measured and calculated Fig.2 Comparison between measured and calculated Nusselt number for pool boiling Nusselt number for channel boiling REFERENCES
1 I.P.Vishnev, General relationship for heat transfer during boiling of cryogenic liquids Int. Chem. Eng. Vol. 14,No. 1 8-15 2 Jinping Liu,The experiment study of liquid nitrogen boiling heat transfer in vertical narrow channel of porous layer surface Master Paper of Xi'an Jiao Tong University(1986)(in Chinese) 3 Guili Zhai,The study of enhanced boiling heat transfer Master Paper ofTianjin University(1993 )(in Chinese) 4 Lanping Zhao,The performance study of boiling heat transfer of nitrogen in channel of porous layer surface Master Paper of Shanghai Jiao Tong University(1993)(in Chinese) 5 Rongshun Wang, Study and experiment of saturated liquid nitrogen boiling heat transfer in porous layer pipe Master Paper of Xran Jiao Tong University(1988)(in Chinese) 6 Yuenian Huang, The experiment study of boiling heat transfer on sintered porous layer surface Master Paper of Xi'an Jia0 Tong University(1985)(in Chinese) 7 Mingheng Shi,etal, The criteria correlation for heat transfer during bubble boiling of cryogenic liquids Cryogenics(1989)2(in Chinese) 8 Yaozhen Song, The experiment study of liquid nitrogen boiling heat transfer in vertical channel of porous layer surface outside tube Master Paper of Xran Jiao Tong University(1986)(in Chinese)
C r i t i c a l H e a t F l u x e s in P o o l Boiling of S u b c o o l e d L i q u i d N i t r o g e n a t E l e v a t e d P r e s s u r e s Koichi Hata, Masahiro Shiotsu, and Akira Sakurai* Institute of Atomic Energy, Kyoto University, Uji, Kyoto 611, Japan .Future Energy Research Assoc., Pasteur Building, 103-5 Tanaka Monzen-cho, Sakyo-ku, Kyoto 606, Japan The critical heat fluxes, qcr, on a 1.2 mm diameter platinum horizontal cylinder in liquid nitrogen due to quasi-steadily increasing heat inputs were measured for the subcoolings ranging from zero to about 40 K at pressures ranging from 0.3 to 2 MPa. The qcr values for saturated conditions almost agree with the corresponding values derived from the Kutatela.dze correlation for saturated conditions except those for the pressures higher than around 1.5 MPa. The critical heat fluxes for subcooled conditions, qcr,sub, at the pressures cannot be described by existing correlations based on hydrodynamic instability, except those near zero subcoolings at pressures lower than around 1.5 MPa. It was clarified that there exist two mechanisms for qcr,sub; two qcr, sub correlations which depend on subcoolings and pressures were suggested.
INTRODUCTION With remarkable advancement of high temperature superconductors, knowledge of boiling critical heat fluxes under steady and transient heat inputs will be required to understand stability of the superconductors resulting from local thermal disturbances in tllem. The steady and transient boiling critical heat fluxes for saturated conditions in a wide range of pressures have been clarified by the authors [1]. Knowledge of critical heat fluxes for subcooled conditions under steady and transient heat inputs is also important for the stability design of the superconductors. This paper deals with the critical heat fluxes on a horizontal cylinder in subcooled liquid nitrogen under steady-state heat inputs at pressures higher than about 302 kPa. Those for transient heat inputs will be published elsewhere in the near future. EXPERIMENTAL
APPARATUS
The apparatus used is shown in Figure 1. Tile boiling vessel(l) is a vacuum insulated cylindrical pressure vessel of 20 cm inner diameter and 70 cm height capable of working up to 3.5 MPa. The vessel has two sight
Fig. 2 Heat transfer processes for pressures of 101.3, 302.2 and 582.4 kPa
Fig. ] Schematic of experimental apparatus 585
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ports (5) and is equipped with a sheathed 1 mm diameter C C ( C o p p e r - C o n s t a n t a n ) thermocouple that is used to measure the bulk liquid temperature. The vessel is connected via a valve to a liquid feed tank (7). System pressure in the boiling vessel is automatically controlled to within +1 kPa of a desired value by a pressure control system consisting of a pressure transducer, a pressure controller (10), and a control valve (9). A platinum cylinder test heater of 1.2 mm in diameter was used. Two fine 30 #m diameter platinum wires were spot welded as electrical potential taps at about 10 mm from each end of the heater. The length between the potential taps was 62 nun. The heat transfer on the cylinder surface between the taps was measured. The test heater was anneMed and its electrical resistance versus t e m p e r a t u r e relation was obtained for the t e m p e r a t u r e s between 77 and 453 I( in liquid nitrogen, water and glycerin baths. The test heater was heated electrically by an output current from a power source (12 V, 600 A max.). The average t e m p e r a t u r e of the test heater between the potential taps was measured by resistance thermometry using the heater itself. A double bridge circuit with the heater as a branch was first balanced at the bulk liquid t e m p e r a t u r e . The signM voltage of the bridge circuit, together with the voltage drops across the potential taps and across a standard resistance by the output current were amplified and passed through analog-to-digital converters of the digital computer, respectively. The average t e m p e r a t u r e is calculated from the signM voltages by using tlte previously calibrated resistance-temperature relation. Measurement errors for the heater surface temperatures and the heat fluxes were estimated to be within about 4-1 I( and -t-2 percent, respectively.
EXPERIMENTAL
RESULTS AND
DISCUSSION
The critical heat fluxes, qcT, on the 1.2 mm horizontal cylinder in liquid nitrogen were measured for the range of pressures front 302.2 to 2047 kPa, and range of subcoolings from zero to 37 K under quasi-steadily increasing heat inputs given by exponential time function with a period of 20 s. The heat transfer processes at the atmospheric pressure, 302.2 and 582.4 kPa pressures under saturated conditions, and the process for the subcooling of 6.3 K at 582.4 kPa are shown in Fig. 2. Direct transitions from natural convection regime to film boiling occur at the pressures lower than around atmospheric pressure as mentioned in previous papers by the authors [1,2,3]. For saturated and subcooled conditions at pressures above atmospheric pressure, after initial boiling, the transitions from the natural convection regime to nucleate boiling occur, and then the heat fluxes gradually increase in the close proximity to the values corresponding to the fully developed nucleate boiling(FDNB) curve on q vs. ATtar graph up to the critical heat fluxes at which the the transitions to fihn boiling rapidly occur. This paper deals with the critical heat fluxes in F D N B for subcooled conditions at various pressures. The critical heat flux in FDNB for saturated conditions, qc,.,sot, at pressures ranging from 302.2 to 2047 kPa are shown in Fig. 3. The values for the pressures up to about 1016 kPa agreed quite well with the corresponding values derived from Kutateladze correlation for qc,',.~at, Eq.(1), with the coefficient K1 of 0.17. On the contrary, the values for 1.t67 and 2047 kPa were lower than the predicted values. This fact suggests that there can exist another mecllanism for the critical condition, which is different from that based on the h y d r o d y n a m i c instability. The critical values for various subcoolings, qcr.s,,b, at the pressures 582.4, 1016 and 2047 kPa are shown
Fig. 3 Critical heat fluxes under saturated conditions at elevated pressures
Fig. 4 Critical heat fluxes for liquid subcoolings at 582.4 kPa
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Fig. 5 Critical heat fluxes for liquid subcoolings at 1016 kPa
587
Fig. 6 Critical heat fluxes for liquid subcoolings at 2047 kPa
as a function of ATsub in Figs. 4, 5 and 6, respectively. As shown in these figures, the qcr,sub values at the pressures are different in their trends of dependence on subcoolings from those predicted by well known correlations of Kutatela.dze [4], and Ivey and Morris [5]: the q~,~,,.b values at 582.4 kPa are higher than the predicted values except those for a very narrow range of subcoolings from zero to about 5 K, and those at 2047 kPa are significantly lower than the predicted values even for that for zero subcooling. values obtained for wide ranges of subcoolings Sakurai et a.1. [6,7,8] have a.lready made it clear tha.t qr at pressures and heaters of various geometries in liquids such a.s water, liquid He I and ethanol could not be explained by existing correlations at all, and they suggested that the qcr,sub values were expressed by the following two correlations depending on subcoolings and pressures:
(1)
qcr,sat -- KI LPv[Og(pl - Pv)lP~] 1]4
qcr,sub = qcr,sat[l + K2(Pl/Pv)~
1"5]
(2)
+ B(%t~T,.,,.b/L) 2 + C(c.pzAT,~b/L)3}]
(3)
and
qr
-- qc*,.,,~t[1 + (Pl/P,,)~
where qcr,sat are values of qcT,,at which correspond to those due to the heterogeneous spontaneous nucleation. The power value of 1.5 in their correlation, Eq.(2), which is a modified Kutateladze correlation to take into account the non-linear effect of subcoolings was determined based on the experimental data for water [6,7] obtained so that the correlation can fit to within +5 % the experimental data for the subcoolings lower than 40 K at pressures up to 2063 kPa.. On the other hand, the q,:T values for the subcoolings higher than about 40 K at the pressures higher than about 200 kPa could not be expressed by the correlation, Eq. (2); this fact means that there can exist another mecltanism of heat transfer crisis at qcr,s,,b for higher subcoolings different from that based on the hydrodynamic instability. Sakurai et al. [7] suggested that the heat transfer crisis at the qc~,,..b would occur in FDNB due to explosive heterogeneous spontaneous nucleation (HSN) in originally flooded cavities on tile solid surface at IISN tempera.ture, ttowever, derivation of the correlation for qcr,sub based on HSN model is difficult at present time by the lack of fundamental database concerned; therefore, they presented the empirical equation, Eq. (13). The comparison between the experimental d a t a obtained for liquid nitrogen and the corresponding values by the correlations obtained for water is performed as mentioned below. The curves of qr versus AT.,,,b for each pressure obtained from Eqs. (1), (2) and (3) with the suitable coefficients I(1, 1(2, A, B, C given in Table 1 are also shown in Figs. 4 to 6. The %~,~,,.5 for the subcoolings lower than about 10 K at pressures lower than 1016 kPa agree quite reasonably with those derived from Eqs. (1) and (2). The qc~,.~,,.t, values for the subcoolings higher than about 10 K at pressures lower than 1016 kPa, and the qcr,sub va.lues for all subcoolings at 20,t7 kPa agree well with the corresponding values derived from Eq. (3). It is recognized that the %,. for the subcoolings at elevated pressures in liquid nitrogen ahnost occur due to the HSN in FDNB similar to corresponding cases for water except those for narrow range of subcoolings from 0 to 10 K at pressures lower than 1016 kPa.
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Table 1 Coefficients in Eqs. (1), (2) and (3) for LN2 Heater
Horizontal Cylinder (1.2 rnm dia.)
K1
K2
A
B
C
0.17
0.7
0.378
-0.104
0.0111
Table 2 Coefficients in Eqs. (1), (2) and (3) for LHe I Heater Horizontal Cylinder (0.2 mm dia.)
K1
K2
A
B
0.19
0.87
0.378
-0.104
C 0.0111
Fig. 7 Critical heat fluxes on a. 0.2 mm-diam wire in liquid He I versus liquid subcooling with pressure as a parameter The qcr,sub for subcoolings ranging from zero to about 2.5 K at the pressures of 101.3, 142 and 196 kPa on a 0.2 mm diameter horizontal cylinder in liquid He I have been measured by Sakurai et al. [8] and are compared in Fig. 7 with the values derived from Eqs. (1), (2) and (3) with the corresponding coefficients given in Table 2. The coefficients A, B, and C are the same as those for liquid nitrogen. The qr values almost agree with the values derived from Eq. (3). As shown in the figure, the trend of qc~,~,b for subcoolings and pressures is similar to that for liquid nitrogen obtained here.
CONCLUSIONS The critical heat fluxes on a 1.2 mm diameter platinum horizontal cylinder in liquid nitrogen due to quasisteadily increasing heat inputs were measured for the subcoolings from 0 to about 40 K at pressures ranging from 302 to 2047 kPa. The q~,~,b values obtained for wide ranges of subcoolings and pressures were expressed by Eqs. (1) and (2) and Eq. (3) depending on subcoolings and pressures. The qc~,.~,b for the pressures higher than around 1.5 MPa were expressed by Eq. (3) for all subcoolings including zero. The qcr,sub for wide ranges of subcoolings and pressures for liquid nitrogen occur due to the hydrodynamic instability or the HSN similar to those for other liquids. REFERENCES
1. Sakurai, A., Shiotsu, M., and ][ata K., New transition phenomena to film boiling due to increasing heat inputs on a solid surface in pressurized liquids, Instability ill two-phase flow systems, (1993) ASME HTD-Vol.260, 27-39 2. Shiotsu, M., Hata, K., and Sakurai, A., IIeterogeneous spontaneous nucleation temperature on solid surface in liquid nitrogen, Advances in cryogenic eng. (1990) 35,437-445 3. Sakurai, A., Shiotsu, M., and I[ata, K., Boiling heat transfer from a horizontal cylinder in liquid nitrogen, Heat transfer and superconducting magnetic energy storage (1992) ASME HTD-Vol.211, 7-18 4. Kutateladze, S.S, Heat transfer in condensation and boiling (1959) AEC-tr-3770, USAEC 5. Ivey, H.J. and Morris, D..J., On the relevance of the vapour-liquid exchange mechanism for subcooled boiling heat transfer at high pressures (1962) AEEW-R 137. UKAEA 6. Sakurai, A., Shiotsu, M., and tiara, K., Transition phenomena from non-boiling regime to fihn boiling on a cylinder surface in highly subcooled and pressurized water due to increasing heat inputs (1995) ASME Paper 95-WA/HT-17. 7. Sakurai, A., Shiotsu, M. and Fukuda, K., Pool boiling critical heat flux on a horizontal cylinder in subcooled water for wide ranges of subcooling and pressure and its mechanism, to appear in Subcooled boiling phenomena session in 1996 national heat transfer conf. (1996) 8. Sakurai, A., Shiotsu, M. and tIa.ta, K., 1996, Incipient boiling superheats and critical heat fluxes due to increasing heat inputs in subcooled tie I a.t various pressures, to appear in Advances in Cryogenic Engineering (1996) 41
The Measurement of Vapor Bubble Vibration During Noisy Film Boiling in Superfluid Helium
Mahito Yamaguchi, Masahide Murakami Institute of Engineering Mechanics University of Tsukuba Tennodai l-l- 1 Tsukuba 305 Japan
Noisy film boiling phenomena in He II had been visually investigated with the aid of Schlieren and shadowgraph methods. The frequency of vapor bubble formation and crush during noisy boiling was measured from high-speed video image. In the present experiment special attention is paid to the investigation of the fundamental mechanism of noisy film boiling phenomenon. For this purpose, we also measured the pressure oscillation in He II caused by vapor bubble formation and crush with a pressure sensor mounted above a heater. The wave form and the FFT result of the pressure oscillation are minutely examined.
INTRODUCTION It is known that there appear two film boiling modes in He II; noisy and silent boiling modes. The occurrence either noisy or silent film boiling depends on the magnitude of the hydrostatic pressure and on the He II temperature as well as on the heat flux. For larger hydrostatic pressures and at lower temperatures, noisy boiling accompanied by a laud acoustic noise and a mechanical vibration exclusively ap pears [ 1]. Loud noise results from unstable violent motion of a large vapor bubble [2]. Noisy and silent film boiling phenomena on a planar heater in He II had been visually observed with the aid of Schlieren and shadowgraph methods [2]. It is a well known experimental fact that the heat transfer coefficient for noisy boiling is smaller than that for silent one [1,3,4]. However detailed physical mechanism of noisy boiling has not been fully made clear. In view of the status of researches, the study of the fiandamental mechanism of noisy boiling is of importance for effective applications of He II to cooling of superconducting ma~ets. In the present study, our attention is focused on the cause of audible laud noise generation in order to investigate the fundamental mechanism of noisy boiling.
EXPERIMENTAL APPARATUS In the present experiment, the specially designed cryostat with optical windows is utilized for the visualization study as illustrated in Figm'e 1. Boiling phenomena in the test section are optically observed through these windows. Two kinds of planar heaters are used in the visualization experiment. One is a transparent heater consisting of indium oxide thin film (1000 A-thick) vacuum deposited on an optical 589
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glass substrate with dimensions of 25mm • 25mm mounted vertically in the test section, through which boiling phenomena can be observed. The other is an opaque heater, consisting of evaporated Ni-Cr thin film on a Pyrex substrate, mounted horizontally. Several heaters of this kind of various sizes are also used. The side view of boiling phenomena can be observed with these horizontal heaters. Experiments are conducted under a constant temperature condition. A heater is excited by step-wise electrical current for one second. The resulting phenomenon is visualized with the aid of shadowgraph and Schlieren methods. In addition, the pressure fluctuation is measured with a pressure sensor (PCB106M 122) mounted above the heater surface in He II in order to investigate detail mechanism of audible noise generation as illustrated in Figure 2. The variation of the pressure wave forms with the distance of the pressure sensor from the heater, d, and the FFT result are minutely examined.
RESULTS AND DISCUSSION A large vapor bubble is found to repeat expansion and crash nearly periodically accompanied by a laud audible noise in noisy boiling[2]. It is another distinction between noisy and silent boiling states that the density fluctuation in bulk He II is observed only in the noisy case. The density fluctuation is considered to be resulted from pressure fluctuation caused by violent motion of a vapor bubble. The pressure s i s a l s from the pressure sensor mounted 30mm above the heater surface (25ram • 25mm) are shown in Figure 3 for both noisy and silent cases. The ordinate is the out put voltage from the s i s a l conditioner for the pressure sensor as the sensor is not calibrated under the liquid helium condition. It is seen that the frequency of nearly periodic spiky pressure peaks correspond to that of vapor bubble formation and crush. Large amplitude pressure fluctuation is not observed at all in the silent case as seen in l,'iL,ure 3. The frequency can be obtained from the FFT result of the pressure oscillation data. The frequency of vapor bubble formation and crash during noisy boiling was also measured from a high-speed video image. The ageement of the visualization result with the pressure measurement data is satisfactory as seen in t"igm'e 4. It is found that the larger heat flux or the larger heater area causes the lower frequency. The variation of the frequency with the depth of the immersion of a heater, h, or the hydrostatic pressure, is plotted in l"i[,,,ure 5 by taking the heat flux as a parameter. It is found that the smaller hydrostatic pressure causes the lower frequency. It drastically decreases in the transition r e , on between the noisy and silent boiling modes. It is also found that as the hydrostatic pressure becomes small, the amplitude and the frequency of the pressure wave becomes irregular as a typical behavior in the transition region as shown in l"igure 6. It is of importance to note the distinction between the two pressure oscillation data measured at two different distances from the heater surface, d, as shown in l~Tgure 7. The pressure sensor, mounted 60ram above the heater surface, is always located outside of a vapor bubble, while that 5ram above the heater surface is involved in a vapor bubble in its g o w n up stage. It is seen that in the latter case, the signal become quite small between the pressure spikes though in either case the pressure wave is generated almost periodically. The reason is considered that after a pressure spike occurs the pressure sensor enters a vapor bubble in the latter case. It can be concluded that the pressure spikes appear at the moments of a vapor bubble crush.
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CONCLUSION 1. The frequency of pressure spikes measured with a pressure sensor corresponds to that of vapor bubble formation and crush. 2. As the hydrostatic pressure becomes small, the amplitude and the frequency of pressure wave become small and irregular as a typical behavior in the transition region. 3. The pressure spikes appear at the moments of vapor bubble crush.
REFERENCES Leonard, A. C., Helium II noisy film boiling and silent film boiling heat transfer coefficient values, Proc. ICEC 3 (1970) 109-114 Katsuki, Y.,et al. Visualization study of film boiling onset and transition to noisy film boiling in He II Cryogenics (1995) 35 631-635 Betts, K. R. and Leonard, A. C., Free convection film boiling from a fiat, horizontal surface in saturated He II . . . . . . . . . . . .
Eng. (1975) 21 282-292
Steed, R.C. and Irey, R.K., Correlation of the depth effect on film-boiling heat transfer in liquid Helium II Adv. Cryog. Eng_ (1970) 15 299-307
Figure 1 Schematicillustration of experimental cr3,ostat with optical windows
Figure 2 Pressuresensor is mounted above a planar heater. The pressure oscillation caused by a vapor bubble formation and crush can be measured.
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F i g u r e 3 The distinction of pressure wave forms between
F i g u r e 4 Tim comparison of the frequency ofvapor bubble
noisy and silent film boiling. (,'1) noisy boiling, Tu=I.8K,
formation and crash measured from high speed video inaage
q=6W/cm 2. h=20cnr (b) silcnt boiling, Tr~=I.SK. q=6W/cm ~,
and pressure sensor.
h=5cm.
F i g u r e 5 The variation of the fiequency of vapor bubble
F i g u r e 6 Tile comparison of tim amplitude of pressure
formation and crash with the depth of tile hcater, h. ~ q = 6 W/c m2. @q = 10 W/cm~. ~kq= 15 W/cm 2,,~q= 20 W/cm 2.
waves for two depths ofa hcatcr, h. (a) Tn=! .8K, q=6W/cm 2,
Trs= 1.8K. hcaicr size S=25mm X 25ram.
Figure
h=27.5cm; (b)Tn=I.8K, q=6W/cm 2, h=7cm. hcatcr size S =25ram • 25ram.
7 The dislinclion between Ihc Iwo prcssure wave forms measured at two different dislances frOlll tim hcaler, d.
(a) Tu=I.8K, q=6W/cm 2. d=60mm: (b)T~=I.8K, q=6W/cm ~, d=5mm : heater size S =25tuna X25nma.
Influence of Surface Roughness on Transient Nucleate Boiling of Cryogens Shuichiro Fuchino, Noriharu Tamada, Itaru Ishii, and Makoto Okano Electrotechnical Laboratory, 1-1-4, Umezono, Tsukuba-shi, Ibaraki 305, Japan The influence of the surface roughness on nucleate boiling has been studied well, because it is known to be an important parameter. However in liquid He I, no marked differences have been found. In this paper, this anomalous behavior in liquid He I is investigated by observing the transient bubble formations for different surface roughness using a high-speed camera and an optical sensor. Marked differences of the bubble formation time were observed in liquid nitrogen for different roughnesses but not for liquid He I. These results are discussed on the basis of critical radius.
INTRODUCTION The surface roughness in nucleate boiling is known to be an important parameter. It would trap foreign particles or gas which constitute nucleation sites and liquid can be superheated inside the small cavities until the molecules attain the energy of vapor phase. This behavior is significant for room-temperature liquids because changes in the rms (root mean square) roughness values are known to affect the surface excess temperatures AT above the saturation temperature of the bulk liquid. In contrast, in liquid He I, no marked AT differences between various rough surfaces have been found [1-3 ], though polishing caused an increase in A T. Some reason is that liquid helium has some peculiarities connected with good wetting of the solid surface and complete absence of gaseous impurities, hence the surface temperature is easier to reach the homogeneous nucleation temperature before an appreciable number of heterogeneous nucleation site have had the time become active[4]. Optimum surface conditions, however, are desirable for efficient cooling in some components of liquid-helium cryosystems. Therefore, it is very important to investigate the influence of the surface roughness in liquid helium nucleate boiling. In this paper, this anomalous behavior is investigated by observing the transient bubble formations for the different surface roughness using a high-speed camera and an optical sensor. EXPERIMENTS Figure 1 shows a heating surface. The surface of the quartz substrate was polished as a mirror at first, which is called a smooth surface. Then the other half of the smooth surface was polished with different grit abrasive, which is called a rough surface. The surface conditions were described by means of the root mean square (r.m.s.) roughness. The surface finish range between a r.m.s, value of 0.1 and 0.4 tz m. A transparent thin film heater, which is made of ITO (Indium Tin Oxide), has been deposited on the quartz substrate. Figure 2 shows a schematic view of the experimental set-up. The heating surface is hanging in the cryostat vertically. As light from the lamp goes through the heating surface perpendicularly, the formation of bubbles can be observed directly by a high-speed camera. The high-speed camera has maximum recording speed of 10,000 frames per second to observe the fast boiling phenomenon. Another effective measurement system [5] is also used in this work. Figure 3 shows the configuration of the
Figure 1 Heating surface. 593
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measurement system for transient boiling phenomenon. The principle of the sensing system is as follows. Light from a emitted GaAs diode passes through a light channel, of which diameter is lmm, 9 is 10 mm, and reflects on the heating surface and goes to a Si photo diode. When bubbles appear on the heating surface, output signal of the Si photo diode fluctuates with the appearance of bubbles because of different refractivity. We used an ultra red type GaAs photo emitting diode and a Silicon type photo diode to make them work in liquid helium temperature. RESULTS AND DISCUSSION At first we demonstrated the influence of the surface roughness in liquid nitrogen. Figure 4 shows the difference of the bubble generation for different surface roughness in liquid nitrogen. It shows clearly that bubbles are easier to generate for the rough surface ( the r.m.s, value is 0.1 tz m). Figure 5 shows the bubble growth time ( the interval from the start of heating to first bubble appearance). It also shows that bubbles are easier to grow on the rough surface. It means that the surface roughness influences the transient nucleate boiling in liquid nitrogen. Figure 6 and 7 show other interesting results. Figure 6 shows the values of tD 2 for various heating powers, where f is the frequency of bubble formation and D is the departure diameter of the bubble. However Jacob[6] have proposed fD=constant and McFadden and Grassmann [7] have proposed ff)~ as against tD2=constant in our experiments. Foster and Zuber [8] found an approximate relation for the formation of vapor bubbles on solid walls within boiling liquids D .._Ctg0.5
Figure 2 Schematic view of the experimental set-up.
Figure 3 Configulation of the measurement system.
where C is a constant depending on the properties o f the liquid and t~ is the growth period (time during which a bubble grows on the heating surface up to separation from it). Considering f-1/tg, the above equation could be written as D~f=C: This is just the relation found in our experiments. It is worthwhile to note, that the values for boiling nitrogen, as found by Bewilogua et al. [9], was described by the same equation (with C2-~7.6 mm2/s). Figure 7 shows the values of tD3N, where N is the number of the site where bubbles generate. This value is almost equal to the total amount of gas formation per second. If most of the heating power is used to produce bubbles, this value must be the same. The power in excess of producing bubbles may be used for heating up the superheated layer or bulk liquid temperature by natural convection. The promotion of bubble generation is necessary for avoiding local
Figure 4 Differences of the bubble generation. (heat flux:
0.5W/cm 2, at 1s after heating)
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superheating. When we tried to demonstrate effect of roughness in liquid helium, however, we could not recognize any difference by photographs. The reason is that the time difference is too short and the diameter of helium bubble is too small to recognize in spite of using a high-speed camera. Also, bubbles in the back-ground impeded the recognition. So the optical sensor measuring system was used. Figure 8 shows a sample output from the optical sensor for smooth and rough surfaces at 0.5 W/cm ~ heater input. This result shows that bubbles grow faster on the rough surface. This time difference is not so large compare to that in liquid nitrogen. This anomalous behavior in liquid helium is discussed on the basis of critical radius. The critical radius r e is given as follows [ 10]: rc~.
20"% Ov~.v ~T
Figure 5 Bubble growth time dipendence on heat flux.
where a is the surface tension, T~ is saturated temperature, o~ is the density of the gas, ,t., is the latent heat and A T =Tw-T~ is the temperature difference between the temperature of the wall and the saturated liquid. Table 1 [ 11,12] shows the critical radius of several cryogens. From the data, we can see that the critical radius of liquid helium is very small compared to those of other cryogens. The surface roughness in this experiment is almost equal to the critical radius of liquid nitrogen. This roughness may be too large to become an nucleation site for liquid helium. CONCLUSIONS The influence of the surface roughness has been investigated for liquid nitrogen and helium by observing the transient bubble formations using a high-speed camera and an optical sensor. Marked differences in bubble formation time were observed in liquid nitrogen. However, the difference was not clear in liquid helium. The optical sensor could detect the bubble formation time difference in liquid helium. This time difference is seen to be very small. These results are also discussed on the basis of critical radius. The surface roughness in this experiment is too large to become an nucleation site for liquid helium. The influence of the smaller roughness will be investigated in a future work. Table 1 Liquid
Figure 6 Dependence of ~ 2 on heating power.
Critical radius of liquids
He
H2
Ne
N2
Ar
02
ATmax(K) 0.7
2.0
2.8
16
20
22
0.12
0.1
rc(tZm ) 0.0035 0.068
0.11 0.12
,,
Figure 7 Dependence of fD3N o n heating power.
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Figure 8 Outputs from optical sonsor for 0.5W/cm 2heat flux.
REFERENCES
6 7 8 9 10 11 12
Boissin, J. C. et al. In: Adv. Cryog. Eng. Plenum Press, New York (1968) 1__33607-616. Subbotin, V. I. et al., Heat transfer and hydrodynamics in cooling channels of superconducting devices Cryogenics (1985) 25261-265 Chandratille, G. R. et al., Pool boiling heat transfer to saturated liquid helium from coated surface Cryogenics (1989) 2___99588-592 Brodie, L. C. et al., Transient heat transfer into liquid helium I J. AoD1.. _ Phys. (1977) 4__882882-2885 Tamada, N. et al., Optical measurement of transient phenomenon of liquid helium boiling under very high centrifugal operation Cryogenics (1993) 331023-1027. Jacob, M.In" Heat Transfer John Wiley, New York (1949)1 McFadden, P. W. and Grassmann, P. In: Int. J. Heat Mass Transf. (1962) 5169 Foster, H. K. and Zuber, N. J. Appl. Phys. (1954) 25474 Bewilogua, L. et al., Studies on bubble formation in low boiling liquids .Cryogenics (1970) 1___0_069-70 Tanazawa, I. and Enya. S. In: Engineering Heat Transfer Ohm Sha, Tokyo (1982) 155 Astruc, J. M. et al., Comparison of heat transfer to hydrogen, deuterium, and neon boiling with free convection at atmospheric pressure Cryogenics (1969) 9248-250 Lyon, D. N. In: .Chem. Eng. Progress Symp. Series 64 (1968) 8__7_782-92
Superheating of Liquid Mixtures of a He and
4
He
Kazu Nishigaki, Minoru Takeda, and Yoshio Maruno Kobe University of Mercantile Marine, Higashinada, Kobe 658, Japan
In order to investigate the effect of 4 He atoms on vapor bubble nucleation in liquid a He, we have measured the superheating of mixtures of 4 He and a He by means of the continuous heating of a liquid specimen: The sample mixtures (ca.l.4 • 1 0 - 4 cc) with different 4 He concentrations (5.4 ~" 75%) were made to superheat in a small glass capillary (0.55 mm in i.d. and 1.05 mm in o.d.) by electric heater along various isobars (0.10 "~ 1.04 atm.). The experimental data were analized with particular attention to the superheating limit. It is found that the maximum attainable temperatures (T s ) clearly raised with increasing the concentration of 4 He in liquid 3 He, indicating that the nucleation barrier was increased by adding 4 He atoms to liquid a He. The values of T ~ were also analized on the basis of the thermal fluctuation theory of nucleation.
INTRODUCTION Accurate information on the nucleation rate, and energy and size of bubble nuclei at the onset of liquid-gas phase transition can be obtained by measuring the superheat limit of a liquid (T ~ ) and comparing the result with that predicted by the nucleation theory. Liquid helium can provide high-purity samples with distinct quantum characteristics when used in this type of study. Therefore, liquid helium is of great interest in research on nucleation. We have conducted research on the mechanism of bubble nuclei generation focussing on the T s of pure liquid a He and liquid 4 He [ 1,2] . As a result, we have established that the T ~ value predicted by the homogeneous nucleation theory agreed with the experimental T ~ value for pure liquid a He or liquid 4 He, a quantum liquid; and that quantum deviations from the principle of corresponding states are clearly observed in these liquids. We are currently conducting research on bubble nucleation using a mixture of liquid a He and 4 He. A major characteristic of the liquid helium mixtures which is attracting attention is the effect of the mixed molecules on boundary energy. Recently, Suzuki et al. [3] reported that the surface tension increases when a small amount of liquid 4 He is mixed with liquid 3 He. Since the liquid-gas boundary energy greatly affects bubble nucleation, the energy is also considered to affect T ~. However, no reports have been published to confirm this point. The aim of the current study is to clarify the effects of a mixture of liquid 4 He molecules with liquid 3 He on the superheat limit, namely, the homogeneous bubble nucleation. The conventional nucleation theory of Hirth and Pound [4] was used for predicting the superheat limit. 597
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EXPERIMENTAL SET-UP Rated purities of 3 He and 4 He used in the experiment were 99.98% and 99. 999%, respectively. Four types of gas samples with 3 He tool concentrations of 25.00%, 50.01%, 74.99% and 94.57% were prepared by mixing these gases. A liquid sample was prepared by cooling the gas samples to less than 2K in order to induce liquefaction in a small glass capillary, sample cell (inner diameter: 0.55ram; outer diameter: 1.05ram; length: 30ram; volume: 7.13 • 10-3 cc). A continuous isobaric heating method was used to heat the sample and the heating time was approximately 1 to 3 sec. The values of T ~ were determined from a thermogram which was obtained by recording the temperature change of the liquid sample during heating. A carbon sensor was used to measure T ~. The temperature measurement range in the experiment was from 1.5 to 4 . 1 K and the accuracy of the temperature measurement was +__0.01K. The measurement range of the pressure applied to the liquid sample was from 0.05 to 1.00 atm and the accuracy of the pressure measurement was _ 9.87 • 10-s atm. Our experimental setup system is given in Figure 1.
EXPERIMENTAL RESULTS AND ANALYSIS Figure 2 shows one of the results from this experiment, in which T ~ and saturation temperature are expressed as a function of pressure for the sample with a He mol concentration of 74.99%. In this figure, the experimental values of saturation temperature are represented by open circles, and those of T ~ are represented by open triangles. The dashed line indicates results obtained by the best fit method using experimental values of the saturation temperature, and dotted line represents values of T ~ predicted by the nucleation theory. In the calculation, T ~ was assumed to be the temperature, which corresponds to a nucleation rate of 1. The experimental data of Kerr [5-1 were used to determine the density of liquid 3 He4 He mixture, which was necessary for the calculation, and the experimental data of Esl'son et al. !-6-] were used to determine the surface tension. As shown in Figure 2, T s rapidly increases as the pressure increases in the low-pressure region (approximately less than 0.25 atm), and the increase almost becomes linear in the high-pressure region. Also, at a pressure of approximately 0.1 atm, the experimental and theoretical values of T are almost identical; however, as the pressure increases further, the experimental values become larger than the theoretical values. The same tendency of the superheat limit being different from saturation temperature was also observed in the other three samples with different a He mol concentrations. However, such a difference was not observed in pure liquid a He or 4 He. Slight differences in the saturation temperature were observed between the current experimental data and the experimental data of Roberts et al. [7] at temperatures less than about 2.3K; however, both sets of data showed a similar temperature dependence. The 3 He mol concentration dependence of the experimental and theoretical values of T ~ will be discussed next. For example, at a pressure of 0.15 atm and 0.50 atm, the experimental values of T ~ were expressed as a function of a He mol concentration and are shown in Figure 3. In this figure, the experimental values of T ~ are represented by open triangles and open circles, and the theoretical values are represented by solid line. Both theoretical and experimental values of T change as a He mol concentrations change. However, experimental values become larger than the theoretical values as the pressure applied to the liquid samples increases as shown in Figure 2; As described above, the experimental and theoretical values of T ~ of liquid helium mixtures differ in a wide range of pressure and a He mol concentration, whereas those of pure liquid helium do not
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differ.
The reasons for this are investigated from both experimental and theoretical perspectives. In conclusions, we have first presented experimental data on the maximum superheating temperatures of mixtures of a He and 4 He. Our data demonstrated the inadequacy of the conventional theory of nucleation in describing the bubble nucleation in a mixture of liquid helium.
REFERENCES 1
2 3 4 5 6 7
Nishigaki, K. and Saji, Y. International Conference on Low Temperature Physics - 18 Kyoto, JAPAN (1987), Nucleation in Superheated a He Jpn. J. Appl. Phys. (1987) 26 Supplement 26-3 (BH25) Nishigaki, K. and Saji, Y. Superheat Limit of 4 He and its Ouantum Deviation from Classical Behavior J. Phys. Soc. ]pn. (1983) 52, 7, 2293-2296 Suzuki, M. Yamanabc, M. Ohtani, N and Sato, A. Surface Tension of Dilute Solutions of 4 He in a He Physica B(1994), 194-196,873-874 Hirth,J.P. and Pound, G.M. Progress in Materials Science (1963) 11, 149 and references cited therein Kerr, E.C. Low Temperature Physics and Chemistry (1958) 158 Esel'son, B.N. Ivantsov,V.G. and Shvets,A.D. Soviet Phys. JETP (1963) 17, 330 Roberts, T.R. and Sydoriak, S.G. Vapor Pressures of a He --4 He Mixtures Phys. Rev.
(1960) 118, 901
Figure 1. Experimental setup system
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Figure 2.
Superheat
limits Ts
and s a t u r a t i o n temperatures for
mixtures as a f u n c t i o n of pressure. /k
and
9 denote the
the 74.99moi% 3He
The broken curve shows the t h e o r e t i c a l T s ,
experimental data of the superheat
limit
and saturated
temperature, r e s p e c t i v e l y .
Figure
3.
Superheat
limits
(/k)and P=O.5Oatm(C)).
as a f u n c t i o n
of
3He tool
concentration
at
P=O.15atm
Nucleate Pool Boiling Heat Transfer to Slush Hydrogen
Katsuhide Ohira, Hitoshi Furumoto Nagasaki R&D Center, Mitsubishi Heavy Industries, Ltd.
1-1 Akunoura-machi, Nagasaki 850-91 Japan
Laboratory data has been obtained concerning the nucleate pool boiling heat transfer of slush hydrogen. The heat transfer surface was a flat disk made of copper, and three different orientations of the heat transfer surface were used. In addition to that for slush hydrogen, heat transfer data for normal boiling point liquid hydrogen, triple point liquid hydrogen, slush nitrogen, normal boiling point liquid nitrogen, and triple point liquid nitrogen was obtained. Comparison between these sets of heat transfer data and that for slush hydrogen, including the results of observation of the heat transfer surface, revealed the nucleate pool boiling heat transfer characteristics of slush hydrogen, which has hardly ever been investigated.
INTRODUCTION Slush hydrogen (52.8Torr, 13.8K) is a mixture of solid and liquid hydrogen showing an increased capacity to absorb heat and higher density than liquid hydrogen. Technical development is being undertaken for its application as a fuel for space planes and clean energy systems [1 ]. This paper reports data obtained concerning nucleate pool boiling heat transfer characteristics. Coeling [2] and others [3] have extensively reported on the heat transfer characteristics of liquid hydrogen, but there is hardly any heat transfer experimental data available for slush hydrogen. In fact, the only paper that was identified in our search was by Sindt [4], who used a 0.025m diameter flat disk made of stainless steel for the heat transfer surface.
EXPERIMENTAL APPARATUS AND PROCEDURE The experimental apparatus consisted of a heat transfer unit, a vacuum insulated glass dewar for slush hydrogen in which the heat transfer unit was placed and which allowed visual observation of heat transfer conditions, internal pressure adjustment apparatus for the glass dewar, measuring and data processing equipment, heater current and voltage measurement device, and a slush hydrogen densimeter. Glass Dewar The glass dewar used in the experiment, shown in Figure 1, consisted of three glass dewars nested together: the outer dewar was for liquid nitrogen, the second dewar was for liquid hydrogen, and the inner dewar was for liquid or slush hydrogen. The outer and second dewars were to prevent heat leakage to the inner dewar. When the inner dewar was filled with normal boiling point liquid hydrogen, heat leakage to the liquid hydrogen was measured at 1.2W under experimental conditions. Heat transfer experiments with liquid or slush nitrogen were conducted with the second dewar for liquid nitrogen and the inner dewar for liquid or slush nitrogen. The capacity of the inner dewar was approximately 12L, and it was equipped with an observation slit and a lighting slit. Heat Transfer Unit A cross section of the heat transfer unit is shown in Figure 2. The heat transfer surface was a flat disk of electrolytic tough pitch copper 0.025m in diameter. It was polished with emery paper such that surface roughness was measured at below ___l/~m. Manganese wire wound around the heater block was used to 601
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heat the heat transfer surface, and a germanium resistance thermometer was inserted at the center for measurement of the temperature of the heat transfer surface. Apiezon grease was used in order to achieve high thermal conductivity of the heater block and the thermometer, thereby reducing temperature measurement error. In addition, a vacuum was established in the interior of the stainless steel housing using a vacuum pump, and multi-layer insulation was used around the heater block to preclude radiant heat. Experimental Procedure In addition to slush hydrogen (52.8Torr, 13.8K), heat transfer tests were conducted on normal boiling point (NBP) liquid hydrogen (1 atm, 20.3K), triple point (TP) liquid hydrogen (52.8Torr, 13.8K), as well as on slush nitrogen (96.4Torr, 63.2K), NBP liquid nitrogen (1 atm, 77.4K) and TP liquid nitrogen (96.4Torr, 63.2K). The slush hydrogen and slush nitrogen were between 20 and 35% solid by weight at the start of experimentation, and the heat transfer surface was placed in each of the three orientations (horizontal upwards, vertical, and horizontal downwards) for each of the heat transfer tests. The liquid hydrogen used was more than 95% para-hydrogen. After liquid hydrogen was placed in the inner dewar and a vacuum pump was used to reduce the pressure to triple point, the Freeze-Thaw method was used to produce slush hydrogen [1 ]. After slush hydrogen production was stopped, the heater inside the heat transfer unit was used to heat the heat transfer surface; the heater current was used to control the heat flux. Data for temperature, pressure, heater current, voltage, etc. was recorded at the point when the temperature of the heat transfer surface stabilized. The state of the heat transfer surface and the slush hydrogen could also be observed through the observation slit.
RESULTS AND DISCUSSION Figure 3 shows the test results for liquid nitrogen and slush nitrogen for the horizontal upwards orientation of the heat transfer surface, and Figures 4 through 6 show the results for liquid hydrogen and slush hydrogen for each orientation. All of the test results are given in the form of power per unit of area (heat flux) vs. the temperature difference between the bulk liquid and the heater surface. Open symbols indicate increasing heat flux at the data points, while solid symbols indicate decreasing heat flux. Liquid Nitrogen and Slush Nitrogen Results Figure 3 shows when vapor bubbles were first observed during the heat flux increase, when burnout occurred, and the last vapor during heat flux decrease. The following general conclusions can be drawn from the results for the three orientations: (1) For each orientation of the heat transfer surface with TP liquid nitrogen, transition from the nonboiling natural convection region to boiling occurred at temperature differences of over 10K. (2) For each orientation of the heat transfer surface, except in the nonboiling natural convection region, the heat transfer characteristics of TP liquid nitrogen and slush nitrogen exhibited the same tendencies, and these tendencies were clearly different from those of NBP liquid nitrogen. (3) Heat flux at burnout was the highest for NBP liquid nitrogen at each orientation of the heat transfer surface. Also, for each of the liquids (NBP liquid nitrogen, TP liquid nitrogen, and slush nitrogen), heat flux at burnout was the lowest at the horizontal downwards orientation of the heat transfer surface. Liquid Hydrogen and Slush Hydrogen Results Figure 4 provides a comparison with the heat transfer test results for liquid hydrogen by Coeling et al. [2] Their study used both polished stainless steel and copper heat transfer surfaces (0.025m flat disk), and was conducted at a pressure of 879Torr. The current authors used copper for the heat transfer surface, and the results closely match those obtained by Coeling et al. with a copper heat transfer surface. Figure 4 also provides a comparison with the results obtained by Sindt [4] with a stainless steel heat transfer surface for TP liquid hydrogen and slush hydrogen. Differences owing to the heat transfer surface material were observed, but the general tendencies of heat transfer characteristics were highly similar. However, while Sindt asserts that heat transfer characteristics for TP liquid hydrogen and slush hydrogen (as shown in Figure 4) are the same, the authors found that, with the exception of the region near burnout, there were distinct differences between TP liquid hydrogen and slush hydrogen. As shown in Figure 5, Sindt observed strong hysteresis during increases and decreases in heat flux, but the authors observed
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only weak hysteresis. As shown in Figure 6, for the horizontal downwards orientation of the heat transfer surface, the heat transfer characteristics of TP liquid hydrogen and slush hydrogen were found to be the same by Sindt. However, the authors found differences in heat transfer characteristics between TP liquid hydrogen and slush hydrogen. In addition, the following general conclusions can be drawn from the experimental results: (1) For liquid hydrogen and slush hydrogen, hysteresis caused by the transition from nonboiling natural convection to nucleate boiling was weak. (2) Heat flux at burnout was the highest for NBP liquid hydrogen regardless of the orientation of the heat transfer surface. Furthermore, for normal liquid and slush hydrogen, heat flux at burnout was the lowest at the horizontal downwards orientation of the heat transfer surface.
CONCLUSIONS A heat transfer surface made of electrolytic tough pitch copper was placed in liquid or slush hydrogen at three different orientations (horizontal upwards, vertical, and horizontal downwards), and data for heat transfer characteristics in the nucleate boiling region was obtained. The same equipment was also used to obtain data for liquid or slush nitrogen. Differences caused by the orientation and material of the heat transfer surface as compared with the experimental results obtained by Sindt were clearly identified.
REFERENCES 1 Ohira, K. et al., Experimental investigation of production and density measurement of slush hydrogen Cryogenics (1994) 34 397-400 2 Coeling K. J. et al., Incipient and nucleate boiling of liquid hydrogen J. Eng. Indus. (1969) 91(2) 513520 3 Brentari, E. G. et al., Boiling heat transfer for oxygen, nitrogen, hydrogen, and helium NBS Technical Note 317 (1965) 4 Sindt, C.F., Heat transfer to slush hydrogen Adv. Cryog. Eng. (1974) 19 427-436
Figure 1 Schematic of glass dewar
Figure 2 Heat transfer unit
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Figure 3 Heat transfer of nitrogen, horizontal surface facing up
Figure 4 Heat transfer of hydrogen, horizontal surface facing up
Figure 5 Heat transfer of hydrogen, vertical surface
Figure 6 Heat transfer of hydrogen, horizontal surface facing down
Heat Transfer Characteristics of a Prototype Pool Boiling Superconductor to Liquid Helium A. lwamoto, T. Mito, K. Takahata, N. Yanagi, and J. Yamamoto National Institute for Fusion Science, 322-60roshi, Toki, Gifu 511)9-52, Japan
The heat transfer characteristics of a prototype superconductor to liquid helium was investigated as a function of heat transfer surface orientation and cooling channel width. The sample was a copper block with 18 x 18 mm and 72 mm long. The sample orientation was changed from 0~ to 90~ at the interval of 15 ~ The sample was surrounded by four stainless steel plates to make up a cooling channel. The cooling channel width was chosen as either 2 or 3 ram, furthermore measurements were conducted without a channel for comparison. The heat transfer characteristics of the prototype superconductor are compared with those of a copper plate.
INTRODUCTION The heat transfer characteristics of liquid helium for the stability analysis of pool boiling superconductors have been studied[I,2,3]. Usually, the samples were plates or thin wires. Recently, large sized superconducting devices are developed, for cxample, thc large hclical device(LHD) in NIFS[4]. Shape of the superconductors is different fiom the plates or the thin wires. The superconductors almost have rectangular cross section as the superconductor of the helical coil for the LHD. It is thought that heat transfer characteristics of the pool boiling superconductor are different from those of the plate sample because they are different shape of heat transfer surfaces. Liquid helium surrounding the superconductor in a magnet is limited by a cooling channel formed with insulating spacers and superconductors. In this study, for reliable stability analyses of the superconductors, the heat transfer of the prototype superconductor is measured as a function of the sample orientation and the cooling channel width. Heat fluxes of the prototype superconductor are estimated, using those of a copper plate. Comparisons between two samples are discussed. EXPERIMENT The heat transfer characteristics of the prototype superconductor to liquid helium have been investigated as a function of sample orientation and cooling channel width. The measurements were carried out in a glass dewar with saturated liquid helium. The schematic illustration of the sample is shown in Fig. 1. A polished copper block, 18 x 18 mm and 72 mm long, was used to simulate the superconductor for the helical coil. The surface roughness was less than 1() btm. Temperature difference between the surface and the liquid helium was mcasurcd using AuFc-Chromcl thermocouples. Thermocouples were mounted hi the place with apiczon grcasc as shown in Fig.1. To input hcat load for thc sample, a cylindrical shape heater was used and inserted into the ccntcr of a cross section with 18 x 18 mm. The sample was surrounded by four stainless steel plates to make up a cooling channel. The channel width was either 2 or 3 mm, furthermore mcasurcmcnts were conducted without a channel for comparison. The sample orientation was changed from ()~ to 9()~ at 15 ~ interval. The schematic illustration of a rig to change the sample orientation is shown in Fig.2. RESULTS AND DISCUSSIONS Dcpcndcncc of hcat transfer on samplc oricntation and channcl width Table 1 sulnmarizes the pool boiling hcat transfcr of thc prototypc supcrconductor. The heat transfcr curve fl)r each case was approximated as h(T)=a T+b. Figurc 3 shows thc dcpcndcncc of critical and minimum 605
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Figure 1 Schematic illustration of the prototype superconductor
Figure 2 Schematic illustration of a fig to change sample orientation
Table 1 The heat transfer of the prototype superconductor (at bath temperature of 4.2 K) Angle
Critical Heat Flux (W/cm2) (arT1 K)
Channel Width 2 mm 0~ 0.279 (at 4.69 30 ~ 0.359 (at 4.85 60 ~ 0.383 (at 4.77 90 ~ 0.366 (at 4.85 Channel Width 3 mm 0~ 0.338 (at 4.77 30 ~ 0.476 (at 4.61 60 ~ 0.494 (at 4.61 90 ~ 0.447 (at 4.61 Without a Channel (1~ 0.449 (at 4.69 30 ~ 0.5(i)9(at 4.69 60 ~ 0.508 (at 4.69 90 ~ 0.467 (at 4.68
Minimum Heal Flux (W/cm 2) (atT2 K)
Boiling Curve Nucleate
Transition
Film
T
b
T1 < T a
T2< T a
K) K) K) K)
0.0574 0.0674 0.0674 0.0664
(at (at (at (at
5.67 5.82 5.74 5.82
K) K) K) K)
0.569 0.552 0.672 (I.562
-2.39 -2.32 -2.82 -2.36
-0.227 -0.300 -0.324 -0.3(17
1.35 1.81 1.93 1.86
().()139 ().()211 0.0233 0.0241
-().0210 -0.0556 -0.0663 -0.0738
K) K) K) K)
0.0622 (al ().()687 (at ().(1686 (at ().()655 (al
5.74 5.74 5.74 5.74
K) K) K) K)
0.594 1.17 1.22 1.1()
-2.49 -4.91 -5.11 -4.62
-0.283 -().359 -().375 -().336
1.69 2.13 2.22 2.()()
0.0164 0.0214 ().()237 ().()243
-0.0320 -0.0540 -().()678 -().0737
K) K) K) K)
().()558 (at 5.82 ().()643 (at 5.82 0.0658 (al 5.66 0.0658 (at 5.82
K) K) K) K)
().92() 1.04 1.04 0.960
-3.86 -4.38 -4.38 -4.03
-().346 -0.391 -0.454 -0.354
2.()7 2.34 2.64 2.13
I).0187 0.0236 ().()256 0.0275
-().0529 -0.0731 -0.()791 -().0942
heat fluxes on sample orientation and channel width. The critical heat f l u x c s ( C H F s ) depended on the sample orientation and increased as the sample orientation increased from ()~ to 45 ~ and decreased from 45 ~ to 9() ~ The C H F s had a peak at 45 ~ The C H F s were significantly affected by the channel width and were i m p r o v e d as increasing the channel width. The channel width of more than 3 mm makes the hc~t transfer be as s a m e as that without a channel. It is a s s u m e d that a narrow cooling channel prevented liquid helium from being supplied to the heat transfer surface. On the other hand, the m i n i m u m heat f l u x e s ( M H F s ) were less dependent on the surface orientation as well as the channel width than the critical heat fluxes. According to these experimental results, it is thought that the stability of a pool boiling superconducting m a g n e t with more than 3 m m channel width is as same as that of its s u p e r c o n d u c t o r .
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Comparisons between the prototype superconductor and the copper plate The heat transfer characteristics of the copper plate with and without a channel were already reported[ 1,5]. The heat transfer characteristics were compared with those of the prototype superconductor. Whole surfaces of the prototype superconductor were exposed to liquid helium. The sample orientation was changed as shown in Fig.2. Therefore, top and bottom surfaces can be changed with a gear, while the side surfaces are kept as constant. If the sample orientation is x ~ the top and the bottom surfaces are x ~ and 180~ ~ respectively. The heat transfer characteristics of the copper plate cannot be compared with thosc of the prototype superconductor since only one surface was exposed to the liquid helium for the plate. To compare be~,een the heat fluxes of the ~,o samples, a rule of mixtures on the surface orientations in the prototype superconductor was defined as following expression.
• HFptate ,x HFproto = x HFproto HFplat< x x
(1)
4 9Calculated heat flux of the prototype superconductor : Measured heat flux of the copper plate at x :Surface orientations of x ~ 90~ and 180~ ~ in the prototype conductor
The heat fluxes of the prototype superconductor were estimated, using those of the copper plate and eq.(1). To calculate the heat fluxes, it was assumed that there was no temperature gradient in the sample and only the surface orientation affects the heat transfer characteristics for each surface. The critical and the minimum heat fluxes of the prototype superconductor without a channel is shown in Fig.4, comparing calculations with measurements. The calculated critical heat fluxes almost agree with those of measurements. On the other hand, the calculated minimum heat fluxes are approximately 1.5 times larger than the data obtained from measurements. Figure 5 shows the calculated heat flux without a channel at 30 ~ using eq.(1), compared with measured result. In only film boiling region, the measurement is about 35% smaller than the calculation. The heat transfer coefficient depends on supply of liquid helium for a heat transfer surface. It is thought that, in film boiling region, LHe supply for each surface in the prototype superconductor is less than that of the copper plate because of sample shape. Then, critical and minimum heat fluxes of the prototype superconductor with 3 mm of channel width were tried calculating, using those of the copper plate with a channel. The calculated and the measured results arc shown in Fig.6. The calculated critical heat fluxes take a different tendency from the data of the measurements. It is thought that different channel arrangement bem, een the copper plate and the prototype superconductor affects the critical heat fluxes. Channel volume for each surface of the prototype superconductor was bigger than that of the copper plate so that the calculated CHFs underestimated. On the other hand, the measured minimum heat fluxes with the channel are also about 35% smaller than those of calculations. The minimum heat
Figure 3 Dependence of the critical and the minimum heat fluxcs on sample orientation
Figure 4 Critical and minimum heat fluxes without a channel (Comparison between calculated and measured results)
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Figure 5 Calculated and measured heat fluxes at 3() ~
Figure 6 Critical and minimum heat fluxes with a channel (Comparison between calculated and measured results)
fluxes have less influence of the channel than the critical heat fluxes. According to comparing the calculations with the measurements, the heat fluxes of the prototype superconductor with the channel cannot be estimated, using those of the copper plate and cq.(1). It was found that the heat transfer measured by the prototype superconductor made the stability analyses of superconducting magnets more reliable than the analyses applying that of the copper plate. CONCLUSIONS The heat transfer characteristics of the prototype pool boiling supcrconductor to liquid helium was investigated as a function of the sample orientation and the cooling channel width. The measured results were compared with those of the copper plate. The following conclusions were obtained; The critical heat fluxes depended on the sample orientation and the channel width. The CHFs had its maximum at 45 ~ On the other hand, the minimum heat fluxes were lcss dcpcndcnt on the sample orientation and the channel width than the CHFs. The channcl width of more than 3 mm makcs the heat transfer characteristics be as same as that without a channel. The heat transfer characteristics of the prototype superconductor were calculated using those of the copper plate and a role of mixtures on the surface orientations. The critical heat fluxes without a chmmcl could bc estimated, but the calculated CHFs with a channel took a different tendency from those of the measurements. On the other hand, the minimum heat fluxes both with and without a channel obtained flom measurements were about 35% smallcr than thosc of calculations in only film boiling rcgion. It was found that the heat transfer measured by the prototype superconductor made the stability analyses of the superconducting magnets more reliable than thc analyses applying that of the copper platc. REFERENCES
l:ll 121 i-31 141 [51
lwamoto, A. ct al., Heat Transfer of a Large Copper Plate to Liquid Helium Applicable to Mlrgc Scalc Superconductors Cryogenics (1994) 34 ICEC Supplement 321-324 Chandratillckc, G.R., ct al., Pool boiling heat transfer to saturated liquid helium from coated surface Cryogenics (1989) 29 588-592 Chcn, Z., ct al., Channel heat transfer in Hc I - steady statc orientation dependence Advances in Cryogenic Engineering (1986) 31 431-438 Yanagi, N., ct al., Experimental observation of anomalous magncto-rcsistivity in 111 - 20 class aluminum-stabilizcd supcrconductors fi)r the L lrgc Helical Device Advances in Cryogenic Engineering (1994) 4!1 459-468 lwamoto, A. ct al., Heat transfcr from an oxidized large copper surface to liquid hclium: Dcpcndcncc on surface oricntation and treatment to be published in Advances in Cryogcnic engineering 41
Temporal Deterioration of Helium Heat Transfer at Moderate Pulse Heat Load
Yuri P. Filippov and Igor A. Sergeyev Particle Physics Laboratory, Joint Institute for Nuclear Research, Dubna 141980, Russia The report presents new experimental data on helium boiling dynamics demonstrating a short-term reduction of the heat transfer intensity. The experiment is carried out in a static He I bath under saturation. Temperature histories are measured of the cooled surface at single pulse heat load input. The heat pulse amplitude is close to the steady-state value of critical heat flux and the leading edge is in the range of 20 #s (step pulse) to 1 s (quasi-steady heating). Temporal deterioration of heat transfer exhibits itself as a pike of superheat. The pike characteristics, amplitude and width, are obtained for the various parameters of the heat load pulse.
INTRODUCTION Stability margin is the key design parameter of a cryogenically cooled conductor carrying high density current. Underestimation of the margin results in the inefficient system operation or poor characteristics. From the other hand, overestimation can cause breakdown: burnout of an ordinary conductor or quench of a superconductor. The modern approach to the stability analysis is to consider the transient pattern of heat transfer from a conductor to a cryogen in detail. Since the pilot experiment performed by Jackson [1] numerous investigations have been carried out of transient helium heat transfer. Most of them deals with the delay of film boiling transition, t~, (referred as "incipience of crisis", "moment of departure from nucleate boiling", "takeoff time") and its dependence on heat flux, q, and bulk fluid temperature, Tb. It has become common to correlate t~ with q by the power-law dependence: t ~ = f u n c ( T b ) . q -'~ at q>>q~t or t r -m at q>__q~t, where n ~ 2 , m ~ 2 . 4 [2] and qst is the steady-state heat flux of nucleate boiling crisis. There are a few models of the first stages of transient heat transfer predicting such a behavior: vaporization of diffusion layer [3], coalescence of vapor bubbles [4], evaporation of liquid macrolayer [5], and others representing some combinations of the listed mechanisms. The curve t~(q) is usually accepted as the bound of the dynamically stable states on the {q; t} phase plane. Nevertheless, some scanty information is available showing the existence of one more bound which reduces the stability domain. These data are the superheat pikes (AT(t) overshoots) occurring at the heat transfer surface in a few hundred milliseconds after the input of moderate heat load, q < q~t, see [6,7,8]. Because of tile amplitude of such overshoots can be as high as a few Kelvin, the corresponding {q; t}-region has to be regarded as the domain of temporally deteriorated heat transfer. The present work is the first one devoted to the systematic study of the temporal deterioration and discussion of its nature.
EXPERIMENTAL TECHNIQUE Experiment is carried out with a cylindrical test sample (length:diameter .~ 6:1, see Fig.l) placed in a static He I bath under saturation. A carbon fihn of 1 #m thickness is deposited onto its surface and serves both as a heater and thermometer. Sample axis orientation is either 0 ~ or 90 ~ with respect to gravity. Some tests run under the so-called confined conditions when the sample is put coaxially into a thin-walled tube (see Fig.l) forming an annular channel with the open outlets. The width 609
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Figure 1 Test sample design and the way of confinement (all the dimensions are in millimeters). of annular gap between the heat transfer surface and the restricting tube is either 0.4 or 1.3 or 2.25 or 4.0 mm. Helium bath pressure is kept constant within a few percent with an automatic system regulating the rate of He vapor pumping-off. Bulk fluid temperature range is 2.2 to 4.9 K. To avoid the drift of thermal power, W, generated in the heater-thermometer while its heatingup, the special power supply electronics [9] is used. With a feedback circuit, it controls continuously just the product W(t)=voltage • current. Experiments are performed with the pulse heat load of trapezoid shape: W(t; r ) = Q . t/r at t < r and W(t; T)--Q--const at t > r. The leading edge r is in the range of 20 #s (step-like pulse) to 1 s (quasi-steady heating) and the amplitude ranges from 0.2 to 20 Watts (heat flux from 100 to 104 W/m2). Measurements are made of the normalized values of both the current and voltage drop across the heater-thermometer with the 12-bit 33 kHz 2-channel 8K-storage ADC. A microcomputer is aided to control the power supply and readout systems [9] and to facilitate the data processing. The readout circuit is calibrated before the every measurement session with the substitution of a standard resistor (resistance from 1 to 6 kOhm, accuracy 10 -4) for the test sample. Resulting errors are: superheat - below 10 mK at 2.2 < T < 5.2 K and about 1% at T > 5.2 K; thermal p o w e r - about 4% at 0 < t < ~" and below 1% at t > r. Figure 2 shows a representative pair of A T and W histories. For further analy.sis, the measurement results are processed in the following manner. The 'j-set' of data, (AT, W, t)~, i - 1 . . . Nj, is fitted by: spline functions built on the unequally spaced nodes - for the power W(t)j, where for the superheat AT(t)j ; and curve Q.(t/r).r162 r is the sigmoidal function. Standard deviation e of the Q fit is ~0.2%. Figure 2 Typical traces of the pulse power W(t) and superheat AT(t) at the absence of confinement, vertical sample orientation, Tb - 4.24 K, Q - 7.97 +0.01 W, T - 0.2 s. (W(t) at t < T does not seem straight due to the logarithmic scale).
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Figure 3 General pattern of transient helium heat transfer at the absence of confinement, vertical sample orientation, Tb = 4.224 K, T -- 1.11.10 -3 s. DISCUSSION Results are obtained for more than 50 combinations of the experimental conditions. Some of them are presented in Fig.3 and exhibit only the minor pikes AT(t..~0.11s) at Q = 7.5 and 6.5 W. As the Fig.4 shows, another combination of parameters is more favorable for an overshoot to be pronounceable. In dependence of the experimental conditions, either the single pike is observed, or two separate pikes, or two overlapping pikes, or even some irregular oscillations (multiple random pikes). The overshoots reproduce reasonably well at the repetitive runs (with the exception of the particular case when the heat load is very close to the heat flux q~t of steady-state boiling crisis and the multiple random pikes occur). Such a regularity is in contrast to the purely stochastic behavior observed in liquid nitrogen [10]. The contradiction can be reasoned by the features of transient nitrogen heat transfer as well as by the different ways of thermal power supply (experiments [10] did not employ the stabilization of power and so allowed the considerable drift of Q). Quantitative characteristics of an overshoot, namely the amplitude A, center B, and width C are obtained with the fitting procedure illustrated in Fig.4. Some of the results are presented in Fig.5 and demonstrate that the overshoots do exist over the broad range of T, i.e. at varios degrees of the power unsteadiness Q,/T including those ones corresponding to the quasi-steady heating. So, the power transients themselves do not cause the pikes of superheat. Thus, the origin of A T ( t ) overshoots lies in the basic mechanism of boiling helium heat transfer. Figure 4 Overshoot of superheat at the presence of confinement (1.3 m m annular gap), vertical sample orientation, T b - 4 . 2 4 K, Q - 5 . 3 0 W, T-- 2.50.10 -2 s. Model" Sum of two Gaussians A o + Z A i exp( t - B i )2/Ci2 ) 2
=0.00027
A o= 0.492488
+ 0.0027312
AI = 0.719407
+0.044168
A2 = 0.185515
•
B1 = 0.226281
• 5.6516E-4
B2 = 0.160730
+0.011372
C 1 = 0.0341590 + 0.0013649 C 2 =0.0711444 a- 0.010229
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Figure 5 Characteristics of superheat pikes at the presence of confinement (1.3 mm gap), vertical sample orientation, Tb = 4.24 K, Q =5.29 W (error bars indicate the fitting errors). Analysis has been made over all the collected transient patterns and the data on steady-state boiling under the same conditions. It shows that the AT(t) pikes appear at heat loads satisfying the condition of q,f _ q < q~t, where q~] corresponds to the steady-state crisis of film boiling and q~t the nucleate one. This proves that the overshoots are caused by the unstable nature of the so-called "mixed boiling regime". While the heat flux into helium q is fixed or controlled, the state of coexisting nucleate + film boiling is unstable in the majority of heat transfer configurations. So, the overshoots AT(t) can appear at the moderate values of thermal load q < q~t and affect the operation of a helium-cooled device that is believed to be dynamically stable. ACKNOWLEDGMENTS Authors would like to express their gratitude to Mrs. T.I.Smirnova assisted to manage the large database of the measurement results. The work was made possible in part by the support of Russian Foundation for Basic Research under grant 95-02-05668. REFERENCES 1 Jackson, J. Transient heat transfer and stability of superconducting composites Cryogenics (1969) 9 103-105 2 Sakurai, A., Shiotsu, M. and Hata, K. In: Advances in Cryogenic Engineering Plenum Press, New York, NY (1994) 39B 1759-1768 3 Schmidt, C. Transient heat transfer to liquid helium and temperature measurement with a response time in the microsecond region Appl. Phys. Lett. (1978) 32 827-829 4 Pavlov, Yu.M. and Babitch, V.I. Transient burn-out in liquid helium with rapid rise of heat flux Cryogenics (1987) 27 641-644 5 Giarratano, P.J. and Frederick, N.V. In: Advances in Cryogenic Engineering Plenum Press, New York, NY (1980) 25 455-466 6 Schmidt, C. and Turowski, P. In: KfK-Report No. 1958 Kernforschungszentrum, Karlsruhe, BRD (1974) in German 7 Steward, W.G. Transient helium heat transfer. Phase I - Static coolant Int. J. Heat Mass Transfer (1978) 21 863-874 8 Kingsbury, D.L., Huang, X. and Van Sciver, S.W. In: Advances in Cryogenic Engineering Plenum Press, New York, NY (1994) 39B 1631-1638 9 Minashkin, V.F. et.al. In: JINR Communications P10-88-902 Joint Institute for Nuclear Research, Dubna, USSR (1988) in Russian l0 Malkovsky, V.I. et.al. Transient nucleate boiling of liquid nitrogen with a stepwise change of heat flux Cryogenics (1992) 32 1131-1136
ClTogenic Characteristic Investigation on Heat Transfer between Gas and Solids of an Adiabatic MovingBed
Youyi Guo.* Zizhi Li,* Li Wang,** Liang Zhang** * School of Chemical Engineering, Xi'an Jiaotong University, Xi'an, Shaanxi, 710049, P.R. China **ClTogenics Laborato~T, Chinese Academy of Sciences, Beijhlg, 100080, P.R.China
A rubber thermal conductivity profile vs. temperature at low temperature is presented tS"stly. B e l o w - 8 0 :'C the heat transfer coefficient between gas and solids is measured, and based on which the con'elation in mahl factors influencing heat transfer of an adiabatic moving-bed is obtained. The results are useful not only fbr desigming a scrap th'es cooling chamber, but for making up deficiency of research in heat transfer between gas and solid at low temperature.
INTROD U C TI ON Ill a veVtical adiabatic bed the solid pavticles of tire and c13'ogenic gas are flowing at uniform velocities oppositely. A! the exit of bed scrap tires can be chilled down to their embrittlement temperature by contact with the cryogenic gas. The characteristic of this kind of heat transfer is velT complex, because there aie so many influencing factors, such as the distribution of gas velocity and temperature, size and relative displacement of the pal"ticles during flow ect. Attention has rarely been paid to this fiekl, especially in cryogenic range. The cooling velocity of the scrap tires with certain size is chiefly depended on its thermal conductivity, and the experiment results of cooling scrap tires are quite different fiom the results of heating scrap tires [!], so it is necessal3/to measure thermal conductivity of rubber. MEASUREMENT OF RUBBER THERMAL CONDUCTIVITY In order to get the thermal conductivity of the tire robber its specific heat must be measured at tirst. The experimental installation is shown ill figure 1.
1-- thermocouple of surface temperature 2 - - thennocouple of center temperature 3 - - a piece of test tyre rubber 4 - - vacuum space 5-- electric heater Figure I measuring set-up for the rubber specific heat After tile center temperature of test tyre rubber drops to tile liquid nitrogen temperature by immersing it in liquid nitrogen, a certain quantity of heat Q is given to the rubber at a vacumnized state. The temperature increase AT of the rubber is measured when it reaches equilibrium. Therefore its specific heat is given as fbllows: 613
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Cp = Q / (M . AT)
In terms of the transient heat conduction theory the thermal conductivity of rubber profile with temperature can be measured. A long cylinder initially at uniform temperature to is cooled by immersing it in a liquid of lower temperature too. The dimensionless temperature distribution of time dependence within the cylinder is got by assuming cx and rc to be constant [2].
|
/9
= ~
O0
=
t - to~ t o - t~
71 = o o
X--" '
- z.., A .
,,=l
-- a z ' ~ p2
U.E
(2)
R~
equation (2) may be expressed as
aTs
0 = A U exp ---R-g-)
=AUe
-=~
, Fo>0.55
(3)
where m, cooling velocity of the cylinder, can be expressed as m = ln(9, - l n 9 2 ) / ('t2
-
(4)
"l~l)
where 01, t92 is the temperature difference at time T1, 7;2 respectively, hence, the thermal diffusivity a=m
9
,
e =2.4048
(5)
consequently, thermal conductivity is stated as (6)
~:= a . p . C p
EXPERIMENT RESEARCH The experimental set-up for measuring the thermal conductivity of tyre is shown in figure 2, and the expe-
1-- switch unit 2, 3, 4 - - thermocouple 5-- multimeter 6 - - liquid level-gauge 7 - - induction regulator 8-- liquid nitrogen dewal
Figure 2 experimental set-up for rubber thermal conductivity rimental results are shown in figure 3. The variation of thermal conductivity near the embrittlement temperature is significant, and the maximum thermal conductivity appears at about -30 ~ It is clear that the heat transfer velocity descents along with the rubber temperature decreasing.
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Figure 3 thermal conductivity profile with temperature of tire Temperature distribution, with respect to a one-dimensional unsteady heat conduction, can be obtained by Heisler Chart at some certain boundary condition. On the contrary some unknown information also can be solved by Heisler Chart at a definite temperature distribution, such as the heat convective coefficient between the surface and fluid [3 ]. For simplicity some assumptions are essential: (1) Dt/Dp>> 10, so the wall effects and the non-tmiform of gas velocity distribution can be neglected. (2) The moving-bed should be consistent with title, so the error associated with neglecting the nonuniform of radial temperature distribution is small. (3) Radiation, conduction along axial direction are ignored. (4) Thermal resistance between particles and the effects arised by movement of particles are ignored also. Based on these assumptions the heat transfer forms in bed consist of heat convection between gas and solid and unsteady heat conduction within solid. The scheme of experimental system is shown in figure 4. The bed length H is 1.5 m, with diameter Dt = 140mm. Initial temperature of gas nitrogen is -80 ~ at velocity V=0.75 - 1.25 m/s, while the Prandtl number is assumed as a constant. Scrap tires, with equivalent diameter Dp=5.5, 7.2, 14.5 mm~ enter the top of the bed with velocity of 0.6 mm/s. Gas-solid temperature distribution of Dp=7.2mm is shown in figure 5. Using the thermal conductivity of rubber above, the heat convective coefficient between gas and particles is obtained, the relations of ct with h, Dp, and V is illustrated in figure 6. The average heat convective coefficient for the whole bed is shown in table 1. 1--LN 2 supply tank 2 - - L N 2 control valve 3--vaporizer 4--mixed regulated container 5--mixed gas pressure build-up valve 6--flux meter 7--moving-bed 8--insulator 9--screw propeller 10~filtrator 11--compressor 12--regulated container 13--oil-gas separator 14--air pressure build-up valve Figure 4 the experimental system
Figure 5 the temperature distribution of scrap tires and gas with Dp=7.2mm
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Table 1 Dp(mm) 5.5 7.2 7.2 7.2 14.5 14.5 14.5 ,
V(m/s) 1.00 0.75 1.00 1.25 0.75 1.00 1.25
' ,
|
Re 558 547 730 912 1102 1470 1837
c~(w/m2.k) 66.323 48.458 52.010 63.121 35.319 46.048 48.926 .,
c~'(w/m2~k) 64.174 48.458 57.384 65.376 36.269 42.938 48.926 ,,,,
Nu 18.24 17.44 16.68 22.72 25.61 33.39 35.47
errors(%0) 3.2 0.0 9.9 3.6 2.7 6.7 0.0
Figure 6 the profile of c~ with Dp, V and h ( distance from the top of the bed ) The Nusselt number and Reynolds number in table 1 can be formulated by regression as follows:
(7)
Nu = 0. 4342259. Re~ Hence, a ' = 0.4342259. Re ~
2 / Dp
and the range of applicability of equation (7) and (8) is: (1) Dp=5.5mm- 14.5mm. (2) V=0.75m/s- 1.25m/s. (3) t=-80~
(8) - 0~
CONCLUSION It is clear that the variation of thermal conductivity of tyre rubber near the embrittlement temperature is significant, and the maximum thermal conductivity appears at about -30 ~ A Nu-Re correlation is obtained for different size of scrap tires and different gas velocities at low temperature. REFERENCES [ 1] Youyi Guo, Xuesong Li, Yan Liang Characteristic Investigation on Unsteady Heat Transfer between Gas and Solids Cryogenics, Vol.34.P.369 ICEC Supplement, 1994 [2] S.M.Yang, Heat Transfer, Second Ed., Higher Education Press, Beijing, 1987 [3] H.J. Zhang, Heat Conduction, Higher Education Press, Beijing, 1992
Transient Heat Transfer from a Silver Sheathed Liquid Nitrogen
High-Tc Superconducting
T a p e in
Masahiro Shiotsu, Koichi Ha.ta, Akira. Sa,kurai*, Chizuru Suza.wa**, Shigeki Isojima** and K enichi Sa.to** I n s t i t u t e of Atonfic Energy, Kyoto University, Uji, Kyoto 611, J a p a n . F u t u r e Energy Resea.rch Assoc., Pa.steur Building, 103-5 Ta.na,ka Monzen-cho, Sakyo-ku, Kyoto 606, Ja.pan * * S u m i t o m o Electric Industries Ltd., 1-1-3 Shimaya, I(onohana,-ku, Osaka. 554, J a p a n The transient heat transfer from a silver sheathed BiPbSrCaCuO tape in liquid nitrogen was measured for the exponential heat inputs, Qoe t/r, with the period, 7-, ranging from 10 ms to 10 s. Average temperature of the sheathed tape was estimated by using the result, of newly developed estimation method. The critical heat flux (CItF) increases slightly with the decrease of 7- from 10 s to 1 s, decreases significantly at 7- around 900 ms, slightly increases for 7- down to 100 ms, and increases almost proportional to 7--1/2 for furt.l~er decrease of 7- down to 10 ms. Direct transition from non-boiling to film boiling was observed for 7- < 900 ms. The CHF at the periods of 900 ms and 10 ms are about 15 % and 40 % respectively, of the steady-state CHF.
INTRODUCTION The knowledge of steady and transient heat tra.nsfer characteristics on a sheathed high Tr superconducting tape in liquid nitrogen (LN2) for a very large electric current under an abnormal condition is necessary for a stability design of various electric appa.ratus using high-T~ superconducting tapes. However, there have been few studies on the steady-state a.nd transient heat transfer from a sheathed high-T~ superconductor in LN2 due to a current beyond its critical value, Ir On the other hand, several resea, rchers [1, 2] studied the steady and transient heat transfer from fine normal resistive wires in LN2 and reported tha, t there existed direct transition (they called this as premature transition) from transient conduction heat transfer to fihn boiling without nucleate boiling caused by a step heat input at atmospheric pressure. Some of the authors [3-5] studied the heat transfer from a 1.2 mm diameter horizontal cylinder in LN2 and reported that the direct, tra, nsition occurred not only in transient conduction heat transfer caused by rapidly increasing heat input, but also in natural convection heat transfer caused by quasi-steadily increasing heat inputs at va,rious pressures near atmospheric. They [5] suggested that the direct transitions occur due to the instantaneous levitation of the liquid on the solid surface by explosive initiation of heterogeneous sponta.neous nuclea.tion in originally flooded cavities on the surface. The aim of the present study is to develop a,n estimation method of transient temperature for Ag sheathed high-T~ superconducting tape, and to clarify by using this method the transient heat transfer characteristics on the sheathed high Tr tape in LN2 for exponentially increasing heat inputs with a wide range of periods 7- from quasi-steady to rapidly increasing ones. APPARATUS
AND
METHOD
Pool Boiling A p p a r a t u s The experimental appara.tus is shown schematically in Fig. 1. The test vessel (1) is a vacuum insulated, cylindrical pressure vessel of 20 cm inner diameter and 70 cm height, with a lnaxinlum working pressure of 3.5 MPa. The vessel has two sight ports (15) and is equipped with a sheathed I mm diameter CC thermocouple to measure bulk liquid teml)erature. The vessel is connected via. a valve to a liquid nitrogen feed tank (7). System pressure in the test vessel is automa.tically controlled within -4- 1 kPa of a desired value by a pressure control system consisting of a pressure transducer, a pressure controller (10) and a control valve (9). A silver sheathed BiPbSrCaCuO high %. superconducting tape of 80 mm in length, 3.2 mm in width and 0.23 mm in thickness was used as the test sample. The critical current I~ at 77 K and critical temperature T~ for the sample are 39.3 A and 1 10 K, rest)ectively. This test sample was horizontally supported in the test vessel with its width direction vertically oriented. Potential taps were soldered on the sample surface with the length between them of 46.3 ram. 617
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Experimental Method The test sample was heated by an electric current from a power amplifier (12 V, 600 A) whose input signal was controlled by a high-speed hybrid computer so as to give a desired time function for the heat input, in spite of rapid variation in the electrical resistance of the sample due to the variations of electrical current and t e m p e r a t u r e of the test sample. A double |)ridge circuit with the test sample as a branch was used to measure the electrical resistance of the sample. The signal voltage of the bridge circuit, together with the voltage drops across the potential taps of tlle test sample and across a standard resistance, was amplified and passed through analog-to digital converters to a digital computer. The average temperature of the test sample, T, was estimated from the measured values of its electrical resistance, R, and electrical current through the test sample, I, by the following way: First, the following empirical equation of the critical current for silver sheathed BiPbSrCa.CuO high T~ tapes with various [c values at 77 K was obtained by fitting the Ic versus temperature rel~ttions reported by Sato et al. [6].
I~ =
S~,o[1.o-
(T/T~)~~] 2~
1 Test Vessel (Cryostat) 2 Test Heater 3 Potential Taps 4 Electrodes 5 SightPort 6 Sheathed Heater
7 8 9 10
LN2 Feed Tank Pressure Relief Control Valve Pressure Controller 11 He Gas 12 Vacuum Pump
(PT) Pressure T~ Transducer Thermocouple ~ Resistance ~ Thermometer ( ~ Liquid Level Transducer ( ~ Pressure Gauge
Fig. 1 Schematic of experimental apparatus
(1)
By inserting the I~ = 39.3 A at 77 K and T~ = 110 K into Eq. (1), the L=,0 value was obtained to be 237 A. The equation of Ic for the test saJnple is their, Ic = 23711.0 -(T/110)15] TM
(2)
Second, the following function was selected to express approximately the electrical resistance of tlle test sample, R, based on the voltage w~rs~ls transport current curve for the same high-Tc tal)e of 200 mm in length in LN2 shown in Fig. 2. R = R N ( T ) [ 1 . 0 - 1.0/{(1 - a ) + a ( l / [ c ) ' " } ~
(3)
where R N ( T ) is the electrical resist.alice of tile Ag sheath correlated by RN(T) = 1.505 x :10-:311 +,1.128 x 1 0 - 3 ( T - 273.15)], and the wdues of a and m. are around 0.1 and 5, respectively. Definite values of a and m cannot be determined from the V-] curve because the temperature of the tape for each current is unknown. On tlle other hand, theoretical expressions for non-boiling heat transfer coefficients oil a w-:rtical plate heated by exponential heat inputs with various exponential periods r had already been o btailled by some of the authors (see Appendix). If the values of a and m in Eq.(3) are given, the aw.'rage temperature of the test sample, T, at each time ('a.~l I)e obtained as the value satisfying Eq.(3) and Eq.(2) siJnultaJleously for the measured values of I and R at the tilne. The values of a and m in Eq.(3) were deternlilled a.s follows by a trial and error method for the non-boililtg heat transfer coefficients thus obtained oJ~ the test sample due to exponential heat inputs to agree better with the corresponding theoretical values. R = R N ( T ) [ 1 . 0 - 1.0/{0.86 + 0.14(I/I,:)4s} ~
(4)
Fig. 2 V-I curve for tlle same high-To tape of 200 mm in length
Fig. 3 Time traces of V, I, q and ATtar for quasi-steadily increasing heat input
I C E C 1 6 / I C M C Proceedings RESULTS
619
AND DISCUSSION
Figure 3 shows the time traces of electrical current I, terminal voltage between the potential taps, V, heat flux, q, and average temperature increase, ATtar for a quasi-steadily increasing heat input with r of 8.98 s. As shown in the figure, I and V were varied very complicatedly to realize the desired exponential heat input. The relations of V, I, q and ATsat at each time can be regarded as the steady-state combined electrical and heat transfer characteristics for the test-sample in a saturated pool of LN2 at atmospheric pressure which were obtained for the first time by using the newly developed estimation method of the average test-sample temperature. Figure 4 shows typical heat transfer processes for exponential periods of 8.98 s, 87.8 ms, 27.6 ms and 14.3 ms on the graph of the heat flux, q, versus wall superheat, ATtar. Theoretical non-boiling heat transfer curves for corresponding periods predicted by Eqs. (5) and (6) in the Appendix are also shown in the figure for comparison. The process for the heat input with the period of 9.8 s almost agrees in the non-boiling regime with the theoretical curve for natural convection heat transfer predicted by Eq. (6). The transition from quasi-steadily increasing natural convection to fully developed nucleate boiling ( F D N B ) occurs with rapid reduction in the surface superheat just after the initiation of boiling at the surface superheat of about 10 K. The heat flux then gradually increases with the increase of the wall superheat along the FDNB curve to the critical heat flux where the transition from FDNB to film boiling occurs due to the hydrodynamic instability. The transient heat transfer process for the period shorter than 100 ms is almost in agreement with the corresponding theoretical curve for the conduction heat transfer given by Eq. (5) before the initiation of boiling; the heat transfer coefficients are higher for shorter exponential periods. At the initiation of boiling at about A:/~t -- 20 K, the whole cylinder surface is covered with a thick vapor film and the heat flux rapidly decreases. Direct transition from non-boiling to fihn boiling is observed for the period shorter than 900 ms. Figure 5 shows the results of transient critical heat flux, qcT, versus r. Open circle shows the qcT for FDNB, and solid circle shows that for direct transition. As shown in the figure, the q,:r for FDNB increases slightly with the decrease of 7 from 10 s to 1 s. The qc~ decreases significantly at r around 900 ms due to the occurrence of direct transition, slightly increases for the decrease of 7 down to 100 ms, and increases almost proportional to 7 -1/2 for further decrea.se of r down to 10 ms. It should be noted that the q~ values for r = 900 ms and 10 ms are about 15 % and 40 % respectively, Fig. 4 Heat transfer processes for various r of the steady-state CHF.
Fig. 5 Critical h~at flux versus exponential periods, r
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CONCLUSIONS
An estimation method for transient tempera.ture of a,n Ag sheathed BiPbSrCaCuO tape due to an over current (I > Ic) was developed. The transient heat transfer from the Ag shea.tl~ed ta,pe in liquid nitrogen was measured for the exponential heat inputs with the period, r, ra.nging from 10 ms to 10 s by using the newly developed temperature estimation method for the tape. Transient non-boiling heat transfer coefftcients for the tape was confirmed to be in good agreement with the corresponding theoretical values dependent on 7-. The q~, for r on the tape can be classified into two groups: that for FDNB and that due to direct transition. The direct transition from non-boiling to film boiling without nucleate boiling was observed for 7- _ 900 ms. The q~, due to direct transition is significantly lower than that. for FDNB" the q~, for the periods of 900 ms and 10 Ins, for instance, are about 15 % and 40 % respectively, of the steady-state CHF. REFERENCES
1. Sinha, D.N., Brodie, L.C., Semllra, J.S., and Young, F.M., Premature transition to stable fihn boiling initiated by power tra.nsients in liquid llitrogen, Cryogenics, (1979) 19, 225-230 2. Tsukamoto, O., and Uyemura, T., Observation of bubble formation mechanism of liquid nitrogen subjected to transient heating, A(lvances in cryogenic eng. 25, (1980), 476-482 3. Shiotsu, M., tlata, K., and Sitkurai, A., lieterogeneous spontaneous nucleation temperature on solid surface in liquid nitrogen, Advances in cryogenic eng. 35, {1990) 437-445 4. Sakurai, A., Shiotsu, M., and ]lata, K., Boiling heat transfer from a horizontal cylinder in liquid nitrogen, Heat transfer and superconducting magnetic energy storage, (1992) ASME tITD-Vol.211, 7-18 5. Sakurai, A., Shiotsu, M., and Ha.ta. I,:., Ne.w tra.nsition phenomena to fihn boiling due to increasing heat inputs on a solid surface in 1)ressurized liquids, Instability in two-phase flow systems, (1993) ASME HTD-Vol.260, 27-39 6. Sate, K., and ttikata, T., Critical currents of superconducting BiPbSrCaCuO tapes in the magnetic flux density range 0-19.75 T at 4.2, 15 and 20 I(, Appl. phys. lett., (1990) 57, 1928-1929 7. Sakura.i, A., and Shiotsu, M., Transient pool boililig heat transfer, part 1, incipient boiling superheat, ASME J. tteat Transfer, (1.(_)77) 9i), 547-553 8. Takeuchi, Y., Hata, K., Shiotsu, M., a.~t(l Sa.kurai, A., A gener;d correlation for natural convection heat transfer from horizontal cylinder i~ liqltids a.J~d gases, General l)a.l)ers in heat transfer, (1992) ASME HTD-Vol.204, 183-189
APPENDIX
C o r r e l a t i o n for N o n - b o i l i n g H e a t T r a n s f e r d u e to E x p o n e n t i a l H e a t I n p u t s Sakurai and Shiotsu [7] rel)orted that the nonl)oiling ]lea.t transfer coefficients on a vertical fiat plate due to exponential heat inputs, Qo ct/r, wil/t tile l)eriods 7- shorter than 100 nts c~tn be described by the theoretical values derived from thermal coltduct.ioz~ eq~la.lioll for the liquid, and those for r longer than 1 s by the theoretical values for natura.l colivection ltt,.at trallsfer. The conduction heat transfer coefficients, h,.:, a.r(: M)proxintately given by the following equation at a time longer than t = 3r after tlle initia.lion of tl~e heat inl)ut [7].
(kzpt%z/r) 112
hc-
(5)
Laminar natural convection heat transfer coefficients, b,,,, on a vertical plate with the height H for wide ranges of Rayleigh and Pra, ndtl lllllnbe.rs are exi)rcsscd by the following equation [8]. hn = 1 . 2 ( k l / t i ) x
10 y
(6)
where Y = 0.193.185 + 0.1,t50:37X + (I.6G.132:~IO-"X 2 - 0 . 2 3 9 , t 3 9 1 0 - 3 X 3 - 0.23861310-4X 4, R.f = G r*t:'rz/(,l + 91"r 1/~ + 10P/') .
X = loglo(R.f),
Non-boiling heat transfer coetficient.s for the il,t(:rlne(liate region, h,,~, (1 s > 7- > 100 ms) can be expressed by the following equation[7]. -
+ h,",.
(7)
Surface T r e a t m e n t of A l u m i n u m Heat S w i t c h
Toshinobu Shigematsu, Minoru Maeda, Masatoshi Takeshita, Yoshiko Fujii, Masaki Nakamura, Minoru Yamaguchi, Toyoichiro Shigi and Hiroshi Ishii* Dept. of Applied Phys., Okayama Univ. of Science, Ridaicho 1-1, Okayama 700, Japan *Okayama Ceramics Research Foundation, Nishikatakami 1406-18, Bizen, Okayama 705, Japan
Although aluminum is the most promising as a material of the superconducting heat switch at ultra low temperatures, it has not been widely employed so far. The reason is that it is very difficult to ensure the metallic contact between the aluminum strip and the copper holder. In order to realize such a condition, sputtering and plating of gold on the aluminum surface were tried after mechanical polishing, chemical etching and chemical substitution. Evaluation was done by measuring the electrical contact resistance between the aluminum specimen and the copper holder at 4.2 K. The gold-plated specimen and the gold-sputtered specimen, both treated in the Bonder-dip solution beforehand,showed the same contact resistivity, 5 n ~/cm ~-, the lowest value ever reported.
INTRODUCTION It is no exaggeration to say that at ultra low temperatures the heat switch has a decisive influence on the experiments, especially on the specific heat measurements.
In this temperature region, a metallic
superconductor with the transition temperature higher than 1 K is usually employed for the heat switch.
In
the superconducting state, the heat conduction is very poor, governed by the lattice heat conduction, because the Cooper pair does not carry entropy.
By applying the magnetic field larger than the critical
field, the heat switch restores the metallic thermal conductivity. Aluminum has the high Debye temperature and has no isotope, the former corresponds to the low lattice thermal conductivity at low temperatures and the latter to the high electronic thermal conductivity also at low temperatures. materials.
Therefore, aluminum is the most promising as a material of the heat switch over the other
However, it has not been widely employed so far from the following reason.
Aluminum is
covered with the hard oxide layer which is difficult to be taken off completely and the cleaned surface is easily re-oxidized if the surface is exposed to the air even for a short time.
This oxide layer on the
aluminum surface makes the thermal contact resistance very large. In order to overcome this problem, we tried gold-sputtering and gold-plating On the aluminum surface after removing the oxide layer by using various procedures. 621
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EXPERIMENTAL PROCEDURES The aluminum specimen of high purity (6N) had a shape, 0.1 mm thick, 10 mm wide and 20 mm long,
and its both ends were
gold-sputtered or gold-plated in 5 mm length on both surfaces. For evaluation of the fabricated specimen, it was mounted on the gold-plated copper holder as shown in Figure 1.
The electrical
contact resistances at both ends including the specimen resistance were measured at 4.2 K With four terminal method by increasing the current from - 10 A to + 10 A.
In order to
determine the specimen resistance, two
Figure ! Deviccto measure the electrical contact resistivity.
voltage leads were attached on the surface of the specimen with silver paste. This value was - 5 n Q for all specimens. I. Gold-sputtering after simple chemical etching The gold-sputtering method was adopted at first from the following reasons" (1) Special techniques are not necessary. (2) The surface of the specimen has a possibility to be cleaned by anti-sputtering before gold-sputtering. (3) The specimen is annealed in the course of sputtering. Before gold-sputtering,
the following chemical treatments were performed in the cell vibrated
supersonically and in the nitrogen atmosphere" Wash in the acetone ~ Dip in the 50% HNO, solution.
Dip in the 50% NaOH solution
After anti-sputtering, gold was sputtered on the aluminum surface at 5 kV
target voltage and 0.5 mAJcm 2 target current density in the 0.5 Pa Ar (6N5) atmosphere flowing at a rate of 10 atm 9cc/min. Although many trials were done by changing the time of chemical etching, that of anti-sputtering and that of gold.sputtering, the contact resistivity was only 10.0/z Q/cm 2 at best. We thought that chemical etching was too weak for taking off the hard oxide layer on the aluminum surface. 2. Gold-plating after mechanical polishing and chemical etching In order to remove the oxide layer on the aluminum surface more effectively, the No. 1 specimen was fabricated with the same procedure as Mueller et al.lll as shown in Table 1.
In this case, gold was
electro-plated just as they did to compare the both results. The contact resistivity of this specimen was reduced to about a half of the previous value but still very large, and the plated gold could be easily removed by the soft touch of a finger. The preceding experiments showed difficulty of taking off the strong oxide layer with chemical treatments, we decided to polish the aluminum surface at first with #3000 sandpaper coated firmly with 10/z m aluminum oxide powder. treatment.
The contact resistivity of the No.2 specimen was reduced one order with this
For the No.3 specimen, the acid solution was changed to stronger one.
on the No.3 specimen was not removed off so easily.
The plated gold layer
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Table 1 Specimen
Treatmentprocedure
2
';Contact resistivity at 4.2K
O. 6 /1. f ) / c m 2
:(~)--+(~)--+(~---+@--+(~,- - ~ ( ~ + @ ~ @ ~ @
3
0.5
u n / cm 2
(~ Polish the aluminum surface with #3000 sandpaper. (~) Wash in the acetone. (~ Wash in the alkaline cleanerat 75 0(2 for 60 see (22 ~ fl of Na3PO 4 912H20 and 22 ~ 12 of Na2CO3). @ Dip in the acid bath for 15 see (50% HNO3). @ ' Dip in the stronger acid bath for 15 see (equal volume mixture of HF, HNO 3 and water). (~) Dip in the zincate solution at (22-t-2) ~ for 60 see ( 1 g/~ of FeC13 96H20, 100 g/12 of ZnO, 525 g/12 of NaOH, l0 ~ 12of C4H4KNaO o 94HeO). (~ Dip in the same acid bath as @ for 30 sec. (~' Dip in the same acid bath as @ ' for 30 see. (~ Dip in the same zincate solution as (~ at (22-1- 2) ~
for 10 sec.
(~ Copper strike; at 26 mA/cm 2 for 2 min and then 13 mA/cm 2 for 2 min using copper anode in the following so|ution; (41.3 g/t2 of CuCN, 50.8 ~ 12of NaCN, 30 ~ 12of Na.CO~, 60 g/12 of C4H4KNaO,- 4HeO). (~) Gold plate; deposit 1/x m at 0.5 mA/cm 2 for 4 min. 3. Gold-plating after treatment in the Bondar-dip solution Today, effective zincate solutions are sold for industrial use.
So, we decided to employ the Bondar-dip
solution [2] on the market in place of the zincate solution shown in Table 1.
Accordingly, the processes
(~ and (~ were replaced by the process of dipping in the Bonder-dip solution ((~)). The results are shown in Table 2. The No.6 and No.7 specimens give the best result, 5 n ~ / c m 2, the smallest contact resistivity ever reported [3].
This fact may indicate that the oxide layer on the aluminum
surface was completely removed. Table 2 Specimen
Treatment pr(x:edure
!
5
.
.
.
.
.
|174
!
2
@--+(~)--+(~)--+(~)'--+(j~(~'7~~~
--,|
6 .
Number of peffonrtance of Contact resistivity' the substitution process at 4.2K
.
.
.
7
.
.
.
.
'
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
.
2
--, -+|174 .
.
/
t / / -,b. . . . . . . . . . . . . . . . . . . . . . .
(~--+(~)--+(~)--+@'~(~)--,(~~~~
O. O15/.1. f ) / c m 2
3 !
'
0 . 0 0 5 u o/
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4
0
.
m
~. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
005 # f ) / c m 2
@ Dip in the Bondar-dip solution Comparing Tables 1 and 2, it can be said that the Bondar-dip solution forms a better substitute layer than the zincate solution. substrate.
In the case of the zincate solution, the substitute layer tends to form a thick and porous
While, in the case of the Bondar-dip solution, the layer precipitated in the early stage of the
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substitution contains relatively large amount of copper.
This layer covers the aluminum surface uniformly
to make strong bonding with aluminum and serves to suppress rapid growing of the substitute layer to form a thin and dense substrate. 4. Gold-sputtering after treatment in the Bondar-dip solution At the end, gold-sputtering was performed again. The chemical procedure before gold-sputtering was the same as shown in Tab|e 2. The results are indicated in Table 3. In the case of the No.8 specimen, the substitute layer was dissolved out by the acid solution just before sputtering. But, for the No.9 specimen, gold was sputtered on the substitute layer.
From comparison of
the contact resistivity of these specimens, it can be inferred that the oxidized layer formed during the setting to the sputtering apparatus could not be removed by anti-sputterino~ . The No. 10 specimen has the smallest contact resistivity, 5 n Q / c m 2, just the same as the No.6 and No.7 specimens. Table 3 Spec~en
Treatment procedure
Anti-sputtering time Sputteringtime
Contactresistivity at 4.2K
8
5 min
5 min
0.095 # f'//cm 2
9
5 min
5 min
0.025 # f ) / c m e
5 min
5 min
0.005 # f ) / c m 2
CONCLUSION When aluminum is used for a superconducting heat switch, it is important to reduce the thermal contact resistance between the aluminum strip and the copper holder by removing the oxide layer on the aluminum surface.
We have developed the best procedure to fabricate an aluminum heat switch by applying
gold-sputtering or gold-plating on the aluminum surface after mechanical polishing, chemical etching and chemical substitution with the Bondar-dip solution. Both gold-sputtering and gold-plating gave the similar result, 5 nQ/cm e in electricalcontactresistivity, which is the lowest value ever reported. REFERENCES 1
Mueller,R.M.,Buchal,C.,Oversluizen,T. and Pobell,F., Superconducting aluminum heat switch and plated presscontacts for use at ultralow temperatures, Rev.Sci.lnstrum. (1978) 49 No.4 515-518
2
Bondar-dip solution, CANNING JAPAN Co. Ltd., Kameido I-8-4, KBt~-ku, Tokyo 136, Japan
3
Bunkov,Yu.M., Superconducting aluminum heat switch prepared by diffusion welding, Cryogenics (1989) 29 September938-939
The Study on the Solid Thermal Contact Resistance at Low Temperatures
Xu Lie*,
Zhou Shuliang*,
Yang Jun*,
Xu Jiamei**
*Low Temperature Center, Shanghai Jiao Tong University, Shanghai 200030, China **Shanghai Sunny Research Institute of Environment and Energy, Shanghai 200040, China
This paper mainly studies thermal contact resistance (TCR) of solid interfaces. On the basis of Reedwood-Williamson model, a thermal contact resistance model is proposed. Experiments and analyses of TCR of stainless steel and aluminum have been done which include the effects of pressure, temperature, roughness, stuffing and different materials.
INTRODUCTION In many advanced research and application areas of high technology, it is necessary to keep the measuring equipment or a system in a low temperature environment. Heat has to be taken from it through a passage connected to a cooler or low temperature liquid. The contact resistance of the solid surface will affect the heat transfer performance of the system. Lower values of TCR can improve the performance of heat transfer, while higher values can be used for insulation[ 1]. In the experiments, effects on TCR of pressure, temperature, surface roughness and material properties have been studied, some of the results are presented in this paper which may be useful for the engineering design.
EFFECT OF SURFACE ROUGHNESS When heat is transferred from a solid surface to another through an interface, usually there exits an excess temperature drop due to the interface A T - T~ - ~
(1)
Defining R - q / A T as contact resistance, which is composed of thermal resistance Rs of solids, thermal convection resistance Rf of fluids between the contacting surfaces and thermal radiation resistance R,. of the two surfaces:
R
-
R.,. + Rj~ + R,.
(2)
If the heat transfer occurs in vacuum, the fluid heat transfer can be treated as conduction heat transfer through fluid, the thermal contact resistance is R f - Kt-Aj./6, where 6 is the average thickness of the gap between two surfaces. Compared to solid surface heat conduction and fluid heat transfer, the radiation effect is very small and can be neglected. Actually, the percentage of solid surface thermal resistance or fluid thermal resistance to the total resistance is changing due to the effects of thermal conductivity of the solid and fluid, and also the actual contact areas. The fluid (mainly gas) is under vacuum, its heat transfer effect can be calculated by the formula of molecular flow[2]. So in this paper the effect of solid surface heat conduction will mainly be analyzed. The actual contact area is only part of the total area, so the heat flux will have a constriction at the joint. Under the conditions of constant material properties and no heat source, just like fluid flowing through a hole must obey the equation of the conservation of mass, the equation of conservation of energy is: 625
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+
+
-0
(3)
here z is the axial distance from contact. There are many small channels on the contact surface, the above equation for one channel model can be written using cylindrical coordinates as: 02 T
1 (7/'
~'~
r ~"
C 2T
+
-0
-~-
(4)
according to the boundary condition, when Z --~ oe
; r - c,, + 9
/
z
(s)
when there is no insulating gap between the interface, 71:=. - C., actually, the average temperature at the interfaceis
:
~ ,-, - C,., + q -=
x ~(a')
4ak
(6)
from Equation (5) and (6), the temperature drop due to the constriction of heat flux on the contact is: A T - TI_=,:' -7~:=,:,--
4 aqk
x ~(s)
(7)
~J is the function of shape and size of the contact area. From the analyses above, it can be seen that thermal contact resistance of solid interface is mainly caused by the constriction of heat flux . Practically, there are many small areas connected together on a interface, their sizes and distribution are very complicated. For this kind of surfaces, we give the formulae below on the basis of analyses[3 ].
R-1
or,.' 3D. K
•
E
F
•
V(,5.)
~,_(d/cr. '
. .
h - C(F). D ~ - -
o- ["'"
(8)
(9)
Formulae above indicate that TCR is in inverse relation with the heat conductivity of materials and the density of rough granule on the surface, proportional with surface roughness and the modulus of elasticity of the materials.
RESULTS AND ANALYSES OF THE EXPERIMENTS Experimental results of TCR of aluminum and stainless steel with various conditions are shown in Figure 1 to Figure 5. Effect of Loads Figure 1 shows the thermal conductance data of aluminum pairs (No.1 to No.5) at 135k with different surface conditions under loads from 1MPa to 3MPa. They have the same tendency vs. pressure though the values are not same.
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627
Effect of Temperature Figure 2 is the thermal conductance of stainless steel pair No. 1. This figure shows that thermal conductance increases with increase in interface temperature. Published data of thermal conductivity of bulk material has been used to calculate the heat flux, therefore, the effect of temperature on interface thermal conductance is influenced partly by bulk material thermal conductivity. This data has a big difference with reference [4]. In reference [4], when the average temperature of the contact was changed from 30~:C to 190~C, thermal conductance has changed four times. But we got only 25% change when temperature changed from 130K to 350K. The discrepancy can be due to the differences in the experimental conditions. The data in reference [4] was obtained at the ambient pressure with foam type insulation materials, so thermal conductance of the gas between the interface and also the insulating material would affect the results, while in the present work data was taken under the conditions of high vacuum and with the protection of the radiation shield at lower temperatures.
Figure 1 Thermal conductance versus pressure of A1-A1 interface with different surface conditions (No. 1-No.5) at 135K
Figure 2 Thermal conductance of stainless steel interface at different temperatures (P 1- 1.06MPa, P2 = 1.3 6MPa, P3 = 1.67 MPa, P4=2.09MPa, P 5=2.86MPa)
Figure 3 Thermal contact resistance of stainless steel with different surface roughness at 350K(P 1= 1.06MPa, P2 = 1.3 6MPa,P3 = 1. 67MPa, P4=2.09MPa,P 5=2.86MPa)
Figure 4 Thermal contact resistance at different temperatures under ambient pressure
Effect of Surface Roughness Figure 3 shows thermal contact resistance of stainless steel samples versus surface roughness at 350K. At certain temperature and pressure, thermal contact resistance will increase with the increasing of the
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roughness. At 2.86MPa, when the roughness was reduced from 44/z m to 18 ~z m, thermal contact resistant lowed by 40%, while the roughness changed from 18 g m to 2 tz m, thermal contact resistance decreased by 80%. Effect of Stuffing in Gap The value of thermal conductance of stainless steel interface under 1.06MPa augmented with vacuum grease has been measured as 2473W/mZ'K, which is 7 times better than the data without the augmentation of vacuum grease. The results shows that augmentation of vacuum grease can reduce the TCR effectively. Comparison Between Theoretical Analyses and Experiments The experimental data and the references imply that thermal conductance has a linear correlation with pressure. To a certain sample, at a certain temperature Equation (9) can be rewritten as h=c-F. So theoretically, when the pressure of contact is 0, thermal conductance will also be 0, but we find that there exists a residual thermal conductance. As shown in Figure 5 , thermal conductance h=h0+c.F, this is caused by some other heat transfer mechanism not included in the analyses, and the presumption of fully elastic deformation. So it is better to include the residual thermal conductance to calculate thermal conductance. Fitted formula for thermal conductance of stainless steel shown in Figure 5 is: h=221+267P (350K) Figure 5 Relation between TCR and pressure of stainless steel at 3 50K CONCLUSIONS Theoretical analyses agree with the experimental results qualitatively. Pressure, temperature, surface roughness, material properties and filling the gap will affect thermal conductance as follows 1. the pressure applied on the contact surfaces affect the thermal conductance, and it can be expressed as h=h0+c.F. 2. thermal conductance increases with the rising of the temperature. 3. surface roughness has a large effect on thermal conductance, thermal conductance will increase with reducing of surface roughness, especially at the low roughness condition. the properties of the bulk material will affect thermal conductance of the interface. Filling heat conducting medium into the gap can increase thermal conductance. So it is possible to increase or decrease thermal conductance by changing pressure, roughness, filling the heat conducting grease, and plating with soft heat conduct film, etc. ~
REFERENCES 1. 2. 3. 4.
Snaith, B., Probert, S.D., Callaghan, P.W., Thermal resistance of pressed contacts, Applied Energy (1988)20 31-84. Xu, L., Fang, R.S., Ma, Q.F., Insulating Technology, National Defense Industry Press, China, (1990) 210-219. Yang, J., Research on the Heat Contact Resistance at Low Temperature, thesis for Master's degree, Shanghai Jiao Tong University, China (1995). Gu, W.L., Experimental research of heat contact resistance, Journal of Nanjing Aviation Institute (1992)24 44-53.
Experimental Study on Thermal Contact Conductance at Liquid Helium Temperature Kazumi Sunada,
Yoon-Myung Kang
MEC Laboratory, DAIKIN INDUSTRIES, LTD., 3 Miyuldgaoka, Tsukuba 305, Japan When cooling an object in a cryogen-free, cryocooled system at liquid helium temperature, it is important to quantitatively estimate the thermal resistance that occurs at the contact surface between the cooling stage and the object. Here we employed two methods to estimate this thermal contact conductance at liquid helium temperature : (1) an empirical equation and (2) the Wiedemann-Franz law. We compared the estimated conductance with that measured at liquid helium temperature, both methods were proved valid. Additionally, we measured the thermal conductivity and the electric resistance of phosphor bronze screen stacks at room temperature and at low temperature.
INTRODUCTION When cooling an object in a cryogen-free, cryoeooled system at liquid helium temperature, it is important to quantitatively estimate the thermal resistance that occurs at the contact surface between the cooling stage and the object. A recent comparison shows a large discrepancy in the published values for the thermal contact conductance of nominally similar contacts at room temperature; sometimes differing by more than six orders of magnitude. Furthermore, little data is available on thermal contact conductance measured at liquid helium temperature of bare contacts. Therefore, there is a neeA for bare contacts. Here, we measured the thermal contact conductance of bare contacts at liquid helium temperature. First, as reference, we measured the thermal contact conductance of bare contacts made of copper (OFHC) at liquid helium temperature. We then compared these measurements with estimates calculated using an empirical equation derived by P.W. O'Callaghan and his colleagues [1] for the thermal contact conductance at room temperature under a vacuum in which they assumed elastic contacts. Second, we estimated the thermal conductance of contacts with and without interposers at liquid helium temperature by measuring the electrical resistance of the contacts at room temperature. For electrically conductive contacts consisting of conductive blocks and interposers such as indium foil, we estimated the thermal conductance at liquid helium temperature (CT4K) from the Wiedemann-Franz law, taking into account the temperature dependency of the contact's properties. For reference, we measured the thermal conductance of contacts with a non-conductive thermal grease (Apiezon-N). We then compared these estimates with our measurements, showing that the thermal contact conductance of electrically conductive contacts made of copper at liquid helium temperature could be estimated from the electrical resistance at room temperature within an accuracy range of about 30%. Additionally, we measured the electrical resistance of phosphor bronze screen stacks, which are used in regenerators of cryocoolers, at room temperature and low temperature. We used the measured thermal conductance of the stacks at low temperature to calculate the thermal conductivity of the stacks. Because of the result of this, the Wiedemann-Franz law did not hold between the electrical resistance of the stacks and the thermal conductivity of them. EXPERIMENTAL APPARATUS We chose copper (OFHC) as our sample because k is often used in cryoeoolers, and because the thermal contact conductance of copper has been measured by many researchers at room temperature. The contact specimen was a copper (OFHC) cylinder with a nominal cross-section of 1.4 cm 2, an outer diameter of 14 mm, and an inner diameter of 5.0 mm. To normalize the data as the following equation (1), we measured the surface roughness before and after each experiment. A load was applied to the contact specimen by a stainless steel bolt at pressured from 0 Mpa to 20 MPa and was measured by strain gages. A steady-state longitudinal heat flux method was used to measure the thermal contact conductance. To occur heat flow, a button heater was placed on the upper samples. To monitor the specimen temperature, Ge thermometers were installed in both the upper and lower samples such that the thermometer axes were parallel to and located 5 mm from the contact interface. 629
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The electrical resistance of the contact was measured using a 4-wire DC method. The wires were soldered to both the upper and lower samples. To keep the sample heating negligible, current was controlled by the thermometer response. We thus measured the thermal contact conductance of copper (OFHC) without a thermal interposer (i.e., bare) with indium foil, and with Apiezon-N grease. For the indium foil, we placed a thin donut of indium foil between the contact interfaces, in which the outer diameter of the donut was 14 mm, the inner diameter was 5.0 mm, and the thickness 0.1 mm. For the Apiezon-N, we applied the grease to both contact surfaces, provided the contact was electrically conductive. EMPIRICAL EQUATION FOR THERMAL CONDUCTANCE AT ROOM TEMPERATURE We normalized the load and the thermal contact conductance with surface roughness etc. using the results from following equation by P. W. O'CaUaghan et al. [ 1] as reference.
W*= W
a2H
C*=-.A~ aX
(1)
where W* is the normalized load applying the contact, C* is the normalized thermal contact conductance, W is the load, C is the thermal contact conductance, 6 is the surface roughness, H is the effective elastic modules, An is the nominal cross-sectional area, 2Lis the thermal conductivity, of the sample. By assuming the contacts were elastic contacts, O'CaUaghan et al. derived an empirical equation (2) from previous experimental data (344 points) that included data for contact between various metals ( e.g., copper-copper, aluminum-aluminum, etc.) in a vacuum at room temperature
Figure 1
Comparisons of normalized thermal contact conductance
Hgure 2 Compafis.ons.gf t.he.rma.1conductances measured at llqU1(1 llellum temperature with and without Indium foil or APIEZON-N and those calculated from electric resistancesat room temperature
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631
(2)
Figure 1 shows the measured and empirical thermal contact conductance for copper (OFHC) contacts without interposers at low temperature. [2,3] The agreement between the two confh'rns that the Callaghan empirical equation can accurately explain this thermal contact conductance. EFFECT OF INTERPOSERS AND ELECTRICAL RESISTANCE OF CONTACT The Lorentz Number of most metals depends on the temperature. [4, 5] For copper (OFHC), the Wiedemann-Franz law holds near liquid helium temperature and room temperature. Using this law, Lee. et al. [4] explained the relationship between the thermal contact conductance (CT) and the electrical contact resistance (Rel) as CT=
LT AnRel
(3)
where L is the Lorentz Number and T is temperature. Using copper (OFHC) as the contact specimen, we measured the thermal contact conductance at liquid helium temperature. And the electrical contact resistance was measured at room temperature and liquid helium temperature. The relationship between the thermal contact conductance at liquid helium temperature: CT4K and the electrical contact resistance at liquid helium temperature : ReMK can be explained by the equation (3). Therefore, if we can derive Rd4K, we can calculate CT4K. When RelnK was too small to measure, we extrapolated the value by assuming the pressure that was decreased by cooling. Figure 2 shows the comparison of CTnK calculated using equation (3) for the electrical contact resistance at room temperature with that measured at liquid helium temperature. These two values agree well. MEASUREMENT OF PHOSPHOR BRONZE SCREEN STACKS Additionally, we measured the thermal conductivity of phosphor bronze screen stacks at low temperature and the electrical resistance of them from 4K to 300K. These stacks are used in the regenerators of cryocoolers. Each screen was 25.4 mm in diameter and was a 200-mesh composed of 53 Ixm diameter wire. We measured stacks of 40 and 100 screens. After measuring the thermal conductance of a stack at low temperature, we calculated its thermal conductivity in bulk shown in Figure 3. And condition 1 is a stack that was cleaned ultrasonically before measurement. Condition 2 is a stack that was cleaned ultrasonically before measurement and then degassed 8 hours at 393K. Condition 3 is a stack that was cleaned ultrasonically before measurement and then oxidized 8 hours at 393K. Figure 4 shows the Lorentz Number of the stacks calculated from the thermal conductivity of the stacks and the electrical resistance of the stacks using Wiedemann-Franz law in equation (3). -At,xlO0
W-2-d
[%]
(4)
where d is wire diameter of phosphor bronze consisting screen mesh, At is thickness of a clamped screen, and W is the degree of clamping of the screens. The pressure applied the contact at low temperature was measured from 0.01 to 0.4 MPa by the strain gage. The Lorentz Number of phosphor bronze bulk was about 2.9 * 10* W f g K 2 at 300K. In contrast, that of the phosphor bronze screen stack was large more than 2 orders of magnitude compared to 2.9 * 10.8 WDJK 2, provided the Lorentz Number of the stack decreased as the pressure was increased. For tI' was 104 %, the Lorentz Number of the stack for condition 1 was around 70 * 10"8 WD/K 2 at liquid helium temperature. Further work is needed to determine the causes for the increase in the Lorentz Number of the stacks, namely, measurements in which (a) higher contact pressure is applied and (b) the contact surface is cleaned chemically. The result for condition 3 indicates that an oxidized contact surface strongly influences the Lorentz Number; this influence agrees with that reported in a previous study by Nilles, et al. [2] Therefore, oxidation of the contact surface was a factor in increasing the Lorentz Number of the stacks.
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Figure 3 Comparisons of thermal conductivity for phosphor bronze stacks
Figure 4 Lorentz Number of phosphor bronze stacks
CONCLUSION The equation (2) also helps us to quantitatively understand the thermal conductance of bare contacts at liquid helium temperature. For electrically conductive contacts consisting of electrically conductive blocks and interposers such as indium foil, the thermal conductance at liquid helium temperature could be estimated with WiedemannFranz law when the temperature dependency of the properties are considered. In contrast, the thermal conductance of phosphor bronze stacks whose surface was oxidized could not be estirnated with Wiedemann-Franz law between 4K and 300K. REFERENCES 1 2 3 4 5 6
O'CaUaghan, P.W. and Probert, S. D. 9Journal Mechanical Engineering Scienc~ (1974) 16 No.7. 4155 Nilles, M. J. and Van Sciver, S.W. 9.Adv. Cryogenic Engineering (1988) 34 443-450 Yu. J., Yee, A.L. and Schwall, R.E. 9Cryogenics (1992) 32 610-615 Lee, A.C., Ravikumar K.V. and Frederking T.H.K. 9Cryogenics (1994) 34 451-456 Clark, A.F., Childs, G. E. and Wallace, G.H. 9Cryogenics (1970) 4 295-305 Li, R. Hashimoto, T. Ohta, K. Okamoto, H. 9Proceedings International Cryogenic Engineering Conference (1988) 12 414-417
Cryogenic engineering
Gas properties
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On the Joule-Thomson Integral Inversion Curves of Helium-3, Helium-4 and Hydrogen
B-Z. Maytal* and A. Shavit** *Rafael Institute, Cryogenic Section, P.O.Box 2250(39), Haifa 31021, Israel **Department of Mechanical Engineering, Technion-Israel Institute of Technology, Haifa 32000, Israel The Integral Inversion Curve (I.I.C.) is the contour of all thermodynamic states which exhibit a zero integral cooling under Joule-Thomson (isenthalpic) expansion from a pressure down to ambient atmospheric pressure. This is an alternative and complementary presentation of the Joule-Thomson inversion phenomena. The traditional inversion curve is a differential one. The I.I.C. has a peaking temperature dependence in the plane of pressure and temperature. This is the highest pressure for which integral cooling is still possible. However, most of the gases solidify before the peaking pressure is reached. The quantum gases, helium and hydrogen comprise the exceptional group which enables the study and verification of the predicted peaking pattern of the I.I.C. The helium-4 and hydrogen I.I.C.'s are obtained through an available numerical code. The I.I.C. of helium-3 is evaluated through an advanced, sixteen parameters equation of state. By analogy to the traditional maximum inversion curve, the maximum integral inversion reduced pressures are determined to be about 40 for helium-4 and 30 for both hydrogen and helium-3.
NOMENCLATURE Cp
Cpo
h hK; hR Ah r M
n~ P
Isobaric specific heat, J/(mole K) Isobaric specific heat at zero pressure, J/(mole K) Specific enthalpy, J/mole
•Pc
R
T
Ideal gas enthalpy, J/mole
BOIL
Residual enthalpy, h - h tG , J/mole Integral isothermal Joule-
V
Thomson effect, J/mole Molar mass, g/mole Coefficients of the He 3 equation of state, i- 1... 16, Pressure, Pa
Z Greek 9 | FI
Tc Vc
635
Critical pressure, Pa Universal constant of gases, 8.314 J/(mole K) Isenthalpic temperature drop, K Absolute temperature, K Normal boiling point, K Critical temperature, K Specific volume, m3/mole Critical specific volume, m3/mole Compressibility, P . v / ( R . T ) notation Density, mole/dm 3 Reduced temperature, T/Tc, Reduced pressure, PIPe,
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INTRODUCTION The inversion of the Joule-Thomson effect was extensively and continually studied in its differential form for about ten decades. The essence of this remarkable tradition manifests itself in the differential inversion curve. It is the locus of all thermodynamic states, (P, T),
for which the Joule-Thomson coefficient
vanishes, namely, ~t = (aT/OP)h = 0. These states serve as a border of transition from the differential cooling to the heating zone. An alternative approach of treating the inversion phenomena is the integral one. The posed question wheather integral heating (ATh > 0)or cooling (ATh < 0) is examined under finite pressure drops, rather then infenitisimal from P down to zero (or ambient). The integral inversion states are
- r ( P = 0, h ) - v(P, h ) - 0
(1)
In historical perspective, the very first scientists who explored the inversion phenomena adapted the integral approach. It was Porter [ 1] (1906) that recommended and established the differential tradition replacing those days current integral approach, not by claiming of any advantage but plainly, for simplicity of treatise of the inversion phenomena. Following Porter, scientists organized and arranged their measurements and analysis in terms of the differential version. Maytal and Van Sciver [2] proposed an empirical correlation for the I.I.C. of low acentricity gases. Maytal and Shavit [3] mapped the intensity of the Integral effect forthese gases and obtained a qualitative picture, including a closed form solution for the I.I.C., applying the Van der Waals equation of state. Maytal and Pfotenhauer [4] derived the I.I.C. via the Peng-Robinson equation of state Koeppe [5] was the first that explicitly introduced the concept of the I.I.C. accompanied by a qualitative plot. He proved that the I.I.C. has a peaking shape and the state of maximum integral inversion curve, if reached, satisfies the condition, Cp = Cpo. However, excluding helium and hydrogen, all gases solidify before their I.I.C. peaking condition takes place. The intention of the present treatise would be to (a) present the I.I.C. of helium-4 and hydrogen, (b) verify the peaking condition, (c) determine the I.I.C. for helium-3, through an advanced equation of state, and (d) display the relation between the traditional differential inversion curves and the I.I.C.s.
THE EQUATION OF THE INTEGRAL INVERSION CURVE The definition of the integral inversion curve through equation 2 is equivalent to the condition
AhT-h(P=O,T)-h(P,T)=O
(2)
The residual enthalpy, h R = h - h z~;, is the deviation of the real from the ideal gas enthalpy. Since h'~J(P = O, T) = h'~;(P, T), and h R ~ 0 because of the low pressure, we get Ahy = - h l~. Hence the alternative condition to be fulfilled by the I.I.C. would be, h R = 0. Expliciting [6] h R and setting to zero, finaly leads to the the equation of the I.I.C.,
o1_
o'9
~,-
T
(3)
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DETERMINATION OF HELIUM-3 INTEGRAL INVERSION CURVE Helium-3 is the most outstanding gas within the unique group of quantum gases. It is the strongest violator of the Principle of the Corresponding States. As such, any display of the I.I.C.s would not be complete without helium-3. Driven by the same motivation [7,8,9] the differential inversion curve of helium-3 was evaluated. The hereby applied advanced equation of state [ 11 ], where X = T + 5.6906, P
/
= Z = 1+ n~ + ~ +
x
+
9 +
7
T . n 5 "4-El6 "["
" ~ + ~
v
, 3.9 4 r
"[-
T
(4)
Integrating equation 3 following the substitution of above Z and 0, the I.I.C equation, in the plane (T, p)is obtained. One may solve it in conjunction with equation 4, thus getting the I.I.C. in the commonly used (T, P) plane..
QUANTUM GASES INTEGRAL INVERSION CURVES Information about integral inversion points of helium-4 and hydrogen is directly obtained from GASPAK numerical code [12] of their specific thermophysical properties. For each temperatue, a high pressure state is requested so that enthalpy equalizes the value at zero pressure. Collection of these states comprise the I.I.C . . . . . . . . 1 displays the I.I.C.s for helium-4 and hydrogen, in addition to helium-3 as previously obtained. The differential inversion curve of helium-3 was derived from the same equation of state [9] and supported by direct measurements [10,11] of inversion points. The differential inversion curves are extracted while applying the numerical code, GASPAK of Cryodata Inc. and searching for states of = 0. These, too, are shown in Figure 1. Inversion curves of nitrogen, as a representative of low acenticity gases, are displayed for refemce and amphasis. For each gas the space between the differential and integral inversion curves are marked for better perception of the general picture.
DISCCUSION The pattern of the I.I.C.s demonstates once again the outstanding position of the quantum gases and disobidience of the Principle of Corresponding States. The most remarkable feature is the pressure peaking form of the quantum gases I.I.C. For instance, He 4, while expanding from a higher pressure than 9.51 Mpa (FI = 41.8) down to zero, will allways warm up and never cool down. Lets consider the ratio between the peaking integral and differential inversion pressures for He 3, He 4 and H2 which respectively are, 1.26, 2.45 and 3.20. One may clearly observe a general trend: the I.I.C. approaches closer to the differential inversion curve as the quantum effect of the gas becomes stronger. The two inversion curves of He 3 are indeed relatively close. The helium isotops inversion curves (both differential and integral) are firmly distinguished altough they represent the same chemical element. The gap between helium-4 and hydrogen is similar to the gap between helium-4 amd helium-3.
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REFERENCES ~
,
o
~
~
6. 7. 8. .
10. 11. 12.
Porter, A.W., "On the Inversion Points for a Fluid Passing Through a Porous Plug and Their Use in the Testing Proposed Equations of State", Phil. Mag., Vol. 6, pp. 554, (1906). Maytal, B-Z. and Van Sciver, S.W., "Charecterization of Coolants for Joule-Thomson cryocoolers", Proc. 6th Int. Cryocoolers Conf., Vol. 1, pp. 245, Plymouth, MA, Jan., 1991. Maytal, B-Z. and Shavit, A., "On the Integral Joule-Thomson Effect", Cry_ogenics, Vol. 34, No.I, pp.19, (1994). Maytal, B-Z. and Pfotenhauer, J.M., "The Integral JT Inversion Curve by the Peng-Robinson Equation of State", Proc. 10th Intersociety Cry.ogenic Sym., Houston, TA, (March 1995). Koeppe, W., "Bemerkungen zur Inversionskurve", Ka!tetechnik, Vol. 14, pp. 399-403, (1962). Walas, S.M., Phase Equilibrium in Chemical Engineering, Butterworth Press, Boston, (1985). Maytal, B-Z., "Helium-3 Joule-Thomson Inversion Curve", Cryogenics, (in press). Vortmeyer, D., "The Joule-Thomson Coefficient of Non Polar Gas Mixtures at P ~ 0. A Theoretical Interpretation of Experiments", Kaltetechnik, Vol. 18, No. 10, pp. 383, (1966). Gibbons, R.M. and McKinley, C., "Preliminary Thermodynamic Properties of Helium-3 between 1 K and 100 K", Adv. Cryo. Eng., Vol. 13, pp. 375, Plenum Press, New York, (1968). Duant, J.D., "Preliminary Thermodynamic Data for the Inversion Curve of Helium-3", Cry_ogenic$, Vol. 10, pp. 473-475, (Dec., 1970). Kraus, J. et al., "Enthalpy-Pressure (H-P) Diagram of He3 in the Range 1.0 K
Figure 1"
Integral and differential inversion curves of the q u a n t u m gases" H e 3 , Space between differential and integral curves is shaded.
H E.
He 4
and
A Modified Patel-Teja Equation of State for Cryogenic Fluids +
Guangming Chen, Zhizhong Yin and Guobang Chen Cryogenics Laboratory, Zhejiang University, Hangzhou 310027, P.R. China
This paper introduces a modified Patel-Teja type equation of state which can accurately predict not only for vapor pressures, specific volumes of cryogenic fluids but also for vapor-liquid equilibria of thier mixtures. Deviations between experimental data and the data predicted by the proposed equation are, in most cases, much smaller than those by the Peng-Robinson equation, the original Patel-Teja equation as well as the sottware of PROMIX developed by NIST and Cryodata.
INTRODUCTION Since van der Waals proposed the first real gas equation of state (noted as VDW) in 1873, a number of equations of state have been developed. They have ranged in complexity from simple expressions containing two or three parameters t14~ to more than fifty ones tSl. Though the multi-parameter equations have higher accuracy, they are not generally preferred for phase equilibrium calculations, due to requiting excessive computer time and difficult to obtain generalized forms of the equations for mixtures. The Redlich-Kwong-Soave equation(RKS) t~, the original Peng-Robinson equation (PRI) t2j and the modified Peng-Robinson equation (PR2) TM ,which are based on the VDW equation, have been utilized for precise representation of general properties ofnonpolar fluids, slight polar fluids and some polar fluids. However they have higher errors in calculating saturated properties of fluids. The PatelTeja equation (PT1)t4~ proposed by Patel and Teja on the basis of the RKS and PR equation in 1982 can be used to predict precise saturated properties of general fluids, but it has large error for highly polar fluids and cryogenic fluids, especially for helium. The authors propose a modified Patel-Teja equation of state(PT2) which can be used for cryogenic fluids and their mixtures. NEW EQUATION OF STATE The Patel-Teja equation has its form as:
p : RT -b
a(T) V(V+b)+c(V-b)
(1)
where R is the universal gas constant, a(T) is a temperature-dependent function, b and c are constants. a ( T ) - k (R2T 2 / P ) a ( T ) (2)
§ This project is supported by National Natural Science Foundation of China. 639
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a ( T ) in PT 1 is given by: ct(T)
- [1 + F ( 1 - 1~''2 )]2
(3)
where, F is a constant and Tr - T / T ~ . Studies [4] shown that a nearly linear relationship exists between In a and (1-T~). This suggests that the following temperature-dependent model may be adopted instead of Eq(3): a(T) - exp[m(1 - Tr ) + n(1 - Tr 1/2 )2 ]
(4)
where m and n are adjustable parameters. The second term of Eq(4) is for improving predictions for highly polar compounds.
CALCULATION RESULTS FOR PURE FLUIDS The new equation was used to calculate vapor pressures, vapor and liquid specific volumes of pure fluids. Calculated results of nine common cryogenic fluids have been compared with experimental data and with those obtained from other equations of state(Table 1). In Table 1, AP, AvV and A v t a r e relative percentage deviations of vapor pressure, vapor and liquid specific volumes respectively. Table 1 Comparison of saturation Component Temperature rans;e(K ) PR1 218.15-273.15 19.8 CO2 C~ 120.00-180.00 0.68 90.00-140 00 0.77 02 Ar 85.00-140 00 0.84 CO 79.72-117.21 0.60 80.00-120.00 0.53 N2 Ne 26.00- 35.00 1.20 I-I2 21.00 - 30.40 3.23 He 3.90- 4.30 5.72
properties t6~ AP PR2 PT1 0.13 20.7 0.53 1.52 0.42 1.47 0.35 1.35 0.37 0.42 0.46 1.21 0.55 1.67 0.21 -0.18 --
AV v
PT2 0.45 0.37 0.38 0.30 1.46 0.33 1.14 0.49 0.94
PRIPR2 29.9 0.98 1.86 1.28 1.150.99 1.91 1.43 6.03 2.85 1.72 1.82 1.21 0.45 5.05 1.59 9.62 1.93
PT1 31.3 1.96 1.17 1.82 2.65 1.54 1.15 ---
Av t
PT2PR1 0.33 5.42 1.02 7.12 0.43 0.54 0.61 8.82 1.85 9.60 0.56 7.16 0.39 14.7 0.42 14.9 0.31 11.3
PR2 2.50 7.31 9.57 8.91 9.55 7.17 14.8 15.7 13.6
PT1 8.33 3.11 2.44 2.51 2.31 2.60 3.81 ---
PT2 1.56 6.69 2.34 2.47 1.89 2.66 1.81 4.37 3.59
We can see from Table 1, the average deviation of saturated vapor pressure from the equation (PT2) is lower than those of PT1 or PR1 equation. Furthermore, the average deviations of saturated vapor and liquid specific volumes from PT2 are much lower than those of PR1, PR2 and PT1 equations. Like other cubic equations of state, this equation can not be used to calculate the properties of heliumlI(i.e. T < T ~).
CALCULATION RESULTS FOR MIXTURES Equation (1) can be used for mixtures if the a, b, c are replaced by the mixture parameters respectively. The following mixing rule was used
am
,bin, Cm,
I C E C 1 6 / I C M C Proceedings a m --ZZriXja~j i j
bo - ZZ i
641
.
(5)
,r;bo j
Cm--ZZx, i j
x./cij
where, %. -
b,:,. =
(1-
b~ +b s 2
(6)
Ci + C i
:I
2
Where k,s is a binary interaction parameter which must be evaluated from experimental data. The optimum value of k,s for each binary pair was obtained by minimizing the average deviations between experimental and calculated results in the bubble point pressure and vapor mole fraction at selected temperatures. The determination of the bubble point pressure and vapor mole fraction must satisfy the following conditions: A
V
A
qk, y, - qklx , A V
(7)
A l
Where r q~,, y, and x~ denote the vapor and liquid fugacity coefficients and vapor and liquid mole fractions of the component i respectively. 1. Calculations of Vapor-Liquid Equilibria(VLE) VLE calculation results for 12 binary, systems containing the common cryogenic fluids are shown in Table 2. In the Table, AP is relative percent deviation of vapor pressure and Ay is 100 times absolute deviation of vapor mole fraction in average. Table 2 shows that, the proposed equation gives lower average deviations of both vapor pressure and vapor mole fraction than those of PT 1 and PR1 equations and the PROMIX sottware. It is worth to point out that even the temperature is as low as 20.4K, the new equation can still be used to VLE calculation for helium-hydrogen binary and lower average deviation in vapor pressure and vapor mole fraction can be obtained while other equations or software memtioned above can not give out satisfactory results for this binary system at so low temperature. Table 2
Comparison 0fbinary VLE results
No.
System
1 2 3 4 5 6 7 8 9 10 11 12
AI'-O2
No. of points
Temperature (K)
5 100.15 Ar-CI-h 5 164.00 Ar-CO2 5 253.15-273.15 CO-At 9 123.4 CO-CH4 9 123.4 N2-Ar 12 79.05-89.15 N2-CO 10 110.15 N2-O2 12 95.15 N2-CO.,. 8 270.0 N2"CI-I4 6 95.0 H2"Nz 5 90.15 He-H2 7 20.40 Average Dexdations
PT2
PT1
AP
Ay
0.53 0.33 4.96 0.42 1.99 0.84 0.31 1.26 0.35 0.30 1.62 1.87 1.23
0.16 1.34 1.79 0.57 0.4t3 0.95 0.35 0.10 0.32 0.23 0.10
1.63 0.75 15.70 0.59 1.51 1.50 2.02 1.00 14.37 1.54 . . .
0.55
4.06
o.30
AP
PR1
PROMIX
Ay
AP
Ay
AP
Ay
Sources
0.28 1.68 1.77 0.40 0.42 1.06 0.64 0.22 0.95 0.17
1.14 0.42 14.18 0.41 1.40 2.01 1.94 0.91 13.98 0.64 4.15
0.25 1.49 1.84 0.53 0.39 1.13 0.74 0.10 0.93 0.18 0.15
1.33 2.00 17.24 0.52 4.06 4.00 6.07 1.96 2.89 3.50 20.55
0.28 1.03 4.94 0.51 0.58 0.79 1.63 0.29 0.85 2.03 6.72
[8] [13] [15] [13] [13] [9] [10] [7] [14] [12] [16]
3.74
0.70
5.83
1.79
.
. . . . . . . . . . . . 0.76
[17]
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2. Prediction of Ternary VLE Data The ternary system Oxygen-Argon-Nitrogen was used to test the extension of the proposed equation to multi-component systems. The interaction parameters were regressed from binary systems VLE data alone. The PROMIX software was used for :omparison purpose. The results are shown in Table3. It is apparent that the extension of the proposed equation to ternary system is satisfactory and the average deviation in saturation vapor pressure is lower than that calculated by PROMIX software. Table 3
Ternary prediction for system O2-Ar-N2 ( 85.25-88.35K ) t~q AP
PROMIX PT2
3.14
0.21
1.31
0.36
0.40 0.35
CONCLUSIONS The proposed cubic equation of state can be used to obtain accurate and consistent predictions of the thermodynamic properties of pure cryogenic fluids and their mixtures. Comparisons have shown that, for pure fluids, the proposed equation gives lower average deviations in vapor pressure and vapor and liquid specific volumes than those of the original PT, PR equations; for mixtures, the new equation gives lower average deviati, ts in vapor pressure and vapor mole fraction than those of original PT, PR equations and PROMIX oflware. By the parameters regressed from binary systems, the equation can be used to predict multi-component systems and gives out a satisfactory result. REFERENCE 1
2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
Soave G.S., Chem. Eng Science, (1972)27, 1197- 1203 Peng D.Y. and D.B.Rob.nson, Ind.Eng.Chem.Fundamen. (1976) 15,59-64 Melhem G. A.,et al, Fluid Phase Equilibrium, (1989)47, 203 Patel,N. C. and Teja, S., Chemical Engineering Science, (1982)37,463 Reynolds W. C., Thermodynamic properties in SI, Stanford University, Stanford CA, 1979 ASHRAE Handbook Fundamentals, (1987) Dodge B. F., Chem. Met. Eng. (1928)35,622 Clark A. M., et al, Proc. Roy. Sot., (1954)221, A, 1147, 517 Dolezalek F. Z., Physik. Chem., (1919)93, 585 Yushkevich N. F., et al, Zh. Khim. P., (1936)13, 1273 (in Russian) Fastovskii V. G., et al, Zh. Fiz. Khim., (1957)31, 836268 (in Russian) Kidnay A. J., Miller R. C., Parrish W. R. and Hiza M. J., Cryogenics, (1975)15,531 Christiansen L. J., et al, Cry_ogeni .cs, (1973) 13,412 Fahad A. Somait and Arthur J. Kidnay, J. Chem. Eng. Data, (1978)23,302 Kaminishi G. I. et al, J. Chem. Eng. of Japan, (1968) 1,109 Maimoni A. AIChE. J., (1961 )7,3 71 Streett N. B. et al, Journal of Chemical Physics, (1964)40, 1390
Helium Extraction from Thermal Spring Gases
Debasis Ghose, Bikash Sinha, R. Dey, S.K. Das and D.G.Bhattaeharya. Variable Energy Cyclotron Centre, Department of Atomic Energy, 1/AF Bidhan Nagar, Calcutta - 700 064, India. Helium concentration in natural gases is highly uneven and anomalous. Known natural gas sources in India are lean in it, making recovery uneconomical. On the other hand, many Indian thermal springs contain a high percentage of helium in their emanations, so that it becomes practicable for extraction. A pilot recovery plant has been set up at Bakreswar, West Bengal, about 250 km NW of Calcutta. The paper presents the purification process adopted and future scope of work in this direction.
INTRODUCTION Helium, the second most abundant element in the universe, is scarce in the earth's atmosphere because of thermal diffusion to outer space. It is far more plentiful within the solid earth due to continuous generation through present day radioactivity, retention of fossil helium lett by short-lived (~, ~107 yrs) extinct radioactivity, and the existence of primordial helium trapped deeper within. The solid earth therefore, automatically provides source of commercial helium. For nearly a century, this has been the helium rich mineral gases trapped within a few favourable geological locations [1]. Because of the pressure of large scale contemporary utilisation and the gradual depletion of known reserves there is a compelling need to look for other and alternate terrestrial channels. High temperature thermal springs often provide a continuous escape route for crustal as well deep-seated mantle helium [2]. In recent years, on an experimental basis, we have set attention to extracting helium from such thermal springs. The springs mentioned are located at Bakreswar, West Bengal, about 250 km NW of Calcutta and Tantloi in the district of Santhal Paragana, Bihar, about 20 km further NW of Bakreswar. PRETREATMENT The recorded composition [3] ofthe spring gases show that helium is present to the extent of about 2% while the remainder is mainly nitrogen, approximately 90%. Table-1 gives an idea of the gas composition by volume. The emanating gases escaping under a positive pressure of about two atmospheres is continuously transferred to gas holders. It is rendered free of moisture and carbondioxide by passing successively through four columns of glycol-amine mixture followed finally through a column of zeolite 13X dessieant [4]. The gas at this stage has a dew point typically -70~ C. The dry feed gas is then processed for helium extraction. Table 1 Percentage composition of spring gas Location Bakreswar Tantloi
He 1.88 1.50
N2 90.0 91.1
Ar 2.62 1.92
CO2 2.0 1.8 643
CI-h 2.7 2.8
02 0.8 0.9
1-12 Trace
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HELIUM EXTRACTION The feed gas is considered basically as a two componem mixture. The principle of separation relies in cooling the warm gas until it is partially liquefied and removing the uneondensed helium rich vapour for further processing. The actual purification mode is in 3 stages as shown schematically in fig. 1. The raw gas is compressed to 5 kg/cm 2. Its thermal duty is reduced by passing through a double pipe heat exchanger HX-1 which is cooled counter flow by LN2 vapour returning from the receiver R. The emerging feed gas temperature is lowered from about 320K to 190K, then enters the inner tube of a second heat exchanger HX2, the upper portion of which is a double robe copper coil while the lower half is tightly wound around the cylindrical SS 304 receiver. The outer tube is filled with pressurised helium which performs the cooling process in a closed system. The feed gas temperature is fimher lowered by forced circulating helium gas at 80K in a closed loop with the help of PW7331 circulating pumps mounted on the cryogenic head of a Philips A-20 cryogenerator. The junction boxes, cold exchangers and the cryohead are vacuum insulated. Nitrogen and other volatiles are condensed and collected in the receiver while the uneondensed enriched He (-~ 66% ) is bled out into a gas holder and further purified by passing through columns of LN2 cooled activated charcoal, at a pressure about 20 kg/cm 2 . There are a number of hot springs in the area with variable rates of emanation. In the two springs under present consideration the relative quantities of escaping raw gas and helium yield are shown in Table 2. Table 2 Relative helium yield Location Bakreswar Tantloi
He 1.88 % 1.50 %
Raw gas/day 5000 Its 40000 Its
He/day 90 Its 450 Its
cylinders/year 5 25
The Bakreswar plant, which is already in operation, is to be the focal point where feed gas is to be brought from nearby springs for helium extraction. Infrastructure for harnessing the T antloi gas are presently under way. FUTURE SCOPE As is obvious, the helium escaping from hydrothermal spring vents constitute only a very minor volume in comparison with helium obtained from petroleum bearing natural gases. But there is a two fold advantage. Firstly, these sources are perennial and seemingly inexhaustible. Secondly, the infrastructure needed for collection and purifying are relatively small, so that one may envisage a larger number of such units distributed over different spring zones. It is now recognised that escape of helium and other volatiles from the earth's interior are related to the displacement of the lithospheric plates especially along subduction zones and in areas where new crust is being formed. It has been speculated that in such areas large scale primordial helium enriched in helium-3 is being released. The overall scenario is also similar in many of the volcanic regions. Indeed the volumes of helium that escape from these areas are far greater and more widespread than hitherto believed. Our own findings are that a significant amount of helium is released, from the earth surface in many areas adjacent to thermal spring sites. A detailed study in this direction is currently in progress. As current accumulations decline in conventional reserves there will be a compulsion to seek alternate terrestrial sources, especially in regions where the availability is low. One can therefore foresee, that in the not too distant future, helium from such hydrothermal sources will likely be an alternative, but never in the scales of present-day exploitation of the commodity.
ICEC16/ICMC Proceedings REFERENCES State and Society., Congress considers new helium conservation plan Physics Today (1979) November 91-92 Gupta S.K. and Sharma P., Helium probe for the earth's secrets Science Today (1980) April 17-21 Ghose Debasis and Chatterjee S.D., Genesis of the abundance of helium formation in natural gas emanating from thermal springs Proc.Indian natn. Sci.Acad. (1979) 46A 81-83 Ghose Debasis, Das Nishith Kr., Das S.K. and Bhattacharya D.G., Helium separation process from nitrogen rich natural gas Indian J Cryogenics (1993) 18 190 - 193
Fig 1
Flow sheet of purificotion process
Fig 2
Cold box with heat exchanger HX 2
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Cryogenic engineering
Measurements
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Temperature Measurement under High Magnetic Fields around 1.8 K by using CGR Thermometers Tomiyoshi Haruyama, Nobuhiro Kimura, Ken-ichi Tanaka and Akira Yamamoto KEK, National Laboratory for High Energy Physics, 1-10ho, Tsukuba, Ibaraki 305, Japan
Carbon glass resistance (CGR) thermometers were used for accurate temperature measurement under high magnetic fields around 1.8-1.9 K. For accurate temperature measurement of the LHC (Large Hadron Collider) model magnet, we investigated the magnetoresistance of CGR thermometer up to 9.3 T at around 1.8 K. The magnetoresistance at 4.3 K under magnetic fields of up to 7.7 T was also measured for comparison. At temperature around 1.8 K, the maximum temperature errors of 0.3-0.8 % (AT/T) were observed at 4-5 T. At 4.3 K, the temperature errors increased monotonously, and reached -1.5 % at 8 T. Measured data at 1.8 K-1.9 K are compared with predicted curve based on the results obtained by other researchers.
INTRODUCTION Due to an increase of application of high magnetic fields by using NbTi superconducting magnets, large superfluid helium cooling system is now widely used to cool those magnets at around 1.8 K. For example, the Large Hadron Collider (LHC) project requires to cool several thousands of high field superconducting magnets down to 1.8 K. The precise temperature measurement of the magnet is one of the essential issues, because almost all kind of low temperature thermometers are affected by the existence of strong magnetic fields. The carbon glass resistance(CGR) thermometer is applicable even under the magnetic fields [ 1-3]. However, data on magnetoresistance of CGR thermometer at around 1.8-1.9 K are not available. In order to obtain those data, two CGR thermometers were placed in the bore of the LHC model dipole magnet which can provide the magnetic fields up to about 9 T at 1.8 K. The thermometers were aligned along parallel and perpendicular orientation to magnetic fields in addition to the Hall effect magnetic field sensors. This paper presents the results of the magnetoresistance measurements of the CGR thermometers at 1.8-1.9 K. EXPERIMENTAL SETUP A magnetoresistance measurement of CGR thermometers was carried out during the training quench tests of the LHC model magnet. A schematic drawing of the pressurized superfluid helium experimental system is shown in Figure 1. The cryostat is designed for dedicated test of LHC high field superconducting magnet. The magnet is cooled by pressurized superfluid helium in the 1.8 K bath with a capacity of about 500 liters. A details of the system is presented in another paper at this conference [4]. A type of CGR-1-1000 thermometers were placed in a G-10 sensor bar together with the Hall effect magnetic field sensors (BHA-921, Bell). The sensor bar was put into the magnet bore and used for identifying a quench origin. These CGR thermometers were calibrated by the Lakeshore Cryoelectronics in 1991-2. Constant current of 1 BA was supplied by a battery-operated current source (Model 101, Lakeshore cryoelectronics). Figure 2 shows the set up details of two CGR thermometers and Hall sensors in the quench antenna bar. The magnetoresistance of the CGR thermometer was measured at 4.33 K, 1.89 K and 1.81 K when the magnet was energized for the training quench test. Table 1 lists parameters including temperature, current increasing rate (ramp) and current holding time. A magnet energizing procedure differs from case to case, however, 1.8 K bath temperature was quickly equalized in every case. Reference temperature was monitored by an another calibrated CGR thermometer placed near the magnet in the 1.8 K bath, but outside of the magnet. 649
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Figure 1 Pressurized superfluid helium system
Figure 2 Setting configuration of CGR thermometers and Hall sensors in G-10 sensor bar for measurements: side view (a) and top view (b) of a sensor bar
Table 1 Measuring condition Temperature
Magnetic field
Magnet current ramp condition
4.33 K 1.89 K 1.81 K
0~7.75 T 0~8.97 T 0~9.33 T
20 A/s, 5 min. holding at every 2000A step 10 A/s, continuously increasing 15 A/s, 1 min. holding at every 2000A step
EXPERIMENTAL RESULTS OF MAGNETORESISTANCE OF CGR THERMOMETER We define a percentage of the magetoresistance as follows: AR _- R(Tref,B)-R(Tref,0) X 100 (%) R R(Tref,0)
(1)
where, Tref is reference temperature of 1.8 K helium bath, and R(Tref, B) is the resistance of CGR thermometer at temperature Tref under the magnetic field B in Tesla. Table 2 summarizes the measured data at each temperature. Here, RB//is the resistance of CGR thermometer in parallel magnetic field to the thermometer, and RB+ is for the case when magnetic field was perpendicular to the sensor. Percentage of temperature errors by the magnetoresistance is defined as follows: AT = T[R(Tref, B)]-Tref X 100 (%) T Tref
(2)
where, T[R(Tref, B)] is temperature corresponding to the resistance value R(Tref, B) of CGR thermometer. Parallel magnetoresistance obtained at 1.81 K in this study is inserted in Figure 3 together with already published data[ 1], for comparison. Our results seem to be consistent with the predicted curve based on the results by previous research.
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Table 2 Measured magnetoresistance at various temperature. Magnetic field was measured by a Hall sensor near Re//, and the bath temperature fluctuation is less than 2 mK for 1.89 and 1.81 K. T(K)
B(T)
RB//(k~)
4.33
0 1.6 3.2 4.8 6.4 7.5 0 1.6 3.2 4.8 6.4 8.0 8.8 0 1.6 3.2 4.8 6.4 7.9 8.7
10.50 10.54 10.65 10.76 10.88 10.98 68.46 68.37 66.59 66.25 67.01 68.78 69.74 88.82 87.79 84.95 84.37 85.46 87.42 88.55
1.89
1.81
dR/R//(%) +0.4 +1.4 +2.5 +3.6 +4.6 -0.1 -2.7 -3.2 -2.1 +0.5 +1.9 -1.2 -4.4 -5.0 -3.8 -1.6 -0.3
RB+(I~) 11.17 11.22 11.35 11.47 11.62 11.73 77.24 77.33 75.31 75.47 77.10 79.62 80.98 101.1 100.2 97.50 97.40 99.16 102.3 103.9
dR/R+(%) +0.4 +1.6 +2.7 +4.0 +5.0 +0.1 -2.5 -2.3 -0.2 +3.1 +4.8 -0.8 -3.5 -3.6 -1.9 +1.2 +2.8
Figure 3 Obtained data at 1.81 K is superimposed on published data [ 1] Figure 4 shows the temperature errors in %, due to the magnetoresistance and the orientation dependence of CGR thermometer at three different temperatures. At 4.33 K, the temperature error increases monotonically as the magnetic fields increase and it exceeds -1.5 % at 8 T. A very little orientation dependence was observed at this temperature. At 1.89 and 1.81 K, the temperature errors have peak values near 4-5 T. Concerning field orientation dependence of the magnetoresistance, perpendicular fields affect slightly more on magnetoresistance than parallel fields. These features of the magnetoresistance of the CGR thermometer qualitatively agree with the already published data [ 1-3]. Figure 5 shows the temperature errors in mK to the magnetic fields change at three temperatures. Only the results for perpendicular field is shown here. At around 1.8 K, CGR thermometer can be used for temperature measurement with temperature errors of about +12 mK under the magnetic fields up to 9T.
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Figure 4 Temperature errors in % vs. magnetic fields applied in parallel and perpendicular at 4.33 K, 1.89 K and 1.81 K.
Figure 5 Temperature errors in mK vs. perpendicular magnetic fields at 4.33 K, 1.89 K and 1.81 K SUMMARY The magnetoresistance of the CGR thermometer at around 1.8 K is measured first time. For accurate temperature measurement at around 1.8 K under the high magnetic fields up to 9 T, the magnetoresistance of the CGR thermometer was investigated for perpendicular and parallel magnetic fields orientation to the CGR thermometer. For the magnetic field range of 0-9 T, temperature errors due to the magnetoresistance were within +12 mK, indicating positive peak at around 4-5 T. In these temperature ranges, as small difference as 10 mK was observed between two fields orientation. The data obtained in this study are consistent with the expectation based on previous data [ 1]. REFERENCES Sample, H.H et al., Low-temperature thermometry in hgh magnetic fields. V. Carbon-glass resistors, Rev Sci Instrum (1982) 53 1129-1136 Rubin, L.G. et al., Some practical solutions to measurement problems encountered at low temperature and high magnetic fields, Adv Cryog Eng (1986) 31 1221-1230 Hua, J. et al., Behaviour and accurate thermometry of carbon-glass resistance thermometers at low temperature and in high magnetic fields up to 7 T, Cryogenics (1987) 27 90-92 Kimura, N. et al., Cooling performance of pressurized He II cryostat for LHC superconducting model magnet test, presented in this conference ICEC16/ICMC (1996) PS2-el-02
New Type of Thin-Film Germanium Resistance Thermometer for Use in a Wide Temperature Range
Vadim Mitin, Yurii Tkhorik and Evgenii Venger Institute of Semiconductor Physics of the National Academy of Sciences of Ukraine Pr.Nauki 45, 252650 Kiev-28, Ukraine
Two new types of thin-film Ge resistance thermometers were fabricated. These temperature sensors may be used in the temperature range 1.5 to 350 K and under high magnetic fields.
INTRODUCTION In cryogenic engineering and low temperature physics, there is a continuing requirement for sensors to measure local temperatures in the presence of high magnetic fields. This necessitates designing and fabrication of miniature temperature sensors which could operate in wide ranges of both temperatures and magnetic fields. At present there exist cryogenic temperature sensors of different types. To measure low temperatures, the resistance thermometers fabricated on the base of bulky materials (germanium, rhodium-iron, carbon-glass and platinum) are most often used. Some radically new potentialities in developing cryogenic resistance thermometers arise due to application, as sensitive elements, of epitaxial semiconductor films on insulating substrates. A thin-film technology is compatible with integration technology in semiconductor devices manufacturing. Use of microelectronic technology in fabrication of cryogenic temperature sensors makes it possible to improve their technical parameters, to reduce both sizes and costs, and to widen the fields of their application. It also opens opportunities for manufacturing sensors either as discrete devices, or as parts of integrated systems. We have proved [ 1, 2] that Ge films on the semiinsulating GaAs substrates can serve as a generalpurpose sensitive material for fabrication of temperature, deformation and magnetic field sensors. This stems from the fact that, during the process of epitaxial deposition of the Ge in a vacuum onto GaAs substrates, Ga and As atoms strongly diffuse from the substrate into the growing film. Diffusion processes at the film-substrate interface depend primarily on the process conditions of the epitaxy. By changing the conditions of epitaxy, one is able to obtain single-crystal Ge films of both p- and n-type with different doping levels and compensation degrees. Besides, one can vary the film structure from poly- to single crystal one. The above factors permit realizing in Ge films a wide range of electrical properties needed to produce various sensors. Besides, Ge-GaAs is an ideal heteropair, because lattice constants of its components, as well as their thermal expansion coefficients, are very close. The lack of thermal strains in a thermosensitive element is of particular value for device operation in a wide temperature range. Here we present the results of developing two types of resistance thermometers on the base of thin Ge films. We have investigated the conduction mechanisms which are responsible for thermal sensitivity. The effect of magnetic field on the thermometric characteristics was also studied. RESULTS AND DISCUSSION Fabrication Practice We prepared films by thermal evaporation of Ge in vacuum onto the semiinsulating GaAs substrates. We used standard microelectronic techniques to produce sensitive elements. Technology of fabrication included the following operations: (i) deposition of an ohmic contact by thermal evaporation of a metal in 653
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a vacuum; (ii) photolithography to form topologies of both sensitive region and contact areas; (iii) wafer slicing, (iiii) packaging. The sensitive elements, measured (0.5x0.5x0.2) and (1.0xl.0x0.2) mm for different models of sensors, were placed into packages measured: Pl-(Q0.6x3.5), P2-(3.0xl.5xl.2) and P3-(Q0.9x0.6) mm. An external view of the package is shown in Figure 1.
Figure 1 Construction of packages Thermometric Characteristics and Electrical Conduction Mechanisms of Ge Thin-Film Resistance Thermometers Thermometric curves for two types of resistance thermometers are given in Figure 2. The sensitive elements of thermometers were made under various technological conditions and so differed in their thermometric characteristics. The fabrication practice of thermosensitive elements on the base of Ge films on GaAs substrates has a special feature, namely, a possibility of producing the resistance thermometers having different temperature dependencies of resistance. This can be achieved by varying the doping level and compensation degree of the films. The impurity spectrum in the gap of Ge films on GaAs substrates is complicated. There exist three main sources of free charge carriers in Ge films, namely, Ga atoms (acceptors), As atoms (donors) and structural defects. Both doping level and compensation degree of the films essentially depended on the conditions of epitaxy.
Figure 2 Characteristics of resistance thermometers
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For doped and compensated, as well as disordered, semiconductors low-temperature conduction is known to be determined by various hopping mechanisms. A general expression for the temperature dependence of semiconductor resistivity is P = Po
Tm
exp(Tx / T) •
(1)
For x = 1 this expression corresponds to the activated conductivity, the activation energy of the conduction process being ~ = kTx., where k is the Boltzmann constant. Variable activation energy (x < 1) corresponds to the variable range hopping. Depending on the density of state, g(~) , behavior near the Fermi level, x takes various values ( e.g., x = 0.25 in the Mott model [3] and x = 0.5 in the Shklovskii model [4]). Shown in Figure 3 are the reduced activation energy, W, vs temperature curves for Ge films of two types W = (l/T) (d lnp / d ln T)
(2)
From analysis of the temperature dependence of conductivity one can conclude that at low temperatures conductivity cannot be expressed as an exponential function with constant activation energy. This means that in this case conductivity is determined by variable range hopping.
Figure 3 Dependence of the reduced activation energy on the temperature At temperatures T<10 K x values for the two types of thermometers differ essentially (0.25 and 0.66). This indicates at different densities of state g(e) and may result from the differences in doping levels, compensation degrees and impurity space correlation. All these factors are determined by the conditions of film fabrication. So, by varying these conditions, one can produce films having different temperature dependencies of conductivity and fabricate temperature sensors which can operate in different temperature ranges. Resistance Changes in Magnetic Fields We investigated the behavior of thin-film germanium resistance thermometers in magnetic fields at a temperature of 4.2 K. Transverse magnetoresistance ( j l B I[ n , where n is a vector along the normal to the film surface, B and j is a vector of magnetic induction and current, respectively) for two types of resistance thermometers are given in Figure 4.
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Figure 4 Dependence of the magnetoresistance on the magnetic field at 4.2 K Same preliminary investigations have shown that the thermometer resistance change in a magnetic field essentially depends on the conditions of film fabrication, i.e. on the conduction mechanism responsible for thermosensitivity. CONCLUSIONS The resistance thermometers were developed on the base of Ge films on GaAs substrates. They are intended for use in a wide (1.5 to 350 K) temperature range. A special feature of the developed technology is that it enables one to produce miniature resistance thermometers of various thermosensitivity, as well as resistivity, values. This is of particular convenience for some services. In addition, some special-purpose resistance thermometers may be fabricated, which provide high accuracy if measurements in high magnetic fields. REFERENCES Mitin, V. In: International Cryogenic Engineering Conference-15 Genova, Italy (1994) Cryogenics (1994) 34 437-440 Mitin, V.F. and Tkhorik, Yu.A. In: International Conference on Microelectronics-20 Nis, Serbia (1995) 2 553-557 Mott, N.F. and Davis, E.A., Electronic Processes in Non-Crystalline Materials, 2nd ed., Clarendon Press, Oxford (1979) Shklovskii, B.I. and Efros, A.L., Electronic Properties of Doped Semiconductors, Springer Verlag, Berlin (1984)
THERMOACOUSTIC TACONIS OSCILLATIONS IN HELIUM-4 LIQUID LEVEL SENSORS S. YOSHIDA*, K. V. RAVIKUMAR, J. WILLIAMSON, N. PAPAVASILIOU, T. H. K. FREDERKING Chemical Engineering Department, BH 5531, School ofEngng & Applied Science, University of California, Los Angeles CA 90095, USA *) Permanent address: Taiyo Toyo Sanso, Kawasaki, Japan
Abstract In contrast to pulse tube coolers, operated at low frequency in forced oscillation regimes, dipstick liquid level finders have been operated in "free oscillation" conditions. Dipstick operation is characterized by simplicity, ruggedness and low heat input when thin-walled tubes of low thermal conductivity are used. We have measured liquid level responses for a comparison of oscillations. The results are compared with a simplified resonator model. Data are consistent with the model results. INTRODUCTION Heat-induced sound has received attemion by glass blowers at high temperatures, e.g. [ 1,2] ,and at low temperature (T) in liquid helium vessels [2]. Once thermoacoustic oscillations are triggered, fast evaporation of the liquid is among undesirable consequences. Guidelines for liquid helium users, e.g. [3], have included precautions preventing oscillations, In contrast, the thermoacoustic oscillations have been used in liquid level finders [4]. More recently, a re-inspection of low frequency thermoacoustics has taken place in conjunction with pulse tube coolers and similar equipment [5,6]. Our paper has the purpose of understanding details of the oscillations. STABILITY LIMITS AND MODELLING Stability limits for thermoacoustic oscillation regimes have been evaluated by Rott and collaborators [2]. For instance, the order of magnitude of stability limits appears to be the same for helium gas and nitrogen gas oscillations in Sondhauss tubes (sphere connected to a long duct). Further, hydrogen oscillations have been reported for a cylinder geometry [7]. Another oscillation aspect is the convection onset in fluids restricted to confined geometries, e.g. gravitational convection. Insulation experiments and stability tests have been conducted [8,9]. During the multilayer insulation tests [8] it was found that Helium-4 gas conditions near 300 K in a long vertical duet of 2.54 em diameter were influenced but little by the presence or absence of the gas inside the duet. Upon application of a large temperature gradient from the top, oscillatory motion started immediately [9]. This 657
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behavior has been consistent with both, Rott diagrams [2] and Rayleigh-Benard stability predictions [ 10]. On the basis of these observations we model the dipstick gas oscillations using simplifying resonator assumptions: Consistem with the observations of stagnant gas conditions [8], the Helium-4 gas in the top detector section and the warm end of the dipstick acts as adibatic "gas spring". In the lower cold gas mass inside the tube, motion is possible because of a very low shear viscosity. A steep temperature gradiem is assumed in between the cold and warm section in agreemem with experimental evidence [ 11]. We obtain the frequency (f) as a function of the sound velocity (c), of the total gas spring volume V and of the ratio of the effective cross section of the cold mass and its length
(Am).
f= c ~/(A/L )/V /(2 rr)
(1)
The ideal gas value of the sound velocity is expressed as c = ~rk (R/M) T
(2)
R = universal gas constant, M molecular mass, k ratio of the specific heat at constant pressure to the specific heat at constant volume, T temperature of the cold section. For Helium-4 we note that the ratio of the sound speed at 300 K to the 4.2 K-value is 8.45. EXPERIMENTS In the experiments we have used a rubber membrane on top of the dip stick. A small magnet is attached to the center of the membrane. A pickup coil is wound around the circumference of the top section. The voltage signal induced by the gas-membrane oscillations is recorded. Once the tube is inserted into the liquid helium vessel and cooled down beyond the stability limit, oscillations start in the dipstick system. Table I lists geometry details including reference [ 12, 13]. Figure 1 is an example of oscillations in the system # 1. The observed frequency is in order of magnitude agreement with [ 12]. For the tube # 2 with a thick wall, Fourier transforms, power spectra and phase space trajectories have been evaluated. Examples are figures 2, 3 and 4. The usual detection methods described in reference [4] are not sensitive entirely. FFT records and related diagnostics provide an enhanced sensitivity and resolution of oscillation details. CONCLUDING REMARK Our experiments indicate consistency in frequency prediction of our model with data. The dipstick with a heavy wall introduces an enhanced heat leak and a lack of sensitivity. This is compensated in part by FFT diagnostics.
ICEC16/ICMC Proceedings ACKNOWLEDGMENTS. Early stability and oscillation work has been supported in part by NASA and NSF. The partial support, interest and encouragement ofTaiyo Toyo Sanso Co, Ltd. is gratefully aeknowledged. We thank our SRP students involved in this project. REFERENCES
10 11 12 13
Rayleigh, J. W. S. Lord, "Theory of Sound", 2nd ed vol 2 (reprinted Mcmillan London 1944). Rott, N. Thermoacoustics, Academic Press (1980) 20 135 - 1 7 5 Olsen, J. L., Liquid helium user manual, ETH 1967 White, G. K., "Experimental Techniques in Low Temperature Physics" 3rd ed (1989) Oxford Clarendon. Radebaugh, R., Advances in eryocoolers, ICEC 16, (1969) paper PL6 Swift, G. W., Thermoacouste engines and refrigerators, Phys Today (1955) 45 22-29 Gu, Y. F. and Timmerhaus, K. D., Thermal acoustic oscillations and triple point liquid hdrogen systems, Intern Journ Refrig. (1991) 14 282-291 Kamioka, Y., Chuang, C. and Frederking T H K, Simplified techniques for the evaluation ofmukiplayer insulation, Proc. ICEC 9, Kobe, 1982 568-571 Kamioka, Y. Chuang.C, and Frederking, T.H.K.,Kinematics of fluid flow instabilities in low temperature transfer systems, Refi'ig. Sci Technol IIR Paris (1980) 1 311 - 317 Catton, I., Ph. D. thesis, UCLA 1966. Bannister, Liquid Helium Technology, Proc. IIR Boulder Meeting, (1966) Clement, J. R. and Gaffney, J. Thermal oscillations in low temperature apparatus, Proc 1st Cryog Conf., Adv Cryog Eng (1960) 302 - 306 GaIfney, J. and Clement, J. R., Liquid helium level-finder, Rev Sci Instrum (1955) 26 620-621.
TABLE I . TUBE GEOMETRIES INSIDE DIAMETER OF TUBE
LENGTH OF TUBE
DIAMETER HEIGHT VOLUME OF OF OF RESERVOIR RESERVOIR RESERVOIR
#1
0.106 in
49.5 in
1.265 in
0.957 in
1.203 cu in
#2
0.075 in
43.25in
0.923 in
0.464 in
0.310 cu in
G-C [13]
0.125 in 1 in = 2.54 cm
34 in
0.613 in
0.24 in
0.0707 cu in
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Fig. 1. Oscillation re6ord, tube # 1" induced voltage versus time.
Fig. 2. Oscillations, tube # 2, induced voltage versus time.
Use of Strain Gages for Low Temperature Thermal Expansion Measurements Robert P. Walsh National High Magnetic Field Laboratory, 1800 E. Paul Dirac Dr., Tallahassee, FL 32306 The use of strain gages for low t e m p e r a t u r e t h e r m a l e x p a n s i o n m e a s u r e m e n t s is increasing, due to it's convenience, and relative accuracy. The determination of thermal expansion with strain gages is complicated by "thermal o u t p u t strain", a p h e n o m e n a caused by the t e m p e r a t u r e d e p e n d e n c e of the gage. If thermal o u t p u t strain is properly compensated, strain gages may be used to measure thermal expansion in much the same manner they are used to measure applied strain. An algebraic method to compensate strain gage thermal o u t p u t over the temperature range 300 K to 4 K is reported here. The results of the t h e r m a l e x p a n s i o n tests on two s t r u c t u r a l alloys and two superconductors are also reported.
INTRODUCTION Generally, the t h e r m a l e x p a n s i o n of solids is m e a s u r e d using a "dilatometer", an instrument which works on the ability to detect the differential expansion between the test material and a low expansion reference material. The m e t h o d of using strain gages for thermal expansion m e a s u r e m e n t [1] also relies on the use of a reference material with k n o w n expansion. It is attractive because it allows rapid and accurate determination of expansion properties of a b r o a d range of material classes. Another a d v a n t a g e is the capability to test various specimen geometries that cannot be a c c o m m o d a t e d with other techniques. Applications of strain gages for expansion measurements have been reported to be viable over a range of temperatures on various materials [2 - 4]. A l t h o u g h this t e c h n i q u e is applicable to m e a s u r i n g t h e r m a l e x p a n s i o n d o w n to liquid h e l i u m temperature (4 K), little or no information has been reported. Research on strain gages available for use at cryogenic temperature, has shown that gages made with a Ni-Cr alloy conductor have the best gage-to-gage reproducibility [5]. A l t h o u g h this characteristic is favorable for their use in thermal expansion tests, the electrical resistance versus t e m p e r a t u r e characteristics of Ni-Cr alloys shows a slope inversion at a r o u n d 20 K that is unfavorable. This slope inversion coupled w i t h a tendency for gage self-heating in this temperature regime, necessitates the careful gage characterization that is detailed here. PRINCIPLE OF THE MEASUREMENT METHOD W h e n a strain gage is m o u n t e d on a stress-free specimen, and the t e m p e r a t u r e of the 661
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specimen is changed, the resistance of the strain gage changes. The change in the electrical o u t p u t of the gage is referred to as "thermal o u t p u t strain" (TOS) and is caused by a combination of the specimen's expansion and the strain gage's temperature dependence. If the TOS is compensated correctly, strain gages can be used to measure thermal expansion. The technique for compensating the measurement error requires using strain gages from the same m a n u f a c t u r i n g lot to ensure they are as close to identical as possible. T r a d i t i o n a l l y this c o m p e n s a t i o n has been done electrically, using the canceling characteristics of the wheatstone bridge on a d u m m y "identical" gage on a reference specimen. There are drawbacks; the d u m m y specimen must be measured along with the test specimen and kept at the same temperature. In addition if there are problems with the d u m m y gage, the data will not be recoverable. Another method, used here, is to algebraically compensate the TOS data. The strain gage lot is characterized by individually testing a few strain gages from the lot on a reference material with k n o w n thermal expansion. The mathematical difference between the reference specimen's expansion at a particular temperature and the TOS signal at the same temperature, is the error that prevents measuring thermal expansion directly. The magnitude of the error is temperature dependent and must be characterized as a function of temperature. The magnitude of the error is independent of the material to which the strain gage is b o n d e d and can be thought of as the electrical output that the gage would have if it were b o n d e d to an "ideal reference material" having zero expansion or contraction. This measurement error data can then used to algebraically compensate TOS data on a material of interest to obtain it's thermal expansion. TEST MATERIALS Two alloys, AISI 316LN and 2024-T4 A1 are measured and compared to published data. The samples are in plate form, 40 mm square by 10 mm thick. Two low temperature, superconductor materials are also tested. They are composite wire samples of C o p p e r / N b T i and C o p p e r / N b 3 S n . The C o p p e r / N b T i wire has a 3.6 : 1 copper to NbTi ratio. The C o p p e r / N b 3 S n wire composition is ; 0.41 Cu, 0.03 Nb, 0.42 CuSn, 0.14 Nb3Sn. The samples are 50 mm long, having a rectangular cross-section of 1 mm by 2 mm. Commercially pure copper, also referred to as oxygen-free high-conductivity (OFHC) copper, with copper content > 99.95 %, is used as the reference material with a known thermal expansion [5]. EXPERIMENTAL METHOD Commercially available, foil-type, resistance strain gages from the same manufacturing lot are used. The strain gages are made from Ni-Cr alloy and have 350 f2 resistance. Two different gage types (from two different lots) are used because a thin gage width is needed for the more difficult application on the superconductor wires. The adhesive used is a two c o m p o n e n t epoxy (M-Bond AE-15) suitable for cryogenic use. A commercial strain indicator system is used that has an estimated accuracy of +/- 1 microstrain. Specimen temperature is controlled in a vacuum-insulated stainless-steel dewar.
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The temperature measuring system uses a silicon diode temperature sensor and has an estimated accuracy of + 0.5 K. Thermally conductive grease is used to ensure thermal contact of the sensor to the specimen. The active strain gage is wired to the strain conditioner using a quarter bridge configuration with a three lead wire connection. The excitation voltage is low (0.5 Vdc) to avoid gage self-heating that can occur due to the power density at the gage [6]. The room temperature gage factor, supplied by the gage manufacturer is used throughout. Tests are conducted by thermal cycling the test specimen (the material of interest or the reference material) from room temperature (293 K) to liquid helium temperature (4 K). The strain indicator output is balanced at 293 K to give a zero strain readout. The test specimen is cycled from 293 K to 4 K to establish confidence in the response of the gage and associated instrumentation, i.e.. : the strain gage signal returns to zero at 293 K. Data is obtained while the specimen is slowly cooled in a dewar. The specimen is cooled down to 77 K using liquid nitrogen and on to 4 K with liquid helium. A reversal of this scheme is used to obtain data on the warm-up cycle. The characterization of the strain gage lot is done by testing "identical" gages from within the lot on the reference expansion material (OFHC copper). The strain gage error (the difference between TOS and the actual copper expansion) is calculated and plotted as a continuos curve, over the t e m p e r a t u r e range, and then curve fit with a polynomial expression. The descriptive equation is used for compensation of the TOS data taken in thermal expansion tests. RESULTS A s u m m a r y of the test results is found in Table 1. The estimated accuracy of the values are + / - 5 %, based on the characterization of the strain gage lot with the OFHC copper reference material. Correction for the temperature dependence of the gage factor, not done here, can further refine the accuracy (~2 %). Figure 1 shows a graph of the TOS data for the a l u m i n u m alloy along the thermal expansion curve that is obtained by mathematical correction of the TOS. This is a graphical representation of the strain gage error. Figure 2 shows test results along with published data [7] for aluminum alloys. CONCLUSIONS The characterization of "identical" gages from the same strain gage lot is necessary for converting TOS data to thermal expansion data. This is the by far the largest source of error and, it's compensation could be sufficient depending on the accuracy desired. Special attention should be given to the potential for gage self-heating which may not be excessive enough to cause noisy data but could cause a subtle data shift, that is Table 1. Thermal Expansion Test Results TEMP (K)
THERMAL EXPANSION (%) 2024 T4 AI
77 4
-0.38 -0.41 ,
316LN -0.28 -0.29
SS
Cu/NbTi
Cu/Nb3Sn
-0.28 -0.31
-0.25 -0.29
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Figure 1. TOS and Compensated Data.
Figure 2. Comparison with literature.
reproducible [6] and could lead to misinterpretation of the data. Temperature measurement is critical to the accuracy of these measurements. Taking sufficient data, over a number of thermal cycles, to obtain good agreement will help assure that thermal equilibrium conditions are achieved. ACKNOWLEDGMENTS This work was supported by the State of Florida and the National Science Foundation through NSF Cooperative Grant No. DMR 9016241. The author wishes to thank R.P. Reed and C. Ferrero for guidance and encouragement. REFERENCES ,
~
~
,
~
,
o
Measurements Group, Inc., Tech Note TN-513, "Measurement of Thermal Expansion Coefficient using Strain Gages", 1986. Finke, T.E. and T.G. Heberling, "Determination of Thermal Expansion Characteristics of Metals Using Strain Gages", Proceedings, SESA, (now, SEM), Vol. XXV, No. 1, 1978, pp. 155-158. Poore, M.W. and K.F. Kesterson, "Measuring the Thermal Expansion of Solids with Strain Gages", Journal of Testing and Evaluation, ASTM, Vol. 6, No. 2 (March 1978). pp. 98-102. Valentich, J., "Thermal Expansion of Solids from -216~ to 173~ Using Strain Gauges", Materials Testing and Evaluation Laboratories, Westinghouse Electric Corporation Research and Development Center, Pittsburgh, PA. Simon,N.J., Drexler,E.S., and Reed,R.P. "Properties of Copper and Copper Alloys at Cryogenic Temperatures" Feb. 1992, NIST Monograph 177. Ferrero, C., "Stress Analysis Down to Liquid Helium Temperature", Cryogenics, Vol. 30, March 1990, pp. 249-254. Clark, A.F. "Thermal Expansion" chapter from "Materials at Low Temperatures" edited by Reed, R.P., and Clark, A.F.,ASM 1983.
Strain Measurements of Stainless Steel at Low Temperatures using Electronic Speckle Pattern Interferometry Sumio Nakahara, Hidetomo Sakiyama, Shigeyoshi Hisada, and Takeyoshi Fujita Dept. of Mechanical Engineering, Kansai University, 3-3-35, Yamatecho, Suita, Osaka 564, Japan Electronic speckle pattern interferometry is applied to the measurements of out-ofplane deformation and in-plane, and strain of some kinds of stainless steel during the cooling down process from room temperature to liquid nitrogen temperature. Interference fringe patterns represent in-plane and out-of-plane deformation (warp), depend on cooling rate, temperature gradient at the surface of the specimen, and on the rolling direction of the specimen. The deformation and temperature of the specimen are also monitored by using resistance strain gauges and thermocouples, respectively. By way of an application, thermal expansion coefficient of the specimen due to various cooling down processes are examined based on those experimental results such as in-plane deformation and temperature measurement.
INTRODUCTION The thermal contraction as the thermal behavior of materials in the cooling down process, is one of the most important properties. From this point of view, the deformation process has been studied by using various techniques. The deformation and strain of materials has conventionally been measured by applying resistance type strain gauges. In the low temperature regions, however, strain gauges are not so convenient because the gauge factors are not kept constant for the wide range of temperatures. In contrast with the strain gauges, optical methods of deformation and strain measurements, such as the holographic and speckle interferometry, and speckle photography, are generally independent of temperatures, and are able to make the measurement with non-contact. Those optical methods also yield two-dimensional (full-field) information about the deformation over the whole surface of the object. This is an advantage of the optical method in detecting the defects of materials, especially in the regions of low and high temperature. In the high temperature ranges, laser speckle photography (LSP) technique has been used to measure strain distributions up to 1170 K by Stetson [1] and thermal expansion coefficient of up to 1473 K by Jiang [2]. There is also the work of Malmo et al., who used the electronic speckle pattern interferometry (ESPI) techniques to study deformations, oxidation shell growth, and melting zones of several metals up to about 3300 K [3]. In the low temperature region, on the other hand, the LSP work in our laboratory has also shown its usefulness when it was applied to the observation of thermal contraction of stainless steel, aluminum, etc. cooled down from the room temperature to the liquid nitrogen temperature (LNT) [4]. In this study, ESPI method is applied to study the deformation behavior of stainless steel during cooling down process from the room temperature to LNT. By combining video tape recorder with ESPI, out-of-plane deformation and in-plane, and strain of stainless steel are measured for hours. The deformation and temperature of the specimen is also monitored by using resistance strain gauges and thermocouples, respectively. By way of an application, thermal expansion coefficient of the specimen due to various cooling down processes are examined based on those experimental results such as in-plane deformation and temperature measurement.
EXPERIMENTAL ARRANGEMENT In speckle measurement system, the correlation of the speckle patterns generally should be kept during measurement. However, the correlation of the speckle patterns are not easy to keep during the temperature change because of the atmospheric convection around the specimen due to the temperature gradient, and also of the change of the material structure; the condensation of water vapor on the specimen surface prevents the observation. In the present study, the problems of the atmospheric convection and the condensation of the water vapor were eliminated by setting the specimen in a vacuum vessel of the cryostat (see below). He-Ne laser (20 mW) is employed in this ESPI. For the out-of-plane deformation measurement, a conventional ESPI system is constructed. Symmetrical illumination of the specimens is used to measure the in-plane component of displacement, as shown in Fig. 1. Figure 1 illustrates the experimental arrangement, the optical set-up for the ESPI method and a cross sectional view of the cryostat. The cryostat has a vacuum 665
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chamber with an observation window of a glass plate of 10 mm thickness. The disk or rectangular shaped specimen was held horizontally downwards by a hollow Cu tube. As a specimen, commercially available stainless steel (SUS 316, 304, 310 etc.) were examined. The disk shaped specimen is 150 mm diameter and 2 mm thickness. The rectangular shaped specimen has a dimension of 150 x 30 x 2 t or 3 t mm 3. All the specimens have a hole 20 mm diameter at the center for attaching to a cold end. By pouring liquid nitrogen into the inner vessel of the cryostat, the temperature of center of the sample is rapidly lowered to about 77K, while the outer portion of the specimen is gradually cooled down by the thermal conduction through the specimen. The temperature distribution of the specimen was monitored at the cold end and the points 30 and 60 mm away from the center by the chromel-alumel thermocouples (K-type) of diameter of 0.05 or 0.1 mm. The temperature data were recorded through a microcomputer for the whole period of the observation. In order to prevent the atmospheric convection and the condensation of water vapor to the sample Fig. 1 ESPI setup and cross sectional view of cryostat surface, the air in the vessel was first subfor deformation measurement on an object in cooling down stituted by helium gas, and then, was evacuated to a fair vacuum ( ~ 10 -4 torr). The speckle images were recorded using the 1/2 inch CCD (510(H) x 492(V), NTSC-interlaced scanning type), through a f/8 lens of 25 or 50 mm focal lengths. The distance from the specimen to TV camera was adjusted so that the image of the object substantially filled the whole area of the CCD. Intensity correlation in our ESPI is observed by the subtraction process of video signal. It takes more than an hour until the temperature of the stainless steel specimen becomes steady after the initiation of cooling. Although this time is long enough for keeping the stability of the interferometric apparatus, in this ESPI experiment, the two frames of this digital picture corresponding to the specimen in its initial and deformed states are stored and processed at short-range interval in the personal computer. Processed images (fringe patterns) are then displayed on the monitor and simultaneously are recorded in the video tape recorder. At the same time, an analog video signal from TV camera is recorded in another video tape recorder. By reproducing the recorded signal, it is possible to measure repeatedly the deformation of the specimen at the same or different situation. A sequence of interferograms (processed images) is obtained at intervals of some 1 to 10 minutes while looking at an interferogram on the TV monitor that is displaying the processed data. Those interferograms cover the entire time range of the object deformation during the cooling process.
Fig. 2 Temperaturetraces at three points on the stainless steel specimen, 30 mm and 60 mm away from cold end, and cold end (center) during cooling down process
Fig. 3 In-planedisplacement change of the disk shaped specimen during cooling down process at four points of the specimen along rolling line and vertical line of rolling.
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The correlation of the combined speckle pattern with the original pattern will take place to produce fringes when the following relation ship is satisfied [see Ref. 5] 2du sin a
= n a.
d w = n a./2
(in-plane displacement measurement)
(out-of-plane displacement measurement)
(1) (2)
where d u and d w are the x- and z-component of the displacement, respectively, 2 a is the angle between the illumination beams in in-plane displacement measurement, ~. is the wavelength of the light used and n is an integer. The sensitivity of the set-up can be changed by varying the illumination angle, 2 a . In the case of the out-of-displacement measurement, the equation of the observed fringe contours is given under small-angle approximation and in the configuration of collimated illumination at near zero incidence. RESULTS AND DISCUSSION As the temperature changes, the totality of the scatterer movements exceeds a certain limit to keep the correlation between the speckle patterns, then the loss of fringes results. An usual way to obtain the scattering at the surface part would be to paint the surface with mat white, but it was not successful at the temperature range of the present work. Satisfactory results were obtained by sandpapering (#400), and then by etching using a etchant. The temperature traces from the thermocouples are shown in Fig. 2 at the three points Of the specimens. In this equipment, it takes about 70 minutes for the stainless steel, for the temperature to reach the steady state. The temperature differences between the front (observed) and the back surfaces were less than a few degrees for stainless steel. Figure 3 shows the in-plane displacement change of the rolling line and vertical line of rolling of the stainless steel specimen due to thermal contraction at the four points of the specimen during the cooling down process. Figure 4 shows the fringe patterns of out-of-plane displacement for the stainless steel specimen. In the case of out-of-plane deformation, the ESPI method suffers for the intrinsic sign ambiguity in the assignment of the displacement vector field. The direction of warp of the specimen was confirmed by a laser displacement sensor. In Fig. 5 the in-plane deformation fringes due to the thermal contraction are shown for various temperature situation. A fringe pattern was sometimes observed even on the center region of the specimen. We assume that these fringes represent the uniform displacement of the optical system as a whole. The net displacement of the sample due to the temperature decrease would be obtained by vectorial subtraction of this uniform displacement from the experimental ones for each point.
Fig. 4 Typical examples of ESPI fringes showing the out-of-plane deformation during cooling down process
Fig. 5 Typical examples of ESPI fringes showing the in-plane deformation during cooling down process
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Fig. 6
x direction displacement (u)
Fig. 7
Thermal expansion coefficient
The thermal expansion coefficient of the specimen in various situation are examined as an application. Figure 6 shows the fringe pattern for the case when rectangular shaped specimen was cooled down by the thermal conduction through the cold end. The fringe pattern shows the x direction displacement field (u). The strains were obtained from these displacement fields. The thermal expansion coefficient was obtained based on the strain and temperature data. Figure 7 shows the thermal expansion coefficient for rolling direction of the specimen. In Fig. 7, it is seen that the value vary widely. The complex deflection of the specimen may be caused by the temperature difference between the front and back surfaces, the inhomogeneity of the specimen, and the fixing method of the specimen. This problem is still open to question. SUMMARY It has been demonstrated that the ESPI method can be applied to the study of the deformation behavior due to temperature decrease to a low temperature environment. The two dimensional out-of-plane deflection, and in-plane displacement, strain and thermal expansion coefficient were measured on several kinds of stainless steel specimens. The present investigation can be summarized as follows" 9 The ESPI method offers a non-contact and temperature-independent method to measure the displacement of materials subject to large temperature changes, as long as the correlation of the speckle patterns are kept. " By using the ESPI method, the mechanical behavior of structural materials due to the thermal contraction were successfully measured during the cooling down process, and the applicability of ESPI methods to the low temperature engineering are experimentally confirmed. ACKNOWLEDGMENTS This work is partly supported by the Science Research Promotion Fund from the Japan Private School Promotion Foundation, by the Kansai University Research Grant and also by Grant-in-Aid for Scientific Research (B) from the Ministry of Education, Science, Sports and Culture, No. 08455047. REFERENCES 1 2 3 4 5
Stetson, K. A., The use of heterodyne speckle photogrammetry to measure high temperature strain distributions. Pro_c_: Soc. Ehoto_-Opt.~Instrum. Eng. (1983) 370 46-54 Jiang W., Kawasaki A., Date K., and Watanabe R., Measurement of thermal expansion coefficient of Mo/PSZ sintered composites by laser speckle photography J. Japan Inst. Metals (1994) 58 835-841 (in Japanese) Malmo J. T., Lokberg O. J., and Slettemoen. G. A., Interferometric testing at very high temperatures by TV-holography. Exp. Mech. (1988) 28 315-321 Nakahara S., Maeda Y., Matsumura K., Hisada S., Fujita T., and Sugihara K., Deformation measurements of materials at low temperatures using laser speckle photography method Advances in Cryogenic Engineering (Materials) (1992) 38 85-92 Sirohi S. R., In: Speckle Metro!ogy Marcel Dekker, Inc., New York (1993) 67, 100
Friction and Wear Testing at Cryogenic Temperatures Thomas Gradt, Wolfgang HObner, Helmut B0mer Federal Institute for Materials Research and Testing (BAM) Lab. VIII.12 "Fretting; Cryotribology", 12200 Berlin, FRG In a cryogenic environment componems with interacting surfaces in relative motion (tribosystems) like beatings, seals and valves cannot be lubricated conventionally by using oils or greases. Therefore, they are critical in respect to wear and frictional heat generation. To obtain reliable data of the properties of conventional and advanced materials in cryogenic tribosystems, model wear tests at low temperatures are performed. Tests with polymer-steel couples varified that some polymers have a favourable tribological behaviour at low temperatures. Coatings of amorphous carbon are on principle suitable for cryogenic tribosystems, but their behaviour depends strongly on the composition and deposition method.
INTRODUCTION In a cryogenic environment componems with interacting surfaces in relative motion (tribosystems) like beatings, seals and valves cannot be lubricated conventionally by using oils and greases. Thus, they are critical in respect to wear and frictional heat generation. For these applications solid lubricants like MoS2, hard wear resistant coatings like TiN, amorphous carbon, or material combinations with low friction and wear in dry sliding like polymers vs. steel are used. An overview on early measurements of tribosystems at low temperatures is given by Kragelsky [1]. Polymers like PTFE and Nylon, pure or with fillers are often used in cryogenic tribosystems and their behaviour is reported in several papers [2,3,4]. Kensley and Iwasa [5] investigated the behaviour of several polymers as insulation materials for superconducting wires and Gamulya et. al. [6] studied the behaviour of diamond like carbon at low temperatures. Most of the investigations on cryogenic tribosystems were performed for special applications in space or superconducting sytems. For a systematic study of a broad variety of conventional and advanced materials for cryogenic tribosystems special test devices (cryotribometers) are developed at BAM. A first cryotribometer for gaseous environment is in use since 1993. As a next step an apparatus for cryogenic liquids like LHe, LH2, and LN2 is under construction and will be available in 1997. EXPERIMENTAL SET UP The tribometer for the measurements in gaseous cryogenic environment is shown in Figure 1. The sample chamber is thermally insulated by vacuum superinsulation and cooled by a continuous flow cryostat. The heat exchanger is located in the bottom plate where a temperature between 4.2 K and room temperature can be adjusted by regulating the mass flow of the coolant. The friction tests are performed with samples in a pin-on-disc configuration, consisting of a fixed pin, continuously sliding against the circumference of a rotating disc. The rotation is transmitted via a rotary vacuum feedthrough to a shaft with the sample disc at the lower end. A normal force up to 10 N is applied to the pin by a lever which is loaded by a dead weight. During the measurement the friction force as well as the linear wear are measured by means of a beam force transducer and an inductive displacement sensor. Because of the temperature gradient inside the chamber the lowest achievable temperature at the sample is 8 K. All tests were performed in gaseous He environment with a normal force of 5 N and a sliding speed of 0.2 m/s. Every friction couple was tested at 77 K and room temperature, some also at 35, 10 or 8K. The complete 669
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sliding distance was 1,800 m, divided into three 600 m runs, to study the running-in behaviour. After the experiments the wear scars on both bodies were measured by means of profilometry, light-, electron- and atomic force microscopy.
Figure 1 Cryotribometer for gaseous environment For the polymer measurements pins with a diameter of 2 mm were manufactured by turning and grinding. No additional chemical treatment was performed. As counterbodies discs of austenitic CrNi-steel (AISI type 304) with polished surfaces were used. For the tests of amorphous carbon films the steel discs were coated with commercially available coatings and 100Cr6-steel balls (AISI type 52100) of 10 mm diameter were used as pins. Before inserting the mounted pin-on-disc assembly into the cryostat the sample surfaces were cleaned with ethanol. To remove any condensed liquids on the sample surfaces the sample chamber was evacuated for 24 h to a pressure of 10-3 mbar.
RESULTS Polymers: Most of the tests were performed with polymers against steel. The investigated materials were PTFE, PCTFE, PA 6, POM and PEEK and the results are reported in [7]. Compared with room temperature, these polymers have an improved tribological behaviour in dry sliding at low temperatures. Because the hardness and mechanical strength of many polymers increase with decreasing temperature, the deformation of contact surfaces also decreases, resulting in a smaller real area of contact and thus lower adhesion between the surfaces. As an example, a comparison between the frictional behaviour of PA 6 at 8 K, 77 K and room temperature is shown in Figure 3. In this case, a complete change in the wear mechanism occurs. The change in friction coefficient from about 0.14 at room temperature to about 0.05 at 8 K is accompanied by a change in the abrasivity of the components. While at room temperature the wear occurred only at the polymer pin, at 8 K also the steel surface is worn by the polymer. Transfer of steel particles to the polymer, identified by EDX, confirms this effect.
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Figure 3 Variation of the friction coefficients ofPA 6 at different temperatures Carbon films Carbon comings of the Me-C:H-type were tested at 10, 35, 77 K and room temperature. The comings of one supplier worked with high reliability and reproducibility. The typical behaviour of such a coming in sliding friction is shown in Figure 4. At the beginning of the sliding distance the friction coefficiem passes a short minimum with values lower than 0.05, followed by a maximum of about 0.7 that had nearly the same magnitude for all measuremems. Compared with room temperature, the duration of this high friction state was shorter at 77 K and longer 35 K. After passing this maximum the friction coefficient decreases again and reaches a steady state value between 0.05 and 0.3. This steady state friction was higher at 77 K than at room temperature but lower again at 35 K. For both effects the influence of the temperature is evidem but the tendency is not clear. Because the high friction is accompanied by high wear, the coming at 35 K showed an increasing friction after sliding distance of 850 m which indicates coating failure. In all cases the wear at 77 and 35 K was higher than at room temperature. Only at 10 K large deviations between different experimental runs occurred. In one experimem the friction force decreased to a low value at the beginning and did not reach the maximum within the sliding distance of 1800 m. However, a second experiment with a new track on the same sample disc showed the usual behaviour with a maximum as well as a steady state value. Up to now, only the carbon comings of one supplier were suitable for low temperatures. Other comings failed at temperatures of 77 K and lower after a sliding distance of only a few meters. However, the existance of suitable types of comings is verified and the very low friction at the beginning of the sliding distance indicates that carbon comings are promising materials for cryogenic tribosystems.
CONCLUSION The results of the tribological model tests of polymers show that the basic wear mechanisms of polymers, deformation and adhesion, are still acting at low temperatures but the increasing mechanical strength give rise to a reduced wear. Tests with hard carbon coatings show that they are appropriate for tribosystems in cryogenic environmem. Temperature dependem low and high friction states show that further investigations on this material are necessary.
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Figure 4 Variation of the friction coefficient of a carbon coating at T = 10 K; a) typical running in behaviour; b) low friction state at the beginning of running in ACKNOWLEDGEMENTS The authors would like to thank Mr. R. D6ring, Mr. O. Bemdes and Mr. M. Heidrich for preparation and performance of the experiments. The investigations on the carbon coatings were supported by the Deutsche Forschungsgemeinschafl, Project-No.: Gr 1002/3-1. The samples were DYNAMANT-coatings provided by IKOS GmbH, HOsbach/Germany. REFERENCES 1 2 3 4 5 6 7
Kragelsky, I.V., Friction at low temperature. Friction Wear Lubrication- Tribology Handbook Pergamon, 75-91 Gardos, M.N., Self lubricating composites for extreme environmental conditions In: Friction and Wear of Polymer Composites Composite Materials Series 1,397-447 Wisander, D.W., Maley, C.E., Johnson, R.L., Wear and friction of filled polytetrafluorethylene compositions in liquid nitrogen. ASLE Trans. (1959) 2 58-66 Michael, P.C., Rabinowicz, E., Iwasa, Y., Friction and wear of polymeric materials at 293, 77 and 4,2 K Cryogenics (1991) 31 695-704 Kensley, R.S., Iwasa, Y., Frictional Properties of Metal Insulator Surfaces at Cryogenic Temperatures Cryogenics (1980) 20 25-36 Gamulya, G.D., Ostrovskaya, Ye.L., Ostapenko, I.L., Presnyakova, G.N., Strel'nitzkij, V.E., Friction behaviour and wear resistance of diamond-like carbon films under cryogenic temperatures Diamond and Related Materials (1994) 3 1381 - 1384 HObner, W., Gradt, T., B6mer, H., DOting, R., Tieftemperatur-Reibverhalten von PolymerWerkstoffen Tribologie + Schmierungstechnik (1995)42 244-251
A Flexible and Extensible Monitoring System for Large Scale Experiments Joji Kariya a Haruhiko Okumura b, Masahiko Emoto ~, Mamoru Shoji g, Yasuaki Teramachi d, Tokio Ohska ", Yuriko Shamoto f, Satarou Yamaguchi g, Osamu Motojima g, Junya Yamamoto g a Integrated Information Processing Center, Yamaguchi University, 2557 Tokiwadai, Ube-shi, 755 Japan b Faculty of Politics and Economics, Matsusaka University, 1846 Kubo, Matsusaka-shi, 515 Japan System & Application Eng. Div., Nihon Sun Microsystems K.K., SBS Tower, Setagaya-ku 158, Japan d Department of Information and Computer Engineering, University of Industrial Technology, 4-1-1 Hashimotodai, Sagamihara-shi, 229 Japan ~Physics Division, National Lab. for High Energy Physics, 1-10ho, Tsukuba-shi, 305 Japan f 22-2 Tuchikame, Satou-cho, Toyohashi-shi, 440 Japan g Hontai-toh, National Institute for Fusion Science, 322-60roshi, Toki-shi, 509-52 Japan
We constructed a flexible and extensible UNiX-based distributed monitoring system, consisting of data acquisition, distribution and monitoring subsystems. The acquisition subsystems transmit signals from physical targets by network multicasting. These signals and derived quantities (such as flow rate) are reconstructed on each network-connected workstation to form "'virtual targets of observation", from which the monitoring subsystems obtain quantities of physical interest. When separated from acquisition and distribution subsystems, monitoring subsystems can be easily designed. The monitoring performance can be arbitrarily increased by increasing the number of monitoring workstations in the network.
INTRODUCTION The Large Helical Device (LHD), under construction at the National Institute for Fusion Science (NIFS), Japan, will be equipped with the largest superconducting coils in the world, with over 2,000 sensors to monitor cooling processes and to enable early detection of quenches. The sampling intervals range from seconds (for slowly changing quantities such as temperatures) to milliseconds (for relatively fast quantities such as tap voltages), and even to microseconds (e.g. for acoustic emissions). Experiments are now being carried out on each constituent coil with about a hundred sensors. Therefore, the monitoring system must be designed to adapt to vast ranges of numbers of channels and sampling rates. Network technologies and techniques for digital processing of analog data, as in multimedia, have greatly advanced recently. Applying these technologies, and digitizing measured data near the sensors, we could make monitoring systems more extensible, flexible, fault-tolerant, and easy to develop. To this end, we investigated how to cope with large amounts of data while still guaranteeing realtime response. This depends on data transmission modes, computer performance, and the process control scheme of the operating system. To investigate the feasibility of such a system on standard workstation-based systems, we experimented a UNIX-based distributed system on Ethemet. Our conclusion is that it is possible to use such methods on the LHD. PHYSICAL SYSTEM CONFIGURATION Figure 1 shows our physical system configuration. Workstations that comprise the distributed system are positioned around three segments of local-area networks (LANs). Each workstation and network 673
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segment has its own role: Segment R is used for realtime data, segment C for control data, and segment X for external services. The firewall prevents unauthorized communications to enter the inner region. segment(C) ~ Workstation(c)' I -{Workstation(a) [----{Workstation(s)] IsensorH -npH xzl ! -{ Workstation(a) [segment(X) segment(R) ........ ~"t Workstation( e ) ] ' ~ --[Workstation(d) ~ firewall[--4 ' ]--~Workstadon(e)J<~ -{Workstation(d) [- '" I
Figure 1 Physical System Configuration Sensors and other analog devices are exposed to temperatures in the range from 4.2K to room temperatures, high magnetic fields and electric currents. To minimize the effects of such a difficult environment, we use insulation amplifiers [1 ]. Signals from the sensors are amplified by insulation amplifiers, and digitized by analog-to-digital converters (ADCs). The acquisition workstations ((a) of Fig. 1) acquire digitized realtime data, and send them via segment R by multicasting to other workstations. The number of these other workstations can be increased arbitrarily. Workstations (d) display the received data either temporally or spatially, so as to utilize humans' high pattern-recognition ability. Automatic warning is possible, provided the algorithm to warn is given. Workstation (c) maintains a database of basic information such as positions and attributes of sensors and amplifiers, as well as control information for the experiments. It works as the information server and controls the total system. The acquisition workstations (a) not only send slow realtime data, but also summarize fast data into slow data, detect interesting events from fast data, store fast data intermittently to the disks. After data compression, the stored data are sent to Workstation (s) for long-term storage. Stored data are used for later analysis. Workstations (e), outside the firewall, receive data from inner workstations, and service them to the outside world on realtime basis. They send data directly by TCP (the Transmission Control Protocol of the Internet) to distant co-workers, and by access-restricted World-Wide Web (WWW) servers to other workers. LOGICAL SYSTEM CONFIGURATION Our system must be applied to vast (two orders of magnitude) range of sizes of experiments, from experiments on short specimen superconductors to the monitoring of the whole LHD system. Therefore it must be extensible. Moreover, the decision principles that lead to efficient monitoring are not known in advance, so it must be so flexible as to do various observations concurrently and to allow observational methods to be changed in the midst of an experiment. As another requirement of flexibility, the development of the monitoring programs must be easy. On top of the distributed processing system shown in Fig. 1, we constructed the logical system as represented in Fig. 2. Physically measured data are sent to all workstations by multicasting, and compose what we call "virtual targets of observation". Monitoring processes can acquire physical quantities directly from these virtual targets. The virtual targets hold data within a specified time span. They also hold quantities derived fi'om multiple data, such as rates of change and flows rate. Quantities can be derived from combined data originating from multiple ADCs. Monitoring subsystems can thus be designed independently of the acquisition subsystems physical configuration. It is also independent of the configuration and electrical conditions of the measuring devices of the experiment. Monitoring programs can be designed easily, because physical quantities can be obtained by first specifying the name of the virtual target of observation, and then by specifying the name of the observed data. Because data are distributed by multicasting (i.e. a data packet is received by several workstations at the same time), network traffic is independent of the number of workstations that receive the data. The monitoring capacity can be easily extended by connecting workstations to the network, constructing virtual targets of observation within each workstation and invoking monitoring processes, without affecting other system elements. Any number of processes measuring the same virtual target of observation can be independently started and stopped.
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monitoring
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[monitoring]process I
l=a l Pr~=t g l/ c "~ process /
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Figure 2 Logical System Configuration Thus the key to the system cxtensibility and flexibility is that the virtual targets of observation can be arbitrarily constructed, independent of physical observation. Reconfiguration of acquisition subsystems, however, is difficult during experiment. Acquisition configuration must be carefully designed so that each sensor is sampled at sufficient rate. Since the acquisition configuration is determined by the characteristics of physical targets and sensors, it should be easy to accomplish this. Figure 2 also shows how several channels of data, possibly from different workstations, may be combined to form secondary data. For example, flow rate is derived from fluid temperature and differential pressure across an orifice. REALTIME DATA DISTRIBUTION Proper network functioning is the key to the system performance. Three segments shown in Fig. 1 must be clearly separated: Steady data multicasting must go through segment R, random bulk file transfer and control signals must go through segment C by reliable TCP protocol, and external access must be allowed only on segment X. We use 10 Mbps optical Ethemet and multicasting for realtime data transfer. In Fig. 2, a measurement process sends only one data packet, irrespective of the number of virtual targets of observation. The system is extensible in that increasing the number of monitoring subsystems does not affect other subsystems. Since realtime data is transmitted sequentially, retransmission requests can affect data stream. Transmission rate must be preset so that no transmission errors should occur without retransmission. Since receiving performances differ among different workstations and different loads, maximum possible load must be determined by experimentation. We.have tested Sun Microsystem's SPARCstation 2 and SPARCstation IPX with Solaris 2.5. We multicasted 512-channel packets (1064 bytes including headers) at regular intervals. At the receiving end, we compared realtime (RT) and time-sharing (TS) dispatch classes available under the Solaris operating o S: send total: 1(300000 packets system. Figure 3 shows the histogram of packet arrival ,,R: reaJtime class intervals when transmission is done at 4-millisecond lost: 0 packet intervals (250Hz), a 20% load for the Ethemet. o~I~ T: time sharing class , "lost: 1649 packets Variation in intervals is very small for RT dispatch, whereas TS dispatch shows large variation and some packet losses. The packet losses tend to occur in runs: the 1649 lost packets (0.16%) comprise 401 runs, the 8 longest being 19-packet long. It is clear that even for low traffic we should use realtime dispatch class. Table 1 shows the standard deviations of RT-class packet arrival intervals for various transmission intervals. We tested 100,000 packets for each case, but no packet loss occurred. Dispersion increases, however, when transmission interval was 3 4 5 6 below 1.5 milliseconds (50% load). Thus, we interval(rnsec) recommend 50% traffic load for the sake of safety. Figure 3 Interval distribution of packets 1
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APPLICATION TO THE EXPERIMENTS Figure 4 shows a realtime graph tool we developed. The data shown here were taken during a two-month experiment [2,3] on an IV-coil (inner vertical coil of the LHD) in February, 1995. The abscissa is time (in HH:MM:SS format), extending to 200 minutes. The ordinates show the temperatures on various parts of the IV coil, measured with CGR, RuO, and Cemox sensors, transformed to Kelvins by interpolating calibration tables with cubic splines. While monitored with such realtime graph tools, the data were automatically saved to disks. They were also converted to hypertexts and GIF (Graphics Interchange Format) graphics on the fly to enable browsing with popular Intemet World-Wide Web browsers. Time stamps were put on stored data by system clocks calibrated by Network Time Protocol to allow accurate synchronization. Data graphs can be replayed at any speed by reproducing network transmission from saved data. This feature is indispensable for such large-scale experiments as LHD that cannot be repeated easily.
send interval
2500usec 2000usec 1500usec 1200usec lO00usec
traffic load 35% 44 % 59% 73% 88%
dispersion of intervals 43.8usec 42.8usec 58.2usec 62.8usec 139.9usec
packet loss 0% 0% 0% 0% 0%
Table 1 Dispersion of arrival intervals
CONCLUSION We demonstrated that extensible and versatile UNiX-based distributed monitoring system can be realized by using multicasting to distribute realtime data. Multichannel realtime monitoring at a speed of 250k samples (500 packets) per seconds can be easily obtained; 2.5M samples per second is within reach Figure 4 Scrolling Graphs of Measured by using suitable machines and networks. Beginning with small-scale experiments, we have been investigating such issues as extensibility. Also, since this field has been undergoing so much technological innovations that a system can be out of date when completed, we investigated trends of new technology, and aimed at the construction of a state-of-the-art system. We are now investigating the effects of computing load on realtime data acquisition, and maximum load for high-speed networks (100Mbps or higher). Another task is to allow external triggers to start/stop taking multichannel data at 5M to 50M samples per seconds and save them to disks. Such high-speed data currently defy realtime disk storage. REFERENCES 1 0 k u m u r a , H., et. al., A scalable data acquisition system for superconducting coil experiment, Proc. of the 18th Symposium on Fusion Technology(1994)835-838 2 Satow, T. et al., Cooling and excitation tests of an inner vertical coil for the Large Helical Device, MT-14 Tampere, Finland, June 11-16, 1995. 3 Satow, T. et al., Cooling and excitation experiments of a single inner vertical coil (EXSIV) for the large helical device, ICEC 16/ICMC, Japan, May 20-24, 1996.
Development of a Low Temperature Measurement and Control System for Measuring Liquid Oxygen Density
Dun Wang*, Zhixiu Huang*, Guobang Chen*, Jianyao Zheng*, Jianping Yu*, Yayu Li**, Min Shen**, Hongyan Chen* *, Ruiliang Xu* *, Guang Wen Cui* * and Ruimin Liu* * *Cryogenics Laboratory, Zhejiang University, Hangzhou 310027, P.R.China ** 101 Institute, Space General Industrial of China, Beijing 100074, P.R.China
A low temperature measurement and control system which mainly contains microprocessor has been developed for measuring the density of liquid oxygen. Much more software instead of hardware has been used in the system to realize different kind of functions, for example, calibration of A/D conversion, sensor's data process, PID regulation and the self-modification of its parameters, PWM method to complete the precision control of the heater, etc. The temperature sensor is interchangeable and remains its original measurement accuracy. The systems is operated in the form of menu on the screen. A simulation test of the system shows that the measuring accuracy is ___0.1K at temperature range of 90 to 112 K.
INTRODUCTION In order to meet the requirement of high efficiency of LH2/LO2 rocket engine, a project fi~r direct measurement of liquid oxygen density under different pressures is put forward. A schematic of th.~ system of liquid oxygen density measurement is shown in Fig.1. The system consists of cryostat, e ectronic balance, temperature controller, pressure controller, vacuum pump, as well as valves and tu{,es. The electronic balance is set at the top of the cryostat. The cryostat is a special Dewar containing set~sors for temperature, pressure, liquid level and heater except an elevator which holds a crystal ball in the bracket or separate the ball from the bracket. The pressure controller is of load type, which can be used to adjust pressures from 0.1MPa to 0.6MPa and as a safety valve. Before liquid oxygen filling into the cryostat, the system should be cleaned by the vacuum pump. The demand of temperature range for liquid oxygen densimeter is from 90K to 112 K, corresponding to the saturated pressures of 0.1MPa to 0.6MPa. The temperature must be controlled withir,~ _+_0.1K during density measurement. For this purpose, a cryostat, including temperature and pressure control system has been developed. In this paper, the temperature measurement and control system J~ mainly presented. Due to the need of high accuracy of temperature control, it is impossible to satisfy th(, function with ordinary measuring instruments. Therefore, a temperature controller which mainly contains microprocessor has been developed and manufactured. HARDWARE FOR TEMPERATURE CONTROLLER As shown in Fig.2, the hardware consists of analog, digit and interface unit [ l ~ 3 ] . T h e main function of 677
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Fig. 1 Liquid oxygen density measurement system the analog part is to complete the conversion and amplification of the input signals, for ~,xample, temperature, pressure, liquid level, A/D conversion, and then enter the CPU (Central Processi~:g Unit). According to the measurement data and predetermined P1D (proportional-integral-derivatix~:), CPU regulates the heater power with help of sot~ware PWM (Pulse Wide Modulate). Therefore, the temperature and pressure are automatically controlled. Mathematical process for signals and sonte logical functions is achieved by digit unit which consists of logical circuit centering in CPU. Communication unit accomplishes level conversation for the purpose of communication between the controller and computer
Fig.2 Interior block diagram of controller In order to raise maintainability of the constructions. The units of A/D, communication, modular construction corresponds with its circuit plates, for example, A/D etc. can be inserted in the
controller, its circuit is divided into several modular supply source, display and CPU are modularized Each plate while CPU is the main plate among them. Other main plate or connected with it through flat wire. h will
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not affect the integrity of the controller if one of the modular constructions is taken from the who! e circuit. The arrangement makes it possible to reduce interaction between them. The inter hardware layo~t of the controller is shown in Fig.2.
SOFTWARE FOR TEMPERATURE CONTROLLER Some software instead of the hardware is designed to increase operation reliability and decrease manufacturing cost of the system. Calibration of A/D conversion It is necessary to calibrate the result of A/D conversion to increase its precision due to imperative nonlinear and quantization errors. Linear insertion is adopted. The analog signal input terminal of the A/D conversion connects with the terminal of a high precision digital voltmeter (at least 5 1/2 digit display). Changing input voltage, a set of data is obtained and recorded in x and y coordinate, respectively At last, the controller will receive the complete diagram which is used for automotive calibration of the A/D conversion results. The precision of the MD conversion is most approximate to the one of digital voltmeter.
1 DISPLAY UNIT INTERFACE 3 CPU
2 K E Y B O A R D INFEI@'ACE
4 C O M M U N I C A T I O N INSERT SLOT
5 A/D BOARD INSERT SLOT
6 A/K B O A R D INSli'..ad SI,()T
7 INTERFACE DRIVE ZONE
8 SUPPLY SOURCE Id,,ITERt:AC],;
9 C O N N E C T I O N TERMINAL
Fig.3 Interior layout of controller Data processing of platinum thermometer Voltage drops across the platinum resistance which is passed by a constant current are transferred to an amplifier, then by the way of a multi-switch to the A/D conversion. The analogue signal is changed into digital one. PID regulation and self-modification of parameters General speaking, the parameters of proportional factor, derivative time and integrating time are determined by mathematical model, experiment or experience [4]. These methods are rather inconvenient and can not be acquired for optimum parameters. But the self-modification of paramel.ers can automatically obtain the approximate optimum ones. These parameters exist individual optimum value and interaction among them. But for a given control system an optimum matching relationship can be determined with quadrate trial [5]. The trial is based on array-combination theory with quadrate table as a means to arrange experiment and analysis results. The self-modification parameters store in the inner storage of the controller which is needed for PID regulation. Pulse wide modulate (PWM) Heating time is controlled by the control way of electrical pulse wide modulate. Average heating power is changed from 0 to heating power P o . The step is short enough (1/255) to be considered as quasi-
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continuous control. In fact the working period of the heater is about 1 second, a precision of 0.05K of the temperature control is expected. 1..nterchangeability of sensor and A/D conversion Software correction (comparison) interface program is designed in the software for every sensor and A/D. Each correction curve is independent. It can be revised when it is needed.
SIMULATION AND PRELIMINARY TEST FOR THE TEMPERATURE CONTROLLER The temperature measurement is based on the measuring resistor of the platinum. Now a set of 4 parallel connection resistance boxes with high precision and stability is considered as a temperature sen,,or. Each box has a resolution of 0.1 ~ We can see from the theoretical calculation that a resistance reso:ution of 0.025 f~ corresponds to a temperature resolution of 0.05K. Experimental study shows that the controller can measure a temperature of better than 0.05K. And under the simulation state, the software and hardware of the whole control system work in good order and fulfil the design requirements. Preliminary test for the temperature controller is made in a cryostat as shown in Fig. 1. The cryostat is filled with liquid nitrogen to a designated liquid level. Two platinum resistance sensors with an accuracy of 0.05K is installed near the densimeter. The sensors have been calibrated by cryogenic metrolog)~ station, Chinese Academy of Sciences. As soon as a temperature range of 78.00K to 78.10K is set, the controller can automatically search out an approximate optimum parameter of PID by self-modification and regulate temperature by the way of electrical pulse wide modulate. After the temperature is reached to the control point. The digital display of the control shows a temperature reading with a resolution of 0.01K. The stability of the control temperature is 0.05K within 5 minutes which is needed for density mea~,.urement. The controller will be used in liquid oxygen density in near future.
CONCLUSIONS 1. Simulation and preliminary test shows the accuracy of the controller can satisfy the de,hand of measuring liquid oxygen density due to adopting single chip processor. 2. Much more software instead of hardware is used to save cost of the controller and increase its reliaFility. 3. Temperature sensor is interchangeable and remains its original measurement accuracy.
REFERENCES Yucai S., MCS 51 series single chip processor and its application Nanjing college of tec:hnology publishing house (1987) 11-75 Jianhua J. and Qingqing W., Rapid data communication of I B M ~ P C and MCS--51 sil~gle chip processor Selected compilation of the application of single chip processor (1992) Shijiang X., Serial communication of MCSmS1 single chip processor and I B M - - P C Selected compilation of the application of single chip processor (1992), No2 Ming L. and Huaixin L., Parameters modification of PID regulation used for automatic control Measurement and control techniques 13 (1994), No 6, 30-31 Dan S., Quadrate trial with computer aided design Cryogenics and special gases (1994)
Investigation of Thermal and Vacuum Transients on the LHC Prototype Magnet String
P. Cruikshank, N. Kos, G. Riddone and L. Tavian LHC Division, CERN, CH-1211 Geneva, Switzerland
The prototype magnet string, described in a companion paper, is a full-scale working model of a 50-m length of the future Large Hadron Collider (LHC), CERN's new accelerator project, which will use high-field superconducting magnets operating below 2 K in superfluid helium. As such, it provides an excellent test bed for practising standard operating modes of LHC insulation vacuum and cryogenics, as well as for experimentally assessing accidental behaviour and failure modes, and thus verifying design calculations. We present experimental investigation of insulation vacuum pumpdown, magnet forced-flow cooldown and warmup, and evolution of residual vacuum pressures and temperatures in natural warmup, as well as catastrophic loss of insulation vacuum. In all these transient modes, experimental results are compared with simulated behaviour, using a non-linear, one-dimensional thermal model of the magnet string.
INTRODUCTION The Large Hadron Collider (LHC) project [ 1], currently under design at CERN, will make use of high field superconducting magnets operating below 1.9 K in a pressurised bath of helium II. Superconducting magnet cold masses are housed in horizontal cryostats [2] with several layers of heat interception and screening. Prototype full-scale models [3] have been built, tested and assembled into a test String [4,5] in order to validate nominal and accidental operational modes. The superinsulation system is composed of an intermediate shielding at 50-75 K and a radiative insulation at 4.5-20 K. The thermal shield at 50-75 K is covered with 30 layers of multilayer reflective insulation and the cold mass is wrapped with 10 layers of the same superinsulation [6]. Nominal insulation vacuum of 10-4 Pa can be achieved first by insulation vacuum pumpdown and then by cryopumping during cool-down. Pumpdown experiments on the String have been made to identify times, the effect of conditioning of the multilayer reflective insulation (MLI) and the acceptable initial pressure for cooldown. Thermal transients represent standard operation modes which have to be studied in order to be able to assess maximum thermal gradients in the magnet during fast cooldown and warmup. Interventions on the LHC machine may require forced-flow warming of only a few magnets. It is of interest to know the time required for warming up, pumping down and cooling down for nominal operation and to understand how the reminder of the machine will behave in a passive state. Another transient mode is produced by accidental loss of insulation vacuum. It may occur because of an air leak from the ambient surroundings or an hntemal helium leak from a cryogenic circuit. The vacuum vessel for the LHC dipole cryostat is made of ISO 430 carbon steel and in case of loss of insulation vacuum it will progressively cool down, but must not reach the embrittlement temperature of the material.
INSULATION VACUUM PUMPDOWN LHC cryostats must be pre-evacuated, to achieve both rapid and efficient cooldown. The insulation vacuum space of the String is characterised by the 10 m 3 volume and the surface outgassing from 104 m2 of MLI. First pumpdown of the insulation vacuum shows high water vapour gas load from the MLI surface. 681
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Following a number of days pumping and then venting to dry nitrogen gas, pumpdown times are gready reduced. The MLI has been vacuum conditioned. However, experience from the String has shown that prolonged exposure to ambient air during a 3 month shutdown, returned the system to initial conditions. Pumping of vacuum enclosures is characterised by a volumetric and outgassing component. On String pumpdowns where the system has been exposed to air, the outgassing component dominates after a few hours pumping, between 10/ and 10 3 Pa. In the case of conditioning and repumping, the outgassing component dominates in the region 1 to 10 Pa. For the two scenarios, a pressure of 1 Pa was achieved in approximately 30 and 3 hours respectively. A series of experiments have been performed to investigate the upper pressure limit to begin cooldown, and show a pressure of about 1 Pa as acceptable. This value is anyway a requirement for room temperature leak testing of the vacuum enclosure using helium mass spectrometer technology.
FORCED-FLOW COOLDOWN AND WARMUP Cooling down from room temperature to 80 K and warming up from 80 K to 300 K are achieved by flowing gaseous helium respectively at decreasing and increasing temperatures. The inlet helium temperature is regulated on the total string gradient in order to avoid thermal stresses in the magnets. The evolution of magnet and helium temperatures have been simulated by a one-dimensional non-linear computer model, shown schematically in figure 1. The heat transfer coefficient (h) between helium and magnets has been estimated at 200 W/K.m from design characteristics and validated from experimental data. The mass and the specific heat of the cold mass (CpM) have been calculated from magnet design and the helium specific heat (CpHo) from property tables [7]. Helium flows through several channels connected in parallel and only those which see turbulent flow, contribute to the heat exchange. The cold mass laminated structure allows longitudinal heat conduction to be neglected. Following this assumption only convective heat exchange contributes to the heat transfer between cold mass and helium. From the heat balance equation we obtain a system of partial differential equations which can be solved explicitly using the finite difference method:
T~ ,t+l) _W~ ,t)
M .CPM 9
At
= - h . (TM (x,t) -THe (x't))
IT(~: 1,t) - "I'~(~t) 1 T~gt+l'- m(HXgt'1 h (Ti(X,t) THe(X,t)) rh. Cp lie " AX v At .~
Boundary conditions:
__.
o
---
o
--
1) at any location, TM(t=0)=constant 2) at any time, Tm(x=0)=constant
At each node, we calculate cold mass temperature (TM) and helium temperature (Tin) as a function of time (t).
Figure 1" Mathematical model scheme forced-flow string cooldown and warmup
of
Figure 2: Forced-flow cooldown prototype String (300-80 K)
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Table 1 fists the value of the cold mass M, the helium mass-flow rh and the maximum temperature difference (AT~x) between helium and magnet during the different measurement runs. The time needed for cooldown and warmup strongly depends on the imposed maximum driving temperature difference between helium and cold mass as well as on the mass-flow. For the first three runs, string cooldown started once the insulation vacuum was around 10 2 Pa whereas during Run2B residual gas pressure was maintained at about 1 Pa. Residual gas pressure of 1 Pa did not affect the cooldown time because the added heat inleak is negligible in comparison with the high heat capacity of the cold mass. Figure 2 shows the forced-flow cooldown of the LHC prototype magnet String (Run2A) and compares simulated and measured data. Table 1" Main parameters of the string forced-flow cooldown and warmup
Cooldown Run0 Runl Run2A Run2B
M [kg] 45.103
45.103 65.103 65.103
ria [g/s] 50 50 60 80
AT~ [K] 50 60 60 60
Time 300-80 K [h] 85 70 85 70
Warmup ria AT~x Time 50-300 K [g/s] [K] [h] 50 50 100 80 60 55 natural and accelerated warmup
NATURAL AND ACCELERATED WARMUP Figure 3 shows the thermal flow-scheme used for the one-dimensional radial model simulating heat transfer during warmup [3]. Convection in nitrogen gas is negligible in the case of natural warmup, whereas it is relevant in case of accelerated warmup. Solid conduction through the insulation spacers and radiation between the layers of superinsulation are negligible in both the cases. Convection cannot occur in the superinsulation due to the restrict space between spacers. Natural warmup without active pumping on the insulation vacuum started after a quench when the cold mass temperature was around 30 K and the thermal shield at 90 K. After about 20 hours of natural warmup residual gas pressure in the insulation vacuum degraded rapidly to 1 Pa. The end composition at 50 K was mainly composed of hydrogen (10 %) and carbon monoxide (25 %) from the superinsulation outgassing and of nitrogen (40 %) and oxygen (10 %) from a known air leak. The gas species at this pressure is not relevant since the thermal impedance given by radiation is at least a factor 5 higher than that due to conduction in residual gas. The test was stopped after 6 days when the cold mass was at 90 K and the thermal shield at 180 K In order to simulate an accidental loss of vacuum insulation, warmup was then accelerated by injecting N2 in the insulation space of the cryostat. Atmospheric pressure inside the cryostat was reached after 30 minutes. After another 30 minutes, condensation followed by frost was observed in distinct cold spots on the external surface of the cryostat beneath the lower end of each dipole. They could be attributed to the longitudinal heat transfer by natural convection. The lowest of these cold spots was at 230 K. In order to prevent embrittlement of the carbon steel wall of the vacuum vessel the insulation vacuum was pumped down to 5.104 Pa. The String temperatures were allowed to evolve at this pressure for 2 weeks until they reached 300 K. Figure 4 shows the evolution of the String average temperatures during natural (1 Pa) and accelerated warmup (5.104 Pa) and compares experimental and simulated data. Thermal impedance due to conduction in residual gas is not negligible and this underlines the important role of superinsulation in case of an accidental loss of insulation vacuum. The maximum calculated heat flux from the thermal shield to the cold mass was 300 W/m and it was reached after a few hours. In nominal conditions the heat inleak at 1.9 K is about 0.4 W/m and during a magnetic resistive transition is of the order of 100 kW/m.
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Figure 3" Flow scheme of natural and accelerated warmup
Figure 4: LHC prototype String temperatures during natural and accelerated warmup
CONCLUSIONS Thermal and vacuum transient modes in a full scale LHC prototype string have been performed successfully and confirmed the basic design choices. It has been showed that it is possible to cooldown or warmup the 65.103 kg of the LHC string cold mass in less than 4 days with 60 g/s of gaseous helium and a maximum longitudinal thermal gradient in the magnet of 60 K. Simulation of loss of vacuum insulation by nitrogen gas injection at 5.104 Pa is not considered catastrophic and the vacuum vessel minimum temperature of 230 K is still within the limit accepted for carbon steel embrittlement. The power received by the cold mass is small in comparison with that dissipated after a magnetic resistive transition. Mathematical models developed for studying forced cooldown and warmup, natural warmup and accidental loss of insulation have been checked and validated against experimental data. These simple models can be used to predict the behaviour of the LHC machine cryostats under nominal and accidental conditions.
REFERENCES 1 2 3 4
5
6 7
Evans, L.R., The Large Hadron Collider project, paper presented at this conference The LHC Study Group, The Large Hadron Collider Conceptual Design, Report CERN/AC/95-05 (LHC) (1995) Lebrun, Ph., Szeless, B., Tavian, L., Williams, L.R., Experimental investigation of accidental loss of insulation vacuum in an LHC prototype dipole cryostat, paper presented at CEC'95, Columbus (1995) Bdzaguet, A., Casas-Cubillos, J., Flemsaeter, B., Gaillard-Grenadier, B., Goiffon, Th., Guinaudeau, H., Lebrun, Ph., Marquet, M., Serio, L., Suraci, A., Tavian, L. and Van Weelderen, R., The superfluid helium cryogenic system for the LHC test String: design, construction and first operation, paper presented at CEC'95, Columbus (1995) Bdzaguet, A., Casas-Cubillos, J., Guinaudeau, H., Lebrun, Ph., Serio, L., Suraci, A. and Van Weelderen, R., Cryogenic operation and testing of the extended LHC prototype magnet String, paper presented at this conference Benda, V., Dufay, L., Ferlin, G., Lebrun, Ph., Rieubland, J. M., Riddone, G., Szeless, B., Tavian, L. and Williams, L., Measurement and analysis of thermal performance of LHC prototype dipole cryostats, paper presented at CEC'95, Columbus (1995) HEPAK, Version 3.21 by Cryodata Inc., Niwot, Colorado
Innovative Device Producing Double-Layer Cryogenic Pellet Hajime Itoh*, Shigeru Sudo *Nippon Sanso Corp., 6-2 Kojima-cho, Kawasaki-ku, Kawasaki-city, 210 Japan National Institute for Fusion Science, Furo-cho, Chikusa-ku, Nagoya, 464-01 Japan
An innovative device for producing a double-layer pellet has been developed for the purpose of establishment of an accurate transport diagnostic system to measure particle transport both parallel and perpendicular to the magnetic field lines of magnetic confinement devices. The tracer particles are deposited by a doublelayer pellet which consists of a small core as tracer of light atom such as lithium and the major outer layer of hydrogen isotope. For technical demonstration of the device to produce the double-layer pellet, we chose the cylindrical form of the pellet: the outer diameter of 3mm, the length of 3mm with the core diameter of 0.24mm.
INTRODUCTION The double-layer pellet has been expected to be ideal means for the efficient fueling of a specific species as tritium in the core region, and/or controlled deposition of specific minority ions in the plasma core region for efficient ICRF heating, and/or establishment of an accurate transport diagnostic system to measure particle transport both parallel and perpendicular to the magnetic confinement devices[l]. The idea of this double-layer pellet is well known in the field of fusion plasma experiment. The required structure of the pellet is the same in principle; the outer layer of the pellet is made of solid hydrogen isotope which is the same species as bulk plasma ions and, in the core, the inner layer, which is solid tritium for fueling, liquid helium for the deposition of minority ions, and solid material such as lithium, beryllium, and carbon, is installed. Due to the difficulty of producing this type of pellet, there have not been so far such device in the world. To develop the technique to produce the double-layer pellet, we have studied and designed the device especially suitable for transport diagnostics. The pellet size and structure we have chosen is the cylinder made of solid hydrogen or deuterium with its diameter of 3mm and its length of 3mm which contains a carbon sphere of the diameter of 0.24mm. The size of this pellet is adequate for the Large Helical Device (LHD) which is under construction in National Institute for Fusion Science (NIFS)[2]. DIAGNOSTIC PRINCIPLE When the double-layer pellet is injected into a plasma, only outer layer begins to be ablated in the plasma edge region. At a certain position, after the outer layer is completely ablated, the core material will start to be ablated. In order to deposit the core material in the localized area, the size and the pellet velocity have to be adjusted according to plasma parameters such as electron temperature. One of the methods for observing the behavior of the tracer particles is charge exchange recombination spectroscopy. DEVICE FOR PRODUCING THE DOUBLE-LAYER PELLET The important requirement for the double-layer pellet for transport diagnostic is accurate positioning of the core material in the center. The device to demonstrate the technology producing the double-layer pellet has to be designed and constructed to satisfy this requirement. Figure 1 indicates the cross 685
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sectional view of the main part of the device developed this time[3]. It consists of a vacuum chamber, a cryohead, a radiation shield, and driving stages for X-axis and Z-axis. The cryohead made of OFC (Oxygen Free Copper) is the most important part of the device, where the pellet is produced, a core material is positioned in the pellet, and the shape of the pellet is controlled. The cryohead is cooled down with liquid helium. The temperature of the cryohead is controlled around 8K (in case of solid hydrogen pellet) with electrical heaters and temperature sensors installed. As shown in Figure 1, the cryohead is surrounded with the radiation shield cooled by evaporated helium gas. To reduce heat inleak to the cryohead from the room temperature and to make thermal reduction as small as possible, the cryohead is supported by 3 invar supports from the top flange. A high pressure helium gas inlet is installed at the vertical flange and a barrel is installed throgh the other vertical flange at the opposite side. The produced pellet is ejected through the barrel by high pressure helium gas. Production process of the pellet in the cryohead is illustrated in Figure 2. As indicated in Figure 2, we have designed the production sequence in 4 steps, physically from the bottom side to the up side. Inside the cryohead there are two sliding carriers, of which thickness is 1.6mm and 1.4 mm, respectively. Each carrier is made of stainless steel and a hole of ~b 3mm is punched out. In the 1st step the 1st carrier is set at Z1 point where the hole of the 1st carrier is coincident with the hole at the cryohead bottom. Hydrogen gas supplied is frozen and form nearly half size of the required pellet. Then the carrier is pulled up to Z2 position, where a carbon sphere as a tracer core is pushed into the half size pellet by a tungsten wire. Then the 1st carrier is pulled up to Z3 position where the two holes at the 1st and the 2nd carrier coincide. Hydrogen gas is supplied and forms full size pellet. Finally both carriers are pulled up to Z4 position. High pressure helium gas is supplied to eject the pellet. To assure the poisoning accuracy along vertical (Z) axis and horizontal (X) axis, a stepping motor drive is applied for each axis. PELLET PRODUCTION The total system of the device is indicated in Figure 4. To observe and confirm the structure of the pellet, a observation pipe is installed at the barrel outlet. The ejected pellet is photographed by fast flash lamps during the flight. The simultaneous photography from two directions, vertically and horizontally, is applied for the first time in the world. After a number of trials, in mid 1995, we succeeded in production of the double-layer pellet, as shown in Figure 3. This is also the first observation in the world. However it has been found that the optimum condition to produce the double-layer pellet required suitable temperature control and hydrogen gas supply speed. In order to improve the productivity, we have modified the control system to be suitable for automatic control. CONCLUSION An innovative device for producing a double-layer pellet is designed, constructed and operated. The production method applied is based on slide mechanism at cryogenic temperature and is designed to produce a cylinder pellet of q~ 3mm x L3mm with a carbon sphere of q~ 0.24mm in its center. The photograph of the pellet has been taken simultaneously in the two directions, vertically and horizontally to confirm the structure of the pellet. It has been shown that the pellet is a true double-layered having a small carbon sphere inside. This is not only the first achievement of the double-layer pellet production in the world but also the first simultaneous observation from two directions in the world. To increase the reproducibility and to improve the accuracy of the carbon sphere positioning, automatic operation of the device has been studied and now its commissioning is being conducted. REFERENCE 1 2 3
Sudo, S. IN: Symposium on Fusion Engineering, Champaign, ILL. USA (1995) Sudo, S., Diagnostics of particle transport by double-layer pellet, J. Plasma and Fusion Research (1993)69 1349-1361 Itoh, H and Sudo, S, IN: Symposium on Fusion Technology, Karlsruhe, Germany (1994)
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Photo of a pellet
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:Heater :Pressure Gauge P I :Pressure Indlcater PG : P i r a n i Gauge :Rotary VaCUUm Pump :Leak Valve TI2 T e m p e r a t u r e E l e m e n t TMP:Turbo Molecular Pump v : Valve servat ion
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A New Concept to Detect He Leaks on Tore Supra Cryogenic Plant
Minot F., Bon Mardion G., Jager B., Gravil B., Marechal JL., Violet JL. Centre d'Etudes de Cadarache, Association Euratome-CEA Departement de Recherche sur la Fusion Contr616e 13108 Saint paul lez Durance. France
One of the main problems encountered in cryogenic helium plants is leaks. The TORE SUPRA cryogenic system is complex and spread over a large area. In order to provide continuous automatic monitoring of the plant, we have developed a new centralized system able to detect and locate leaks in real time. The system essentially uses a mass spectrometer which continuously analyzes the air helium content. A data storing system provides a posteriori information. This paper presents a comprehensive survey of the equipment developed, along with its performance and sensitivity.
INTRODUCTION The TORE SUPRA Tokamak is presently in operation on the site of the Centre d'Etudes of Cadarache, FRANCE. It is a machine which includes in its technology all the problems linked to the use of superconducting coils cooled at 1.8 K by pressurized superfluid helium. The associated cryogenic system [ 1] provides, on the one hand, the refrigerating powers required for the operation of superconducting magnets, and on the other hand, the liquid helium used by the different teams of the department. In view of its complex nature and of its operating mode in liquefaction, the cryogenic system on average loses 2 10-4 m 3/s of gaseous helium. To locate these leaks, we have designed, built and installed a helium presence detection system in the ambient atmosphere of the buildings which house the helium cryogenic installations. The material used is standard leak detection material which we adapted to our needs. We have measured the sensitivity and the response time of this new system. This new device has been in operation on the cryogenic system for two years. THE DEVICE Princiole The system is a helium detector associated to a network of sampling channels which ensure the surveillance, in industrial environments, of important elements which are likely to leak such as the cold box of the refrigerator, the warm machines or the experimental cryostats. There are two types of surveillance : One near the components, with a measurement channel of which the sampling points can be multiple. - A more global one which consists, in the work areas, in detecting the ambient helium rate. The system is controlled by an instrumentation and control system which provides sequential operation. Processing upon request of the signal measured by the spectrometer such as implementation of alarms, graphic follow-up, measurement archives is also available. -
Description The detector is a commercial mass spectrometer which is specific to helium. It is at the center of a starshaped network with twelve sampling channels of which the shortest takes samples at about 5 m from the detector, the longest at 120 m. 689
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A measurement channel is a tube which brings the gas to be analyzed towards the spectrometer using a primary pump. The distance between the different measurement points lead us to look for a tube which would be easy to install, inexpensive and made in a material which would not be permeable to helium, even after aging. The diameter of the tube to be chosen must be sufficient to limit the suction pressure of the primary pump, while allowing the sample to quickly reach the detector. After an artificial aging study [2], we selected a polyamide tube of 4/2 mm diameter (commercial Rilsan tube). In the flow direction of the fluid, the channel consists of (fig. 1): - at suction, close to the components to be surveyed, a paper filter so as to avoid pluggings due to ambient dusts. There can be several suctions on the same channel. - near the detector, a micro-flow rate valve, and a three-way valve.
figure l : t h e helium leak detector All twelve channels are collected downstream from each three-way valve. The first outlet of these valves sends the sample towards the spectrometer. It then crosses an adjustable calibrated leak to lower the sample pressure to the operating pressure of the detector. The calibrated leak is the same for all the channels. Because of the different lengths of the channels, the pressure drops are not identical, which induces pressure differences upstream from the leak. We therefore installed the micro-flow rate valves on each line to impose on all measurement channels, a pressure of 100 Pa upstream from the leak with the primary pump of the spectrometer. These valves are near the detector in order to make their adjustment easier. The second outlet is connected to a second primary pump which ensures a permanent flow rate in all the channels awaiting measurements to optimize the response time and eliminate the line cleaning time. This innovation allows a sequential operation with short transient regimes to be obtained. We can measure the helium signal on the twelve channels in about 3 5 min.
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One of the channels measures the helium content of the outside air (5 10.9 mbarl/s for our application) and is used as a general reference for the measurement and auto-control system. The control is provided by a computer system which gives off an alarm when the measurement exceeds the pre-defined value according to the area considered. All the measurements are archived over a period of one month. A graphic treatment in real time or on the archived measurements is of course possible. Measurement performance By releasing gaseous helium volumes and/or flows in different parts of the TORE SUPRA cryogenic installation, to simulate punctual or continuous leaks, we have quantified and modeled the helium propagation mechanism in the halls and measured the sensitivity of our detection system. Kinetics 9 The response times of the detector and the size of the signal vary enormously depending on the size of the leak, its position in relation to the closest suction, and the length of the line. The choice of the polyamide tube enables, with a 100 Pa pressure upstream from the calibrated leak, a flow rate speed in the tube of 0.5 m/s. In the halls without draughts, the horizontal propagation speed of helium is of 0.06 m/s, the ascentionnal speed of 0.13 m/s. Given the size of the halls to be surveyed, the response times of the detector are of 2 to 5 minutes. Sensitivity, helium propagation With symmetrical release areas around the measurement point, at constant helium quantity, we observed a constant signal to noise ratio (diffusion symmetry). The signal to noise ratio is a decreasing function of the building volume. As with increasing helium quantities, we observed an increase of the S/B ratio (concentration variation effect). Obviously, the intensity of the leak signal in a hall is directly linked to the size of the leak. Using calibrated leaks (1.6 10 -6 m3/s to 5 10 "4 m3/s), we observed signal behavior depending on the ventilation of the halls. A non-ventilated hall always keeps the remanence of a leak according to its air-tightness. For the exterior reference of 5 10.9 mbarl/s, we measure, in the cryogenic hall, a minimum value of 10.8 mbarl/s. Thus, a leak in the cryogenic hall will decrease in 2 hours, an equivalent one will take 6 hours to decrease in the Toms hall, which is tighter. A hall with air removal removes a leak faster than a non-ventilated hall and thus presents only slight remanence. The helium rate (7 10.9 mbarl/s) is close to that of the outside air. Moreover, the different ventilations of the machines installed (as in the machine room) ensure air movement so that leaks are quickly diluted (Figure 2).
figure 2 machine room leak
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It is therefore important to place, for the detection, the suction of the channel near the air removal ventilators.
CONCLUSION Helium propagates in air under the form of a light smoke. In a non-perturbed environment, it ascends, forming a cone, then a layer at the ceiling of the room. For our installation, this means that for global detection to be efficient, the suctions must be as high as possible. Local measurement taking must be close to and above the objects likely to leak. A draught reduces detection sensitivity because of the air movement it causes and the loss of helium from the room to the outside it induces. A leak smaller than 8 10 .5 m3/s is then not to detectable. We chose thresholds of 3 10 .8 mbarl/s for a non-ventilated room and 1.3 10 .8 mbarl/s for a room with air removal as significant of the presence of a leak. These thresholds could be lowered, but we would be alerted for leaks which we believe are unavoidable and unimportant. The present system is very satisfactory. It detects numerous leaks, either by time correlation with operating incidents, or by geographical localization. The most frequently detected leaks are due to the handling of tanks (introduction and removal of transfer lines), to pressure relief valves, or to the packing boxes of noncollected valves, to pipe joining elements, even low pressure ones. The detection system is implemented in all of the cryogenic installation of TORE SUPRA. Moreover, the channels can be lengthened by flexible polyamide tubes ; it is thus possible to inspect a suspected area. We then have a classic sucking detection. Finally, this system is low maintenance, because of the verification of primary pumps and of the spectrometer every 8000 hours, and is very reliable. This new way of using existing techniques will find other applications, particularly in installations of industrial size using gaseous processes. Units: 1 mbarl/s = 0.1 Pa m 3/s REFERENCES [ 1] : Claudet G. and al., Cryogenics (1986) 26 443 [2] : Bonnel P, Violet JL, Etude de retention de l'hdium 4 par du rilsan internal report 1991
One Consideration For Recovery Heat Flux Of Directly Heated Wires Yanzhong Li*, Yezheng Wu*, Yuyuan Wu*, Udo Ruppert**, Ing,rid Arend**, Klaus LOders** *College of Energy and Power Engineering, Xi'an Jiaotong University, 710049, P.R. China **Physics Department, Free University Berlin, D-I 4195 Berlin, F.R. Germany In many cases of experimental investigations on heat transfer, the electrical resistance r of sample is used for heating itself directly and producing a variable heat flux, but the power supplied to sample is changed if sample resistance is changing during a burnout or a recovery from film boiling. For constant voltage power source it is theoretically found that, by adjusting the circuit resistance, the heat fluxes before and after jump are same, which allows a correct readout of the recovery heat flux. It is also experimentally found that the heat flux read from the onset point of film collapse gives constant value, which is independent of power supplying method and. wiring resistance.
INTRODUCTION Boiling heat transfer is widely investigated at low temperatures. In measurement of the study, the testing sample is generally supplied with a constant power source by using an extra electric heater built in the measuring unit, so that the heat flux of sample keeps constant during the burnout and recovery from film boiling. However due to the small size, for example a thin wire, sample is usually heated directly by using the electrical resistance of sample to produce a variable heat flux. The heat flux should be controlled to be constant with. a feedback inhibition system, but practically normal electric equipment hardly follows the rapid change over a jump. Therefore the value of iheat flux is not constant and much dependent on sample resistance. For directly heated sample, the power supply may serve a constam current[l ] or a constant voltage[2] to sample. The influence of power supply and sample resistance on heat flux is easier to find in a sample with low resistivity and high resistance-temperature coefficient. The preliminary report about the discovery was given in [3]. The work here is further to work out the characteristics of heat flux change under differem conditions and the solution to the problem by means of theoretical analysis and experiment. THEORETICAL ANALYSIS In many practical measurements, the sample is supplied with one of two different power sources (either constant current source or constant voltage source). In different cases, the heat flux may change up or down through burnout or recovery process as shown in figure l a. Figure l b shows a simplified circuit of power supply, which includes only wiring resistance R and sample resistance r in external circuit.
Figure 1 (a)Heat transfer performance (b) Principle of power supply circuit 693
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Power Source with Constant Voltage In steady heat transfer, power source serves a steady voltage E to extemal circuit. The current I through sample and voltage drop U on the sample depend on the sample resistance r, the relations are E I - ~ , R+r
rE U - ~ R+r
(1)
The heat flux produced on the sample surface in the case is I .U q =
r. E 2
(2)
I ~----7-= F ( R + r ) 2
F is heat transfer surface area. Suppose the power source serves an EA at point A (onset of bumout), and sample has a resistance rA, wiring resistance keeps R, then the heat flux is rA
(3)
qA ~ F ( R + r A ) 2
After the jump of burnout, heat transfer locates at point B, same equation is written as .
(4)
q B .... F ( R + r B ) 2
Because the jump happens very fast, no change of power voltage is considered, that is E A = E B = E , then
Aq =qB-qA
=
(R+rBj 2
(5)
(R+rA
The sample resistance through a burnout process always increases, i.e., rB > rA . By solving the equation (5), one obtains three possibilities as following Aq-0atR=4r.,.r
B
,
Aq>0atR>4G.r
~ ,
Aq<0atR<4r,-r
B
It is seen that the change of heat flux is also dependent on the selected value of wiring resistance R. With different value of R, heat flux will increase, decrease or be the same through a jump as shown in figure la ( q B > q A , q B < q A , and q B = qA). When R is equal to the geometric mean value of sample resistances at starting and ending points, the heat flux is the same at two points. If a relative changing ratio n - rB / r A is introduced, the relation R = ,r162 A is necessary to get an identified heat flux. Even though R = ~l-n.r A is met, the heat flux is also not constant at every point during the jump, i.e., d q ~ O. The changing characteristics can be achieved by differentiation with respect to equation (2) dq_
d
d r - -~r
rE 2 F(R + r) 2
t:. 2 d = F dr
r (R + r) 2
1'52
R-r
= F (R + r) 3
(6)
From the equation, d q / d r - 0 is tenable if r - R. Because R = ~ f ~ . r A is already selected and rA <_r <_rB is clearly known, heat flux q has a maximum at r = R = ~l-~.r A . Figure 2 shows the relation of heat flux q to sample resistance r through a burnout jump obtained by calculation, where n is dependent on sample material. For the samples used in our experiment, n ~ 2 for RhFe and n ~ 20 for Cu. Same way can be used to calculate the recovery process. To keep heat flux q equal at points C and D, the condition R = ~/r(, .r D must be required. The heat flux q passing through the recovery
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A q > 0 at R < x [ r c . r D ,
Aq <0 at R> x ] r C . r D
Here define n r = r c / r D , then the condition is R = x f ~ . r D. where n r ~ 1.2 for RhFe and n r ~ 3 for Cu. It is seen in figure 1 that the change of heat flux affects the locations of points B and D, i.e., the ending points of jumps. Point A and D are generally defined as the peak point and recovery point respectively. Nevertheless point A is fixed, but point D is relative to circuit resistance R and sample resistance r according to the analysis above. The location of point D influences the readout of recovery heat flux. Theoretically speaking, recovery point should only be a heat transfer feature and has nothing to do with measurement. Figure 2 Heat flux vs. sample resistance through In fact, the influence is inevitable in the system with a direct heating supply. The a burnout jump ( n = rB / r A ) influence degree relies on the sample material. For the material with high resistance-temperature coefficient (n is larger), pure metal for instance, there is a larger affection, but a smaller affection for alloy materials. It is experimentally found that for different wiring resistance R, the influence on recovery heat flux is less than 10% for RhFe sample and more than 40% for Cu sample. Power Source with Constant Current When a constant current is directly supplied to a sample, the heat flux at any case is as following r/2 q=~ F
(7)
2/ Because I~/~. is the same during a jump, heat flux may always go up through a burnout jump and go down during a recovery jump. It means the recovery heat flux from point D is always lower than it is, which has nothing to do with circuit resistance. The larger the n r of sample material is, the more different the recovery heat flux is. There is no way to keep the heat flux same at starting point and ending point. EXPERIMENTAL APPROACH An experiment was done in superfluid helium for RhFe ~38~t and Cu ~40~t wires with a triangular waveform voltage supply to verify the theoretical analysis. The measuring
Figure 3 Heat flux on a sample vs. running time of a triangular voltage supply with an alternative wiring resistance
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system is described in reference [4]. Figure 3 shows one of the recording curves at 2.0K for RhFe wire with three different wiring resistance (R = ~ . r A, > ~t-n.rA and < ,J-n.rA). The abscissa indicates the elapsing time corresponding to a triangular waveform voltage power supply. It is seen from the figure that the circuit wiring resistance R has much influence on the ending points of burnout and recovery jump. Anyway, the onset of a jump is always the same no matter what the wiring resistance is and no matter how largely the heat flux changes. The discovery implies that the recovery heat flux can be read out from jump onset point, i.e., the constant point C may reflect a recovery heat flux instead of point D. The phenomenon is verified by many repeated measurements. Figure 4 shows a test result for copper sample at Hells bath with two different circuit constructions, in which the constant recovery heat flux from point C is also confirmed in any case. However the corresponding sample temperature or other parameters at recovery point should be still taken from the reversible relation part of heat transfer to liquid by introducing a correct recovery heat flux from point C.
Figure 4 Heat transfer changing performance of Cu sample vs'. sample resistance (a) sample without parallel resistance (b) sample with a stainless steel foil (16.175mf~) CONCLUSION The influence of wiring resistance on recovery heat flux is first found for a directly heated sample with a constant voltage power supply in many repeated measurements. The reasons and affecting pattern are theoretically analyzed for different power supply sources and different wiring resistances. The solution to the measurement problem is given on the basis of analysis and verified in experiments. ACKNOWLEDGMENTS The work is supported by the National Education Commission of China. The experiment is completed in TTL, Free university Berlin, Germany, and supported by the Bundesminister for Wirtschaft, ERPSondervermOgen. REFERENCES
,
.
,
Gradt, T., Szt~cs, Z., Denner, H.-D., Klipping, K., Heat transfer from thin wires to superfluid helium under reduced gravity CEC31 (1986) 499 Wang, R., Peak and recovery heat flux densities in the bath of subcooled superfluid helium Cryogenics (1994) 34 983-990 Arend, I., Li, Y.Z., L0ders, K., Ruppert, U., Heat flux investigations on wires coated with porous insulation in superfluid helium Cryogenics (1996) 36 215-218 Li, Y.Z., Wu, Y.Z., Wu, Y.Y., Ruppert, U., Arend, I., LOders, K., Heat transfer from superconductor wire to superfluid helium ICEC 16 Kitakyushu, Japan, (1996)
A Novel Superconductor V-I Simulator for 77 K B. ten Haken and H.H.J. ten Kate University of Twente, Low Temperature Division, P.O. Box 217, 7500 AE Enschede, The Netherlands A passive circuit of two diodes and two resistors is applied in liquid nitrogen as a cryogenic simulator of the V(/) curve of a (high-To) superconductor. The simulated voltage is determined over a wide range of 10 -9 tO 10 -2 volt. The V(/) curve is described accurately by the diode characteristic and the two resistors in the circuit. For low voltages this leads to an n-value of about 100 for the simulated superconductor. The voltage noise is smaller than the detection limit (+3 nV) and the reproducibility is better than the accuracy in the current measurement (+0.05%). No changes in the V(/) curve are observed after thermal cycling between 77 and 300 K. The advantages of the cryogenic device, compared to the versions working near or at room temperature are: a better stability and reproducibility, smaller size and lower costs. An additional advantage occurs when the device is applied to validate measuring set-ups. The cryogenic V-I simulator replaces the superconducting HTS sample inside the nitrogen bath cryostat and then can be considered as a reference sample. In this way the cryogenic V-I simulator includes additional sources of errors such as the influence of thermal noise, leakage currents, common mode rejection and instrumentation errors as ground loops.
INTRODUCTION The non-linear voltage current characteristic of superconductors can be simulated with a circuit of a diode and two resistors. Such a circuit can be used to test accuracy and reproducibility of critical current measuring set-ups. This device is especially suitable for investigating the influence of different noise sources on the determination of a static voltage-current characteristic V(/). Because of the non-linearity in the V(/) characteristic of a superconductor an AC noise in the current may cause an error in the determination of the DC voltage that is measured, especially in the steep part of the voltage transition. Such an error can easily be observed in a simulator circuit with a V(/) characteristic similar to a superconductor. The novel device can be seen as an alternative to the already exiting system that operates at high currents (1 to 100 A) and is applied to compare the performance of measurement systems for the critical current [1]. The power that is generated in a V-I simulator is significant (about 60W in the example). The stability of the device is effected by the heat that is generated in such a system especially when it is operated at room temperature. A reasonably stable operation can only be obtained by regulating the temperature in the V-I simulator with an active circuit. In order to minimise the influence of thermal drift we decided to develop a cryogenic V-I simulator at 77 K where liquid nitrogen bath acts as a thermal buffer.
OPERATING PRINCIPLE A passive circuit of two diodes and two resistors produces an exponential V(/) characteristic that is very similar to that of a superconductor. The current Id through a diode depends on the voltage Vd as: Ij = 10(exp(V,l/r/Vt ) - 1) with Io as a temperature dependent current parameter and V, = kT/q --6.64 mV at T = 77 K (q = electron charge and k = Boltzman's constant) and 7/= the ideality 697
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factor ranging from 1 to 2 for a real diode [2]. As long as the current in the diode is small compared to the total current I~ that is fed through the simulator, the simulator voltage (Vs) is:
(1)
V.,. = R2 . Io(exp(I,. . R ~ / r I V t ) - I ) . The n-value of this simulator, following the empirical power law V o~ I n - - -din(V,.) ----m= din(/.,.)
I,..R~ (1 - exp(-I,. 9R,/r/Vt) ) = I,..R~ ' " r/Vt r/Vt
n
can be directly calculated with:
for I s R l >> 7/Vt .
For a typical Si-diode in liquid Nitrogen, with a diode voltage of Vd = Is "R1 factor close 1.6, this leads to an n-value between 90 and 100.
(2)
=
1 V and with an ideality
The passive circuit as described here is particularly useful as a V-I simulator for testing equipment that measures the V-I characteristic of superconductors. Compared to the existing V-I simulator, operating at room temperature the voltage transition is much steeper. The V-I simulator operated at room temperature has an n-value of 27 [1]. A1 the major problems regarding the temperature changes in the room temperature device are minimised when V-I simulator is operated in a nitrogen bath at 77 K. The power generated in the diodes can be restricted to a few milli-watts maximum. Such a small power is easily absorbed in the N2 bath, without causing any measurable changes in the diode temperature nor its V(/) characteristic. The power generated in R2 is also negligible, but in RI there is a significant power of Va'L. By selecting a well cooled and stable resistor for Rt this problem can be overcome too.
EXPERIMENTAL RESULTS A V-I simulator is characterised at 77 K in order to demonstrate the feasibility of the device. Two medium sized silicon rectifier diodes (1N5401-DC, rated at room temperature for 3 A) are combined with a 50 W power resistor (Rl=100 mf~). The simulator voltage is measured over a 15 mm piece of resistive wire (R2=60 m ~ ) . The combination of the resistance value for Rl and the diode voltage leads to a voltage transition at 10 amperes. The slope of the transition can be described with equation 1 which leads to an n-value ranging from 100 to 85, in the voltage regime from 0.01 to 10 ktV. The description formulated in equation 1 becomes invalid for high current values where a significant part of the current starts to flow through the diode. Above a simulator voltage of 30 laV the n-value decreases significantly due to the internal resistance of the diode Rd (not drawn in figure 1). A V(/) curve as measured on this simulator at 77 K, over more than 6 voltage decades, is presented in figure. 2. The data points are obtained at a constant current in a step-wise manner. With a careful offset correction at zero current and sensitive (nano-)voltmeter the voltage noise can be reduced to +3 nV. The measured data can be described with an ideality factor of 7/= 1.65 and a conduction resistor of Rd = 3.5 ~ for this particular set of diodes. Is
RI
IOl
ld
Re
Vs Fig. 1 The V-I simulator circuit with a double set of diodes for bipolar operation.
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Fig. 2 The measured voltage across the V-I simulator at 77 K. The solid line represents the data fit with 7/= 1.65 and Rd = 3.5 ~ .
NOISE AND SENSITIVITY. At low voltages (V < 1 ktV) an accurate detection of the voltage becomes a delicate matter. Due to the stable (and low) temperature in the nitrogen bath and the small resistor values present in the circuit, the voltage noise in this simulator is relatively low. The combination of a ground contact in the simulator circuit and a sensitive voltmeter with a direct twisted Cu-pair to the simulator sample leads to voltage noise of (+3 nV). With regard to this noise level, a constant voltage reading is observed over the entire current range from zero to the onset of the V-I transition at 9.5 amperes. A detailed view of this transition is presented in figure 3.
STABILITY AND REPRODUCIBILITY. The stability of the V-I simulator is determined mainly by the stability of the resistor R~ that carries the largest part of the current and the diode characteristic. Due to the stable temperature and effective cooling of the liquid nitrogen bath no variations in the simulator V(/) characteristic could be observed with the available accuracy in the current measurement 0.05%. This excellent stability remains when the device is thermally cycled between 77 K and room temperature. No changes could be observed after multiple thermal cycles (investigated up to 10) within the sensitivity limits of the voltage and the accuracy of the current measurement.
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Fig. 3 The measured voltage across the V-I simulator at 77 K, magnified at the onset of the V-I transition. The solid line represents the data fit.
CONCLUSIONS The performance of a passive V-I simulator for superconductors is investigated. A stable and reproducible V(/) characteristic is shown in a nitrogen bath cryostat at 77 K. The V(/) characteristic is described well with the present descriptions for Si-diodes. If the conduction resistance of the diode is taken into account, then the V(/) curve is defined accurately over 6 orders of voltage magnitude. A stable and reproducible operation is shown within the experimental accuracy (+3 nV in voltage and +0.05% in current) The slope of the V(/) curve at the transition is for low voltages determined by the ideality factor 7/of the diode and the voltage across the diode. For a Si-diode at 77 K this leads to an n-value that is close to 100. For large voltages a higher current through the slope of the V(/) transition decreases towards a constant value that is determined by the power resistor R~ and the ratio between R2 and the conduction resistance of the diode (Rd). Besides the improvement of the stability of the cryogenic simulator compared with a device operated at room temperature, there is the additional advantage that the cryogenic version has a small size and can replace a HTS-sample in the nitrogen cryostat. This enlarges the number of possible error sources that may influence the determination of the V(/) characteristic. Additional noise sources as thermo-couple noise in the current leads and wiring, common mode errors, pickup-noise, ground loops and leakage currents are included in an equipment test with a cryogenic voltage simulator. The demonstrated properties of the novel cryogenic version of the V-I simulator makes the device suitable for use as a reference sample for critical current measuring systems. Especially in round robin tests, as for example in the on-going VAMAS programme, the cryogenic V-I simulator can be used to compare the differences between the test set-ups in the different laboratories accurately. In an industrial environment a cryogenic V-I simulator can be applied as a reference sample in a quality assurance routine of the characterisation equipment for superconductors.
REFERENCES D. Aized, et al. Comparing the accuracy of critical current measurments using the voltage-current simulator, IEEE trans. On Magn. Vol. 30, No. 4, p. 2014, 1994. Tyagi M.S., Introduction to semiconductor materials and devices, John Wiley & Sons, New York 1992.
The Estimation of Critical Current Density Using SRPM and AC Methods Shuichi Koto *, Hiroshi Nakane *, Edmund Soji Otabe ** , Teruo Matsushita ** Shigeo Nagaya ** ,and Shuji Yoshizawa **** *Department of Electrical Engineering, Kogakuin University, 1-24-2 Nishi-Shinjuku, Shinjuku-ku, Tokyo 163-91, Japan **Department of Computer Science and Electronics, Kyushu Institute of Technology, 680-4 Kawazu, Iizuka-shi, Fukuoka 820, Japan ***Electric Power Research and Development Center, Chubu Electric Power Co., Inc., 20-1 Kitasekiyama, Ohdaka-cho, Midori-ku, Nagoya 459, Japan ****Central Research Laboratory, Dowa Mining Co., Ltd., 277-1 Tobuki-cho, Hachioji-shi, Tokyo 192, Japan The critical current density ( J ~ ) of Y-Ba-Cu-O sample made by the zone melting process was estimated in SRPM and AC methods. When it was measured under AC and DC magnetic flux density, J ~ obtained in the both of the methods agrees well. There is a possibility to be able to apply the SRPM method to the measurement of J c.
INTRODUCTION We have been investigating a method which can simultaneously estimate both the resistivity ( p ) and the magnetic penetration depth ( 2. ) by vectrially measuring the difference in the impedance between two circular solenoid coils; one with and the other without a rod-shaped sample conductor. (SRPM method) [1]. It is easy to measure the frequency dependence of p and 2.. On evaluating the property of superconductors, the critical current density ( J ~ ) is taken as one of the most important properties except the temperature properties of p and 2 . In order to estimate J ~, AC method (Campbell method) is commonly used [2]. In this method, ). is estimated when the AC magnetic field ( b ) is supplied for the specimen, and J ~ is calculated from the inclination between b and ). under DC magnetic field. For the SRPM method, J ~ was obtained by using the same analysis. The values of J obtained from the both methods were compared. In this paper, the effectiveness of the SRPM method for the estimation of J ~ is discussed. METHOD FOR ESTIMATING J c The process for obtaining J c by the SRPM method is shown as the flow chart in Fig.1. In the SRPM method, the impedance of the coil was vectrially measured twice at the same temperature: once when a rod-shaped sample was coaxially inserted into the solenoid coil and then it was pulled out of the coil. The differences between the real part ( A R ) and the imaginary part ( A X ) in the impedance of the coil at the different conditions mentioned above were obtained from the measurement, while the impedance change of the solenoid is theoretically expressed in [1]. A R and A X calculated by the equation are shown in Fig.2 as the map using the parameters of 2 and p .The 30-layer solenoid coil of 832 turns with the average radius of 5.31mm, length of 3mm, the specimen of radius of 2.27mm, and the frequency of lkHz were used in Fig.1. From the point on the map the parameter of 2 is obtained. In this case, it is apparent that /! consists in the range of 10-~'~ 10 -" [m]. ). at each b is detected by magnifying the range. In the AC method, AC magnetic field was given parallel to the 701
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ICEC 16/ICMC Proceedings < SRPM Method > Impress b and B
Calculation of A R , A X
Measurement of A R , A X
Map of A R , A X
Estimation
of
Vp, V e
,
I
Calculation to Obtain Property of b - 2 '
Property of b - 2 ( p )
t Jc
Jc Fig.1
Measurement
,
o~
Comparison
< AC Method >
Flow Chart
specimen under DC magnetic field. Using the voltage of pick up coil ( V~ ) and cancel coil ( V c ), the difference between V p and V ~ is proportional to (I) differentiated by time. Using the change of fluxoid due to AC magnetic field, the magnetic penetration depth ( 2 ') is calculated as
1] '=R[i-[1 ~R"i aba~ "T
(1)
where R is the radius of the specimen. The relation between b and /l ' is obtained from eq.(1). The slope between b and /l is concerned with J ~. According to the Bean' s critical state model, /l ' is approximately expressed as (2) )
~
/loJc The slope of b - / l represents 1 / / . t 0 J ~. This method of obtaining J ~ is called Campbell method. In the measurement of Y-Ba-Cu-O by the AC method, two slopes are observed; one is concerned with the transport critical current density, the other is concerned with the local critical current density. Using these two, J c is investigated.
MEASUREMENT AND DISCUSSION The 20 wt.% Ag doped rod-shaped YBCO superconductor which was made by the zone melting process was used as the sample [3]. The sample size is a radius of 2.215mm, length of 12mm. A current( I ) of 1~ [kHz] from 10 to 140 [mA] were supplied into the coil and then b from 9.59 • 10-4 to 1.34 • 10-2 [T] were generated. At the same time, DC magnetic flux density ( B ) from 0 to 0.03 [T] is supplied parallel to the specimen. The specimen is maintained at a temperature of 77.3 [K]. In the SRPM method, /l is obtained from the map of A R and A X as described above. At B - 0 . 0 1 and 0.02 [T], on account of the supplied DC magnetic field, 2 appears at b - 0. Taking it into consideration, /l ' of the SRPM method is shown as Fig.3. It is apparent that the slope between b and /l ' is linear especially when B is 0.01 and 0.02 [T]. It had been expected that the two slopes were available also in the SRPM method, but in this result the only one slope was obtained. Fig.4 shows the profile obtained in the AC method. Compared with b - 2 ' of the SRPM method, the value of /l ' at the region of strong b is quite similar, but in the SRPM method the value at B - 0 is different from the others. In the SRPM method, We tried to calculate the critical current density using the inclination
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Fig.2 Map of A R - A X
Fig.4
b-
703
Fig.3 b - /1 ' of SRPM method
2 ' of A C m e t h o d
Fig.5
J ~ in SRPM and AC methods
of the slope. J ~ obtained in the SRPM and in the AC method is shown in Fig.5. In the SRPM method, the result of J ~ under the DC magnetic field of 0 ~ 0.03 IT] was obtained. J ~ obtained in the SRPM method is similar to that of the AC method. J ~ obtained by the VSM is 1.7 • 10 ~ [A/m ~ ], and by the DC 4-probe method is more than 1.6 • 10~ [A/m ~ ] (as shown in table 1). All of the method could have nearly the same value. However, the appropriate calibration concerned with the effect of the shape of the specimen and the existence of cracks and normal conductive material in the superconductor on the value of 2 could not have been done in the SRPM method. Table1. Comparison of J ~ ( at 77.3 [K] ) B [Wb/m x]
0.01
O.
02
SRPM method
AC
method
1.039 • 10
3.063 • 10 ~
9.549 • 10
1.990 x 10
7.520 • 10 7
1.816 •
10
VSM
1.7 •
10
4-probe method
> 1.6 • 10 ~
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CONCLUSION When J c of rod-shaped sample was measured by the SRPM and AC methods under AC and DC magnetic flux density, J ~ obtained in both methods agrees well. It is obvious that there is a possibility to be able to apply this method to the measurement of critical current density. In the AC method, two slopes concerned with the transport critical current density and with the local critical current density are observed, but in the SRPM they are not observed. We have to investigate the reason. And moreover, the appropriate calibration concerned with the effect of the shape of the specimen and the existence of cracks and normal conductive material in superconductor on the value of /l ought to be done in the SRPM method. REFERENCES H.Nakane et.al, Calculation of The Difference in Impedance for a Solenoid Coil with and without a Sample Conductor, IEEE Trans. Instrum. Meas., (1991) 40 544-548 E. S. Otabe et.al, Estimation of Critical Current Density in a Melt-Processed Superconducting Y-Ba-Cu-O Using AC and DC Inductive Methods, Jpn. J. Appl. phys. (1994) 33 996-999 S. Nagaya et.al, Ag Doping Effects on the Microstructure and Properties of Unidirectional Grown Y-Ba-Cu-O Superconductors, IEEE Trans. Magn. (1995) 5 1564-1567
AUTHOR INDEX
Abe, T., 1907 Abe, Y., 1333 Agapov, N., 139 Agatsuma, K., 1701 Ageta, T., 891,899 Aihara, K., 1261, 1685, 1689 Aiyama, Y., 891 Ajima, Y., 419 Akazaki, T., 1193 Akhmetov, A., 1829, 1833 Akinaga, M., 1647 Akita, S., 933 Aksenova, E. N., 1937 Akune, T., 1443, 1561 Alexeev, A., 395 Amano, K., 1707 Amardas, A., 819 Amemiya, N., 1617, 1775 Amrit, J., 559 An, C.-W., 847, 1269 Ando, M., 493 Ando, T., 763, 767, 771,775, 783,787, 791, 795, 807, 1265, 1661, 1665, 1673, 1677, 1895, 2009 Andoh, H., 1571 Andrikidis, C., 1353, 1545, 1549 Antipov, E. V., 1627 Aoki, K., 867 Aoki, N., 1681 Aoki, R., 399 Arai, K., 1297, 1701 Araki, S., 747 Araki, T., 263, 307 Arasawa, K., 267 Arata, M., 1273 Arend, I., 523, 547, 693 Asakura, H., 203, 207 Asakura, S., 999 Asami, H., 351 Asano, K., 751, 755, 759, 811, 1017, 1049, 1989, 2035, 2061 Asano, T., 1089 Aubin, M., 1621 Awaji, S., 1109, 1121, 1365, 1669, 1719, 1723, 1841 Ayai, N., 1673, 1677 Aymar, R., 53 Azuma, N., 1779 Baba, T., 75, 79, 83, 87, 735 Babic, E., 1557
Baenitz, M., 1561, 1627 Baixeras, J., 1479 Ballarino, A., 1139, 1143, 1147 Ban, T., 1655 Barclay, J. A., 2065 Barranco-Luque, M., 103 Batin, V. I., 439 Baudouy, B., 563 B/iuerle, D., 1533 Beales, T. P., 1541 Beduz, C., 1541 Benda, V., 199 Bender, S., 311 B6zaguet, A., 91, 519 Bhattacharya, D. G., 643 Bhattacharya, J. L., 949 Bhattacharya, N. C., 843 Bi, Y.-F, 151,847 Bian, S.-X., 287 Bianco, M., 1499 Binneberg, A., 505, 509 Biswas, B., 823, 827 Blackburn, T. R., 1037 Blau, B., 1665 Bocquillon, A., 435 Boehm, J., 1117 Bohno, T., 1053, 1151 Boiko, B. B., 1949, 1953 Bon Mardion, G., 689 Bona, M., 1911 Bondarenko, S. I., 1177 B6rner, H., 669 Bose, T. K., 2065 Boucheffa, A., 559, 563 Bourque, R., 427 Brandst/itter, G., 1579 Brehm, H., 259 Breitzke, H., 1627 Bremer, J., 173 Brissette, Y., 1021 Brunovsk2), I., 195 Bruzek, C. E., 1321 Bruzzone, P., 1243 Buhler, S., 449 Bunce, G., 871,875 Burger, J. F., 391 Buschmann, H., 509 Cai, J.-H., 271 Cai, X., 1189 Calvone, F., 1871
Camille, R. J., 803 Casas-Cubillos, J., 91 Cave, J. R., 1021, 1621 Chaffard, P., 107 Chahine, R., 2065 Chandratilleke, R., 501 Chauhan, H. S., 1509 Chaussonnet, P., 435 Chen, C.-Z., 217, 221, 225, 229, 233, 497 Chen, G.-B., 247, 275, 315, 407, 639, 677 Chen, G.-M., 407, 639 Chen, H. Y., 385, 677 Chen, Y., 359 Cheng, S.-K., 1447 Chensky, I., 799 Chiba, A., 1907 Chiba, H., 1847, 1851, 1863 Chida, K., 87 Chiba, M., 1711, 1847 Chien, S. B., 385 Chikaba, J., 1369 Chikaraishi, H., 735, 739, 743, 747, 751 Choi, I.-H., 1467 Choi, K.-D., 941,953 Chubraeva, L. I., 895, 925, 929, 1933, 1937 Claudet, S., 103 Collaudin, B., 711 Collings, E. W., 1397, 1609, 1767 Cowey, L., 1143 Cragg, D., 111, 115, 123 Cruikshank, P., 681 Cui, G.-W., 677 Cullen, J. R., 871,875 Cur6, C., 111, 123 Daido, K., 481 Daniels, P., 1117 Das, S., 823, 827 Das, S. K., 643 Dauvergne, J. P., 107, 111, 169, 173, 199 Decker, L., 195 Delikaris, D., 169, 173 Delruelle, N., 173 De y, R., 643 Dhard, C. P., 819 Din, L. R., 847
12
A u t h o r Index
Doi, Y., 119, 165, 419, 843, 867 Doko, T., 1915 Dolez, P., 1621 Dolgosheev, P., 799 Dong, D., 719 Dou, S. X., 1037, 1353, 1393, 1397, 1401, 1545, 1549, 1557, 1609 Drozd, A. A., 1941, 1945, 1953 Duan, Z., 1373 Durga Prasad, K. A., 949 Duthil, R., 123 Ebihara, K., 1483 Egawa, K., 1731, 1771 Egi, T., 1463, 1467 Emoto, M., 673, 1285 Endo, A., 1509 Endo, S., 182 I Enoki, T., 1193 Erbuke, L., 863 Eriksson, H., 863 Ernesto, M., 983 Escher, U., 2023 Ettlinger, E., 707 Evans, D., 2017 Evans, L. R., 45 EXSIV Group, 739, 743, 747 Ezaki, T., 1045, 1049 Fabian, P. E., 1997 Ferlin, G., 443 Filippov, Yu. P., 439, 609 Fisk, H. E., 863 Fleshier, S., 1613 Flokstra, J., 1181 Flukiger, R., 863 F61de~iki, M., 2065 Franco, C., 983 Frangois, M. X., 559, 563 Frederking, T. H. K., 657 Friend, C. M., 1541 Fuchino, S., 593, 957, 1297 Fuji, H., 1013, 1867 Fujii, M., 473 Fujii, Y., 263, 461,567, 621, 1525 Fujikami, J., 967, 975, 1347, 1597 Fujima, K., 1297 Fujimoto, H., 1655 Fujimoto, S., 331, 1173 Fujinami, T., 331 Fujino, K., 1413, 1583 Fujinuma, S., 1073 Fujioka, K., 291, 513, 2069 Fujioka, T., 767, 783 Fujisaki, H., 779 Fujishima, S., 855, 859 Fujita, T., 183, 665 Fujiyama, N., 381 Fujiyoshi, T., 1505 Fukano, T., 87 Fukasaku, Y., 279 Fukuda, K., 1669, 1685 Fukuda, T., 2069, 2073 Fukuda, Y., 1693 Fukui, N., 1173 Fukui, S., 1775
Fukushima, K., 1521 Fukuyama, S., 1919 Funaki, K., 1009, 1049, 1053, 1325, 1329 Funayama, H., 1727 Furumoto, H., 403, 601 Furuto, Y., 913, 917, 921 Furuya, S., 267 Futaki, N., 1381, 1601 Ganzinov, 1. S., 929 Gao, J. L., 295, 303 Gao, C.-H., 221,497 Gauthier, A., 519 Gayet, Ph., 103 Geilenkeuser, R., 2031 Genoud, J.-Y., 1571 Gerstenberg, H., 2001 Gerster, J., 259 Geynisman, M. G., 143 Ghose, D., 643 Ghosh, A. K., 863 Ghosh, G., 1957 Giesey, R. K., 2053 Gifford, P. E., 363 Gippius, A. A., 1627 Gistau-Bagucr, G., 189 Gladun, A., 2023 Godeke, A., 803 Gong, L.-H., 343, 347 Goodrich, L. F., 1715 Gopal, R. B., 2065 Goto, A., 855, 859 Goto, K., 933, 1073, 1197, 1723, 1841, 1867 Goto, T., 1189 Gradt, T., 669 Grantham, C., 1037 Gravil, B., 689 Green, M. A., 871,875 Gregory, E., 1715 Grimaud, L., 519 Grman, D., 863 Grunblatt, G., 1321 Gu, A.-Z., 581 Guinaudeau, H., 91 Gulko, E., 1715 Gtintherodt, H.-J., 1627 Guo, Y.-Y., 287, 613 Guo, Y. C., 1037, 1353, 1393, 1401 Haberstroh, Ch., 395 Hahakura, S., 1347 Hahn, S.-Y., 941 Hakamata, M., 1775, 1847 Hakuraku, Y., 1491, 1495 Hama, K., 535 Hamada, H., 1525 Hamada, K., 127, 131, 135, 427, 493, 767, 783 Hamada, M., 1061, 1095, 1293 Hamada, T., 1443 Hamaguchi, S., 539 Hamasaki, K., 1197 Hamashima, T., 1727 Han, S., 1513
Han, G., 1919 Hanaoka, Y. W., 1685, 1689 Hangyo, M., 1205, 1209, 1631 Hara, K., 183, 851, 1249 Hara, M., 1009 Hara, N., 1305, 1309, 1313 Hara, T., 963, 967, 975, 979, 1029, 1033, 1413, 1575, 1583, 1597, 1605 Harada, N., 1693 Harada, S., 1583 Haraguchi, E., 1173 Harrison, S., 1143 Hartwig, G., 1977 Haruyama, T., 119, 419, 513,649, 843, 867 Hasanain, S. K., 1517 Hase, N., ! 571 Hase, T., 1409 Hasebe, T., 1109, 1121 Hasegawa, H., 909 Hasegawa, K., 1413 Hasegawa, M., 771 Hasegawa, T., 1361 Hasegawa, Y., 399 Hashimoto, T., 325, 1377 Hata, K., 535, 585, 617 Hatakeyama, H., 501, 1113 Haug, F., 107, 111, 123, 169, 173, 199 Hayakawa, N., 1837 Hayashi, H., 735, 1017, 1151 Hayashi, K., 1009, 1329, 1347, 1413 Hayashi, S., 1129, 1409 He, J.-H., 1919 Heiden, C., 283, 311,453 Heinze, M., 1627 Herz, W., 469 Herzog, H., 147 Herzog, R., 505, 509 Hideto, Y., 335 Higashi, N., 419, 851, 1249 Higo, S., 1491, 1495 Higuchi, K., 1103 Higuchi, N., 1297 Hikata, T., 1347 Hikita, M., 1837 Hilbert, B., 91 Hino, N., 1305 Hirabayashi, H., 513, 851 Hirai, H., 2009 Hirakawa, K., 1451 Hirano, N., 1227 Hirano, S., 1583, 1597 Hirao, Y., 211 Hirayama, T., 1455 Hiresaki, Y., 303 Hirose, R., 1095, 1293 Hirumachi, T., 783 Hisada, S., 665 Hishinuma, Y., 1427 Hiue, H., 135, 743, 747 Hiyama, T., 127, 493 Hoffmann, W., 1627 Hofmann, A., 239 Hojo, M., 1791 Hommei, T., 1309, 1313
Author Index Honda, T., 127, 131, 135, 427, 493, 767, 787, 795 Honda, Y., 1277 Honjo, S., 963, 975, 979, 1575, 1605 Honjyo, S., 967, 1413 Honma, H., 1073 Honmei, T., 1305 Horigami, O., 1361 Horiguchi, K., 1887, 1981 Horise, R., 1129 Horiuchi, T., 1103 Horiya, T., 1895, 1915, 1923 Horvat, J., 1353, 1401, 1557 Horvath, I. L., 863 Hosaka, M., 1003 Hoshino, T., 941,953, 1639 Hosokawa, M., 1041 Hosoya, T., 203 Hosoyama, K., 183, 851, 1239, 1249 Hotta, Y., 811 Huang, B. J., 385 Huang, J.-H., 581 Huang, X., 571 Huang, Z.-P., 359 Huang, Z.-X., 247, 275, 677 Hubler, U., 1627 Htibner, R., 1977 Htibner, W., 669 Humer, K., 1997, 2005 Hussain, M., 2057 Huwiler, R., 863 Ichige, K., 493 Ichihara, M., 1681 Ichihara, T., 771 lchikawa, T., 905, 991, 1847 Ichikohara, H., 1693 Ichimaru, O., 1817 Ichiyanagi, N., 979 lde, Y., 1561 li, H., 1405, 1575 Iida, F., 135 Iida, M., 419 Iida, T., 1915 Iijima, Y., 995, 1003, 1669, 1697 lima, M., 735 Iimura, K., 79 Ijspeert, A., 1139, 1143, 1147 lkeda, H., 1385, 1435 Ikeda, Y., 159, 211 Ikegami, K., 855, 859 Ikegami, T., 1483 Ikeuchi, M., 159 Ikeya, T., 2069 Ikuhara, Y., 1455 Imagawa, S., 735, 739, 751,755, 759, 1825 Imai, Y., 909 Imayoshi, T., 1017, 1049, 1151, 1829 Imokawa, H., 1201 Inaba, S., 751 Inagaki, J., 783 Inoue, K., 913, 917, 921, 1089, 1099, 1103, 1361, 1409, 1669, 1697, 1735 lnoue, N., 735 Inoue, T., 299
Inoue, Y., 1129, 1711 Ioka, S., 735 Ionescu, M., 1353 Ipatov, Y., 1969 Irie, F., 1049, 1151, 1567 Ishibashi, K., 1163 Ishige, K., 1923 lshigohka, T., 513, 1069, 1337 lshihara, M., 1109, I 121 lshii, H., 621,963,967, 975, 979, 1413, 1575, 1583, 1597, 1605 Ishii, I., 593, 957, 1297 lshikawa, J., 831 Ishiyama, A., 513, 1239, 1281 lshizuka, M., 1377, 1385 lsogami, H., 1125 Isojima, S., 617, 967, 971, 975, 1057, 1347 Isono, T., 767, 771, 775, 1265, 1731, 1771 Ito, D., 1553, 1821 Ito, M., 1347, 1775 Ito, T., 779, 791, 1301 Itoh, A., 427, 1643 Itoh, H., 685 Itoh, I., 743, 795, 1325, 1329 Itoh, K., 1103, 1735, 1783, 1787, 1799 Itoh, M., 1499 Itoh, R., 1049 Itoh, S., 431, 1129 Itou, I., 807 Ivano, O., 1895 Iwabuchi, A., 1727, 1907, 1961, 1965 Iwaki, G., 1669, 1685, 1689, 1719 Iwakuma, M., 1009, 1325, 1329 lwamoto, A., 75, 79, 83, 87, 605, 735, 739, 743, 751, 1227, 1253 Iwamoto, K., 71, 155 Iwamoto, S., 807 iwasaki, S., 1723, 1841 lwasaki, T., 203 Iwasaki, Y., 2005 Iwaski, S., 933, 1443 J/ickel, M., 2023, 2031 Jacob, S., 413, 949 Jaffery, T. S., 1235 Jager, B., 435, 689 Jayakumar, J., 763 Jenninger, B., 443 Jeong, S.-K., 1215, 1219, 1223, 1231 Jess, P., 1627 Ji, P., 1447 Jikihara, K., 1109 Jin, J. X., 1037 Jochimsen, G., 423, 707, 715 Juillet, J. J., 711 Kabashima, S., 539 Kabe, A., 183, 851, 1249 Kai, T., 739, 743 Kaiho, K., 1013, 1297 Kaiser, G., 259 Kaito, T., 1057 Kajikawa, K., 1013, 1297 Kakehi, Y., 1201, 1591
13
Kakimi, Y., 291 Kakugawa, S., 1305 Kalinin, V., 131 Kamata, K., 1851, 1863 Kamada, S., 1431 Kamikado, T., 1293 Kamioka, Y., 457 Kamiya, I., 1045 Kamiya, K., 531 Kanari, T., 543 Kanazawa, Y., 339, 351,355 Kanda, Y., 1337 Kanegae, K., 1329 Kanekiyo, T., 183 Kaneko, T., 1347 Kancko, Y., 203, 207 Kang, Y.-M., 629, 1173 Karimoto, S., 1487 Kariya, J., 673 Karunanithi, R. 413, 949 Kasagawa, Y., 1009, 1325, 1329 Kasahara, H., 933 Kasahara, S., 331 Kashima, T., 1993, 2049 Kasthurirengan, S., 413, 949 Kasuga, T., 1333 Kasuu, O., 1155 Kasuya, M., 267 Katada, M., 67, 83 Katheder, H., 1665, 2001 Kato, H., 1929 Kato, S., 419, 843 Kato, T., 127, 131, 135, 427, 493, 767, 771,775, 795, 1347 Katoh, Y., 735 Katsumura, Y., 1357 Kauschke, M., 465 Kawabata, C., 1651 Kawabata, S., 921 Kawaguchi, E., 263, 307 Kawaguchi, T., 855, 859 Kawai, M., 419, 867 Kawakami, A., 1185 Kawamata, H., 419, 843, 851, 1249 Kawakami, K., 551 Kawano, H., 1305, 1309, 1313 Kawano, I., 1491 Kawano, K., 127, 473, 493 Kawano, S., 299 Kawasaki, K., 1435 Kawashima, I., 127 Kawate, Y., 431,513, 1103, 1129, 1409 Kazumori, M., 885 Kesseler, G., 173 Khodzhibagiyan, H., 139 Kiboshi, T., 1753 Kida, J., 1103 Kido, G., 1089 Kido, T., 1169, 1173 Kikuchi, A., 1471 Kikuchi, K., 1685, 1689 Kikuchi, M., 1681 Kim, J.-W., 855, 859 Kim, S., 419, 843 Kim, S.-K., 1281 Kim, S. W., 1767
I4
Author Index
Kim, Tae Hyun, 1133 Kimura, A., 913, 917, 921, 1575, 1605 Kimura, H., 1073, 2069, 2073 Kimura, M., 1669, 1719 Kimura, N., 419, 649, 843, 1129 Kimura, Y., 851 Kirby, G., 837 Kishida, T., 1057 Kisida, T., 909 Kiss, T., 1567, 1587 Kitagawa, K., 211 Kitaguchi, H., 1099, 1957 Kitamura, M., !305 Kitamura, R., 1803 Kiyoshi, T., 431, 1089, 1099, 1103, 1361, 1409, 1735 Klundt, K., 283 Knoopers, H. G., 803 Knoops, S., 199 Koba, S., 1491, 1495 Kobayashi, H., 551 Kobayashi, N., 1669, 1707, 1719, 1739 Kobayashi, S., 1347 Kobayashi, T., 863, 867, 1189 Kobori, T., 1887 Kodama, T., 307 Koga, T., 1069 Kohler, C., 1321 Kohno, O., 933, 987, 995, 999, 1003, 1013, 1073, 1381, 1601, 1723, 1841, 1867 Koike, T., 1083 Koizumi, N., 767, 771, 775, 791, 795, 1301, 1673, 1677 Koizumi, T., 1361 Kojima, Y., 183, 851, 1249 Komatsu, K., 885 Komatsu, M., 1537 Konda, H., 1993 Kondo, T., 119 Kondo, Y., 867 Kondou, Y., 419 Konno, M., 795, 1009, 1053, 1151, 1325, 1329, 1665 Konosu, S., 1895 Kos, N., 681 Kosaka, T., 381 Koshizuka, N., 1463, 1467 Kosuge, M., 1099, 1697 Kosugi, K., 1405 Koto, S., 701 K0uki, N., 335 Kouriki, K., 811 Kovalenko, A., 139 Kovrizhnykh, A. M., 439 Koyanagi, K., 1113, 1707 Kr/ihling, E., 2001 Kreisler, A., 1479 Krempetz, K., 863 Krooshoop, H. J. G., 803 Kubo, S., 1487 Kubo, T., 855, 859 Kubo, Y., 1731, 1771, 1799 Kubota, H., 1749, 1763 Kubota, Y., 1763, 1779 Kuchiishi, Y., 751
Kukano, T., 83 Kuma, S., 1389 Kumakura, H., 1099 Kumano, T., 1799 Kume, A., 987, 1381, 1601 Kundzins, K., 1579 Kurahashi, H., 1739 Kurihara, T., 331, 367 Kuriyaki, H., 1451 Kuriyama, F., 279 Kuroda, K., 1463, 1467, 1833 Kurtyka, T., 1911 Kurusu, T., 1029, 1201 Kusayanagi, E., 1817 Kusevic, I., 1557 Kushida, T., 1277
Maki, N., 1305, 1309, 1313 Makida, Y., 119, 419, 867 Mamalis, A. G., 937 Manzoor, S., 1517 Mao, C., 1421 Mao, C. B., 1417 Mao, D., 1783 Marechal, J. L., 689 Marque, S., 1911 Marti, H. P., 863 Martinez, A., 435 Maruno, Y., 597 Masada, E., 1041, 1079 Masashi, N., 335 Masatomi, H., 399 Masegi, T., 1681, 1707 Masuda, T., 971, 1057, 1347 Landgral', R., 283 Masulnoto, T., 1121 Lang, H. P., 1627 Masuzaki, S., 735 Le Lay, L., 1541 Matsubara, Y., 291,295, 303, 319 Lebrun, Ph., 91, 95, 195, 199, 443 Matsuda, H., 67, 71, 83 Lehmann, W., 489 Matsui, K., 127, 131, 135, 493, 767, LHD Group, 63, 75, 79, 83, 87, 731, 771,775, 779, 795, 1301, 1681 751, 1825 Matsui, T., 299 Li, B. Z., 847 Matsukawa, M., 1073, 1841 Li, J. N., 1037 Matsukura, N., 1711 Li, L.-Z., 1269 Matsumoto, K., 1095, 1129 Li, R., 339, 351, 355 Matsuo, M., 1053 Li, S., 359 Matsuo, S., 403 Li, X. Y., 1037 Matsuoka, S., 1459 Li, Y.-Y., 677 Matsushita, T., 701, 1795, 1803 Li, Y.-Z., 287, 497, 523, 547, 693 Matsuzawa, H., 831 Li, Z., 581 Mawatari, Y., 1529 Li, Z.-M., 461 Mayaux, C., 485 Li, Z.-Z., 613 Mayri, C., 111, 123 Liang, J.-T., 271 Maytal, B.-Z., 635 Lierl, H., 147 Mazaki, H., 1537 Lin, L. Z., 847, 1513 Mazurenko, O. N., 1941, 1945, 1949, Lin, X.-J., 871,875 1953 Liu, H. K., 1037, 1353, 1393, 1397, McIntyre, P., 1235 1545, 1549, 1557 Melaaen, E., 99 Liu, H. L., 1037 Meslmani, Y., 1533, 1579 Liu, J. Y., 1037 Meuris, C., 563 Liu, L.-Q., 225, 229 Michael, P., 1215, 1219, 1223 Liu, R.-M., 677 Michael, P. C., 1231 Liu, X.-Y., 2065 Mikawa, M., 1197 Liu, Z. Y., 1037 Mikumo, A., 1661, 1673, 1677 L6hlein, K., 195 Miller, J. R., 1891 Lokken, O. D., 363 Mimori, K., 419 Lounasmaa, O. V., 27 Mimura, M., 979, 1405, 1575, 1605 Lu, X. Y., 1431 Minami, H., 1083 Luciano, M., 983 Minami, M., 1961, 1965 Lfiders, K., 523, 547, 693, 1561, 1627 Minato, T., 771 Luo, E., 271 Mine, S., 863 Lutset, M., 1341 Minemoto, T., 1499 Minervini, J., 763, 1215, 1219, 1223 Machi, T., 1463 Minervini, J. V., 803 Machida, A., 159 Minot, F., 689 Maeda, M., 461, 621 Misaki, Y., 1197 Maeda, H., 431, 1089, 1099, 1103, Mitchell, N., 763 1377, 1385, 1735, 1757 Mitchell, N. A., 1903 Maehata, K., 1163 Mitin, V., 653 Maekawa, R., 75, 79, 83, 87, 481,735, Mito, T., 63, 75, 79, 83, 87, 605, 735, 739, 743, 1227 739, 743, 747, 751, 1163, 1227, 1253, Maezono, K., 1495 1337, 1825 Maix, R., 2001 Mitrohin, V., 799
Author Index Mitsubori, H., 1109 Mitsui, H., 1273, 2009 Mitsumoto, T., 855, 859 Miura, A., 331 Miura, K., 331,367 Miura, O., 1821 Miura, Y., 775, 1265, 1301 Miyaike, K., 905 Miyaji, T., 751 Miyake, A., 131,203, 743 Miyashita, K., 1851, 1863 Miyatake, T., 1095, 1711, 1735, 1739 Miyauchi, Y., 427 Miyazaki, T., 1095, 1711, 1739 Miyoshi, K., 979 Mizumaki, S., 419 Mizusaki, K., 1443 Mizutani, Y., 831 Moca6r, P., 1321 Mogi, I., 1121 Mohr, D., 423 Momal, F., 199 Mori, K., 1305, 1499 Mori, M., 131,203 Mori, S., 457 Moriai, H., 751 Morimoto, H., 267, 1017 Morisaki, T., 735 Morishita, H., 367 Morita, H., 811, 1261 Morita, M., 1125, 1289, 1731, 1771 Morita, Y., 183, 851, 1249 Moriuchi, S., 75, 79, 83, 87, 735 Moriya, T., 1763 Moriyama, H., 1273, 2009 Morra, M. M., 1903 Motojima, O., 63, 67, 75, 79, 83, 87, 673, 725, 735, 739, 743, 747, 751, 1159, 1285, 1825 Motokawa, M., 1121 Mtick, M., 283 Mukai, E., 941,953 Mukai, H., 1413 Mukoyama, S., 979 Mfiller, K.-H., 1353, 1545, 1549 Munshi, N. A., 1997 Murai, K., 79 Murai, S., 783 Murakami, M., 251, 255, 531, 543, 575, 589, 1439 Murakami, T., 2045 Murakami, Y., 1041, 1205, 1209, 1631, 1783 Murase, S., 1681, 1707, 1795 Murata, Y., 811 Muta, I., 941,953 Nabatame, T., 1685, 1689 Nadi, R., 1021, 1621 Nagai, K., 1957 Nagai, T., 1731, 1771 Nagamura, H., 909 Nagano, M., 1537 Nagata, A., 1431 Nagata, M., 987, 1749
Nagaya, S., 701, 971, 987, 995, 999, 1381, 1601 Nakade, M., 1029, 1033 Nakagawa, H., 1923 Nakagawa, M., 987, 999, 1381, 1601, 1701 Nakagome, H., 319, 501, 1033, 1113 Nakahara, S., 665 Nakahira, A., 2041, 2057 Nakahira, M., 427 Nakai, H., 183, 851, 1249 Nakajima, H., 767, 771,775, 783, 795, 1665, 1887, 1895 Nakamoto, K., 739 Nakamoto, T., 419 Nakamoto, Y., 331 Nakamura, H., 913, 917, 921 Nakamura, K., 79, 83, 473, 1333 Nakamura, M., 263, 461, 567, 621, 1455, 1525 Nakamura, N., 251,255 Nakamura, T., 1567, 1587 Nakane, H., 701 Nakanishi, K., 755, 759 Nakaniwa, T., 1895 Nakao, H., 331 Nakayama, S., 885, 891, 1707 Nakayama, Y., 127 Nara, K., 1929 Narayankhedkar, K. G., 377, 373 Natori, N., 1297 Natu, P. V., 373 Nemoto, T., 1993 Nemoto, Y., 1757 Neo, S., 1041 Neubert, J., 509 Neuenschwander, J., 863 Nicoletti, A., 879 Nicollet, S., 435 Nii, A., 1025 Niihara, K., 2041, 2057 Nijhuis, A., 1243 Ninomiya, A., 735, 1069, 1337 Nishi, M., 1731, 1771 Nishida, K., 127, 131, 135, 427, 493, 767, 795 Nishigaki, K., 513, 597, 1813 Nishiguchi, K., 1173 Nishihara, R., 1505 Nishijima, S., 19, 513, 1273, 1277, 1989, 2013, 2035, 2041, 2049, 2057, 2061 Nishikawa, M., 891 Nishimura, A., 735, 739, 743,751,755, 759, 1877 Nishimura, I., 203 Nishimura, K., 735 Nishioka, T., 1163 Nishitani, T., 263, 307 Nishiura, T., 1989, 2013 Nishiya, T., 941,953 Nisiwaki, Y., 1065 Nitta, I., 1961, 1965 Nitta, J., 1193 Nitta, T., 891,909, 1065 Nobutoki, M., 71
I5
Nogawa, S., 1065 Noguchi, T., 513 Nojima, K., 2035, 2061 Noma, K., 2005 Nomura, H., 1065, 1297 Nomura, K., 1389 Nomura, S., 1029, 1113, 1707 Nonaka, S., 1357 Norris, B. L., 179 Nose, S., 1009, 1049, 1053, 1151, 1325, 1329 Noto, K., 1073, 1841 Nozaki, S., 1061 Nozawa, M., 791,795, 1265 Numazawa, T., 2069, 2073 Nunoya, Y., 767, 771, 775, 779, 795, 1665, 1731, 1771, 1895 Oba, K., 79 Obara, H., 1529 Obert, W., 485 Ochi, T., 399 Oellrich, L. R., 239 Ogasawara, M., 1471 Ogasawara, T., 1763, 1779 Ogata, H., 775, 783 Ogata, T., 1791, 1899, 1915, 1923 Ogawa, H., 735 Ogawa, R., 1103, 1129, 1409, 1711, 1799 Ogawa, S., 1201, 1591 Ogino, O., 513 Ogiso, K., 1041 Ogitsu, T., 183, 419, 843 Ogiwara, H., 513 Ogushi, T., 399, 1443, 1491, 1495 Ohashi, Y., 299 Ohba, K., 87, 735 Ohhata, H., 419 Ohira, K., 403, 601 Ohkita, S., 1895 Ohkuma, T., 1029, 1033, 1413 Ohkura, K., 1057, 1347, 1365 Ohmatsu, K., 967, 975, 1347, 1799, 1855, 1859 Ohno, I., 427, 743 Ohsaki, H., 1079 Ohsaki, O., 783 Ohska, T., 673 Ohtake, I., 87, 735 Ohtani, Y., 319, 501, 1113 Ohtsu, K., 493 Ohtsuka, H., 1957 Ohuchi, N., 419, 843 Okada, H., 1049 Okada, K., 473 Okada, T., 909, 1273, 1277, 1989, 2013, 2035, 2041, 2049, 2057, 2061 Okaguchi, S., 1915, 1923 Okaji, M., 1929 Okamoto, M., 331 Okamura, T., 539 Okano, M., 593, 957 Okazaki, O., 513 Okubo, H., 319, 1837 Okubo, K., 1009, 1325, 1329
I6
Author Index
Okumura, H., 673 Okuno, H., 855, 859 Okuno, K., 763, 791 Onabe, K., 995 Onishi, A., 339, 351,355 Onishi, T., 1025 Ono, M., 739, 987, 1727 Onoda, H., 1361 Onodera, T., 1841 Ooba, K., 75 Ootsu, K., 127 Ootuka, Y., 457 Osaki, O., 767 Osaki, K., 1693 Osamura, K., 1357, 1787, 1791, 1795, 1799 Oshikiri, M., 775, 1665 Ostler, J., 837 Otabe, E. S., 701, 1803 Otsuka, M., 791, 1985 Owren, G., 99 Ozaki, O., 1095, 1129, 1293 Ozeki, M., 13 Pal3vogel, Th., 707, 715 Pai, Chien-ih, 871 Pai, C., 875 Pailler, P., 111 Pan, H.-Y., 359 Panek, J., 571 Papavasiliou, N., 657 Passardi, G., 111, 123, 173 Passvogel, Th., 711 Patel, L. N., 377 Pavese, F., 1499 Peltier, F., 1321 Penny, M., 1541 Pe6n, G., 443, 477 Perini, D., 837 Peshkov, I., 799 Petersen, K., 423, 707 Pfotenhauer, J. M., 363 Plashkin, E. A., 1937 Pradhan, S., 819, 823, 827 Proyer, S., 1533 Pylinina, S. N., 929, 1933 Pyon, T., 1715 Qiao, G.-W., 1475 Qiu, L.-M., 247, 315 Qiu, M., 1513 Qiu, N., 151 Quack, H., 395, 465 Quan, H.-Y., 233 Radebaugh, R., 33 Raju, K. S. N., 949 Randall, R. N., 1903 Ravikumar, K. V., 657 Reed, R. P., 2017 Rehak, M. L. F., 879 Riddone, G., 95, 443, 477, 681 Rieder, H., 107 Rieubland, J. M., 173 Rodriguez Mateos, F., 1871 Rogalla, H., 391, 1181
Rohleder, I., 1665 Rousset, B., 519 Ruppert, U., 523, 547, 693, 945 Rychagov, A., 799, 1969 Ryouman, A., 1057 Sadakata, N., 933,987, 995,999, 1003, 1013, 1073, 1381, 1601, 1701, 1723, 1841, 1867 Saga, N., 967, 975, 1347 Sagner, U., 707 Saho, N., 1125 Sahu, A. K., 819 Saito, A., 1197 Saito, K., 1807 Saito, M., !915, 1923 Saito, T., 995, 1003, 1013, 1073, 1189, 1723, 1867 Saitoh, T., 933, 987, 999, 1381, 1601, 1701, 1795 Saji, N., 203, 207, 743 Sakagami, Y., 1681 Sakai, K., 1205, 1209, 1631 Sakai, S., 1669, 1685, 1689, 1693, 1719, 1791, 1799, 1851, 1863 Sakai, Y., 1089 Sakaki, K., 135, 795, 1053, 1151 Sakakibara, S., 735 Sakamoto, H., 913, 917, 921 Sakamoto, N., 1443, 1561 Sakamoto, Y., 183 Sakiyama, H., 665 Sakuma, S., 539 Sakuraba, J., 1109, 1121 Sakurai, A., 535, 585, 617 Salunin, N. I., 1937 Samadi Hosseinali, G., 1533, 1579 Samoto, K., 1369 Sampson, W. B., 863 Sanada, K., 1887, 1981 Sander, M., 423, 707, 715 Sang, I.-Y., 1439 Sapozhnicov, V. A., 929 Sarkar, B., 819 Sarrhini, O., 1479 Sasaki, K., 843, 1685 Sasaki, T., 771,783, 2009 Sashida, T., 1347 Sata, K., 1173 Satisha, G. V., 413 Sato, A., 431, 513, 1103, 2073 Sato, J., 203, 1389 Sato, K., 617, 913, 917, 921,967, 975, 991, 1009, 1057, 1155, 1329, 1347, 1365, 1413, 1583, 1597, 1673, 1677, 1855, 1859 Sato, M., 2069, 2073 Sato, S., 1017, 1053 Sato, T., 1041 Satoh, S., 63, 67, 75, 79, 83, 87, 481, 735, 739, 743 Satoh, T., 339, 351,355 Satoh, Y., 1049 Satou, K., 1405 Satow, T., 735, 739, 743, 747, 751, 1825
Sauerzopf, F. M., 1579 Sawa, A., 1529 Sawa, F., 1273, 2057 Saxena, Y. C., 819, 823, 827 Scanlan, R., 1235 Scanlan, R. M., 1743, 1767 Schauer, F., 815 Schultz, J., 1219, 1223 Schultz, J. H., 1231 Schultzand, J., 1215 Schumann, B., 505 Schupp, J., 715 Schustr, P., 195 Seeber, B., 863 Segawa, T., 1073 Seidel, A., 423, 707, 711,715 Seidel, P., 259 Seidler, M., 489 Seido, M., 751 Seki, N., 267 Sekiguchi, H., 75, 79, 87, 735 Sekiguchi, S., 127, 493 Sekine, S., 1297 Sekino, T., 2041 Semeonov, I., 799 Senba, A., 1079 Senba, T., 751, 1285 Seo, K., 251,255, 1289 Seppala, J., 863 Sergeyev, I. A., 609 Sergo, V., 199 Serio, L., 91 Serries, J. P., 435 Sgobba, S., 1911 Shamoto, Y., 673 Shavit, A., 635 Shen, M., 677 Shen, S., 1231, 1665 Shevchenko, O. A., 803 Shi, W., 2027 Shibata, K., 1883 Shibata, T., 967, 975, 1347 Shibutani, K., 1129, 1409 Shibuya, J., 775 Shieh, T. F., 385 Shiga, N., 1707 Shigematsu, T., 461, 567, 621, 1525 Shigenaka, A., 811 Shigi, T., 263, 461, 567, 621, 1525 Shimada, M., 1095, 1129, 1409, 1711, 1735, 1739, 1791 Shimakage, H., 1643 Shimamoto, S., 127, 131, 135, 427, 763, 767, 771, 775, 779, 787, 791, 807, 1301, 1665, 1673, 1677 Shimamura, K., 1707, 2073 Shimazaki, T., 575 Shimizu, K., 1069 Shimizu, T., 1727, 1907 Shimonosono, T., 971, 987, 995, 999, 1381, 1601, 1883 Shir.do, Y., 1887, 1981 Shingai, K., 1483 Shinohara, H., 1329 Shintomi, T., 419, 843,851, 1249, 1767 Shiohara, Y., 1455, 1459, 1509
Author Index Shioiri, T., 1239 Shiotsu, M., 535, 585, 617 Shiraishi, M., 251,255 Shoji, M., 673 Siegel, N., 837 Siemko, A., 837 Sigaev, V. E., 929, 1933, 1937 Simamoto, S., 1265 Simon, N. J., 2017 Singo, S., 1583 Sinha, B., 643 Sirot, E., 1321 Sirotko, D. V., 925 Skoczen, B., 1911 Smirnov, A., 139 Smith, B. A., 803 Smith, K., 1143 Smith, R. P., 863 Snydstrup, L. P., 871,875 Sobol, V. R., 1941, 1945, 1949, 1953 Solheim, N., 103 $611, M., 2001 Sotojima, T., 381 Specking, W., 1661 Spiebberger, S. M., 1997, 2005 Sp6rl, G., 509 SST Team, 819, 823, 827 Stamm, M., 489 Stangl, E., 1533 Starchl, B., 1579 Straif, W., 1579 Su, X.-D., 1475 Sfi/3er, M., 469 Sudo, S., 685 Suehiro, J., 1009 Suekane, T., 539 Sueyoshi, T., 1505 Suganomata, S., 831 Sugawara, K., 1431, 1753 Sugawara, S., 419 Sugimoto, M., 767, 771,775, 779, 783, 791,795, 913, 917, 921, 1605, 1665, 1673, 1677, 2009 Sugiura, T., 1571, 1841 Sugiyama, K., 1851, 1863 Sukhanova, A., 139 Sulten, P., 1321 Sumida, M., 1459 Sumita, T., 1749 Sumiyoshi, F., 921, 1049 Sumiyoshi, Y., 783, 2009 Summers, L. T., 1891 Sumption, M. D., 1609, 1767 Sun, T., 275 Sun, X. Y., 1417 Sunada, K., 629 Suraci, A., 91 Suryanarayana, T., 949 Suzawa, C., 617, 967, 975 Suzuki, H., 735 Suzuki, K., 905, 2045 Suzuki, M., 1451, 1487, 1639, 1753 Suzuki, N., 2045 Suzuki, S., 751,755, 759 Suzuki, T., 1239, 1685, 1689, 1985 Suzuki, Y., 1201
Svalov, G., 1969 Sytnikov, V., 799, 1969 Szalay, A., 937 Szeless, B., 443 Szeless, B., 1871 Szfics, Z., 945 Tachikawa, K., 1427, 1471, 1863 Tada, N., 1693 Taguchi, O., 1731, 1771 Taira, M., 1451, 1639 Takabatake, K., 1095, 1103, 1293 Takfics, S., 1253, 1257 Takagi, T., 1129 Takahashi, K., 1525 Takahashi, C., 1073 Takahashi, K., 1155, 1525, 1673, 1677, 1855, 1859 Takahashi, M., 1029, 1033 Takahashi, R., 811, 1261, 1317 Takahashi, T., 851, 1073 Takahashi, Y., 767, 771,775, 779, 783, 795, 811, 1301, 1665, 1681 Takahata, K., 79, 605, 735, 739, 743, 751, 1227, !253 Takano, K., 795, 1895 Takano, S., 1643 Takao, T., 1961, 1965, 1993 Takashi, I., 335 Takaya, Y., 1265 Takayanagi, H., 1193 Takayasu, M., 1215, 1219, 1223, 1231 Takebayashi, S., 1439 Takeda, M., 597, 1813 Takeo, M., 735, 1009, 1017, 1049, 1053, 1163, 1325, 1329, 1567, 1587, 1829, 1833 Takeshima, H., 1305, 1309, 1313 Takeshita, M., 621 Takeuchi, T., 1669, 1697, 1735, 1757 Takeuchi, Y., 535 Takigami, H., 913, 917, 921 Takita, K., 1159 Tallerico, T., 871,875 Tamada, N., 593, 957, 1297 Tamaki, T., 751,755, 759 Tamura, H., 735, 739, 755, 759 Tanahashi, S., 735, 739, 743 Tanahasi, S., 747 Tanaka, Y., 1385 Tanaka, A., 1361 Tanaka, H., 1833 Tanaka, K., 119, 419, 649, 843 Tanaka, M., 307 Tanaka, S., 3 Tanaka, T., 1883 Tanaka, Y., 855, 859, 979, 1377, 1385, 1405, 1575, 1605, 1787, 1799 Taneda, M., 127 Taneya, S., 331 Tang, H., 1475 Tang, X., 1373 Tang, Z.-M., 2027 Tani, M., 1205, 1209, 1631 Tanida, K., 303 Taniguchi, T., 1317
I7
Tanna, V., 819 Taran, A., 799 Tasaki, K., 1361 Tatara, I., 1739 Tateishi, H., 1297, 1701 Tavian, L., 95, 195, 199, 681 Tei, C., 1821 ten Haken, B., 697 ten Kate, H. H. J., 697, 803, 1243, 1743 Teng, M., 1143 ter Brake, H. J. M., 391, 1181 Terai, M., 331 Teramachi, Y., 673 Terasawa, A., 767, 771,791, 1265 Terashima, A., 419, 843, 851, 1249 Teuho, J., 863 Thome, R., 763 Thummes, G., 283, 311 Thfirk, M., 259 Timms, K., 1117 Tischhauser, J., 173 Titus, P. H., 1903 Tkhorik, Y., 653 Tobler, R. L., 1877 Tochihara, S., 1537 Toda, H., 1813 Togano, K., 1757 Tokunaga, M., 1749 Tominaga, A., 243 Tominaka, T., 855, 859, 1305 Tomioka, A., 1053, 1151 Tomioka, K., 331 Tommasini, D., 837 Tomozawa, S., 1205, 1631 Tonouchi, M., 1205, 1209, 1631, 1643 Torii, H., 367 Torii, S., 933 Toyoda, K., 905 Triscone, G., 1571 Troell, J., 453 Tschegg, E. K., 1997, 2005 Tsubouti, H., 979 Tsuchiya, K., 419, 843, 867 Tsugawa, K., 1297 Tsuji, H., 127, 131, 135, 427, 493, 763, 767, 771, 775, 779, 783, 787, 791, 795, 807, 811, 1265, 1301, 1661, 1665, 1673, 1677, 1731, 1771, 1887, 1895, 2009 Tsukamoto, H., 791, 1985 Tsukamoto, O., 1041, 1617, 1775 Tsukamoto, T., 1571 Tsukasaki, Y., 1989, 2013 Tsukiji, H., 941,953 Tsukiyama, M., 941,953 Tsuru, K., 1487 Tsutsumi, K., 1017, 1049, 1151 Tu~ek, L., 195 Tutaev, V. A., 929, 1933 Uchaikin, S. V., 439 Uchida, T., 2045 Uchikawa, F., 1731, 1771 Ueda, H., 1495 Ueda, N., 1083
I8
Author Index
Uede, T., 735, 743, 747, 1151 Uehara, M., 1957 Ueki, T., 2035, 2061 Ueno, S., 2041, 2057 Ueyama, M., 1009, 1329, 1347, 1365 Ulbricht, A., 469 Umeda, T., 203, 457, 1459 Uno, N., 979, 1405, 1575 Uno, S., 1057 Unoki, H., 1463, 1467 Urata, M., 1029, 1033, 1361, 1707 Uriu, Y., 1337 Usami, K., 1189 Usami, S., 1985 Ushigusa, K., 1681 Ushijima, I., 67, 83 Ushijima, M., 783 Utaka, Y., 1169 Uzawa, Y., 1185 Vajda, I., 937 Vallier, J. C., 435 Valthe, S., 2005 van Weelderen, R., 91, 519 van Oort, J. M., 1743 Van Sciver, S. W., 527, 571 Vanni, O., 983 Vanolo, M., 1499 Vecsey, G., 1665 Veldhuis, D., 1181 Venger, E., 653 Vieira, R., 2009 Vins, M., 195 Violet, J. L., 689 Vo, N. V., 1397, 1609 von Schoenebeck, F., 2023 Vullierme, B., 199 Vysotsky, V., 1215, 1219, 1223, 1231 Wachi, Y., 743 Wada, H., 431, 1089, 1099, 1103, 1783 Wada, K., 1427 Wada, N., 1209 Wadahl, A., 99 Wadayama, Y., 811, 1261 Wade, M., 1117 Wagner, A., 423 Wagner, R., 259 Wagner, U., 95, 99, 1139, 1143 Wakabayashi, H., 1301 Wakamoto, K., 1731, 1771 Wakasugi, K., 1749 Wakata, M., 1731, 1771 Wake, M., 867, 1235, 1767, 1807 Wakita, M., 1837 Wakuda, T., 1329 Walckiers, L., 837 Walker, E., 1571 Walker, R. J., 143 Walsh, R. P., 661, 1891 Wang, D., 677 Wang, F. T., 847 Wang, J.-R., 151 Wang, K.-G., 1447 Wang, L., 613 Wang, Q. L., 847, 1269
Wang, R.-S., 581 Wang, R.-Z., 555, 2027 Wang, W. G., 1353, 1393 Wang, W.-Y., 719 Wang, X.-X., 287 Wang, Y.-Q., 1269 Wang, Y.-Z., 1475 Wang, Z., 1185, 1205, 1209, 1631, 1643 Warnes, W., 1231 Watanabe, M., 975 Watanabe, I., 791 Watanabe, K., 79, 735, 1109, 1121, 1365, 1669, 1693, 1707, 1719, 1723, 1841 Watanabe, M., 457, 967, 975 Watanabe, N., 1045 Watanabe, Y., 127 Watazawa, K., 1109, 1121 Weber, J., 423 Weber, H. W., 1533, 1579, 1997, 2005 Weise, F., 2031 Welton, S. J., 527 Wen, J.-G., 1463 Wessel, S., 803 Wild, S., 239 Will6n, D. W. A., 1021, 1621 Williams, L. R., 477 Williamson, J., 657 Winkler, G., 103 Wolf, J., 423, 707 Wong, F. M. G., 1903 Wu, J.-Y., 555 Wu, J.-Y., 2027 Wu, X., 1373, 1421 Wu, X. Z., 693, 1417, 1447 Wu, Y., 217 Wu, Y.-Y., 547, 693 Wu, Y.-Z., 547, 693 Wtichner, F., 469 Wykes, M. E. P., 131,427 Xia, Z.-M., 247, 315 Xu, C., 1451 Xu, J.-M., 625, 1635 Xu, L., 271,625, 1635 Xu, R.-L., 677 Xu, X., 473 Xu, X.-D., 343, 347 Yabu-uchi, K., 399 Yaegashi, N., 1915, 1923 Yagi, N., 381 Yamada, H., 79, 87 Yalnada, N., 1929 Yamada, R., 863 Yamada, S., 63, 75, 79, 83, 87, 735, 739, 743, 747, 751 Yamada, Y., 1109, 1121, 1361, 1427, 1455, 1661, 1665, 1673, 1677 Yamafuji, K., 1009, 1325, 1329, 1567, 1587 Yamagata, Y., 1483 Yamaguchi, K., 905, 1017 Yamaguchi, M., 263, 461, 567, 589, 621, 1525
Yamaguchi, S., 673, 735, 751, 1159, 1285 Yamaguchi, T., 331 Yamamoto, A., 111, 119, 419, 649, 843, 867, 1915, 1923 Yamamoto, J., 63, 67, 75, 79, 83, 87, 605, 673, 731, 735, 739, 743, 747, 751, 755, 759, 1163, 1227, 1253, 1337, 1361, 1825, 1841, 1877 Yamamoto, K., 1029 Yamamoto, M., 263, 1013 Yamamoto, N., 1915 Yamamoto, S., 1133, 1289 Yamamoto, T., 743 Yamamoto, Y., 1525 Yamamura, H., 127 Yamanaka, A., 1993, 2049 Yamaoka, H., 419 Yamasaki, H., 1529 Yamasaki, S., 267, 1025 Yamashita, F., 1427 Yamazaki, K., 735 Yamazaki, T., 1681 Yamazumi, T., 1617 Yanagi, H., 159 Yanagi, N., 605, 735, 739, 743, 751, 1253, 1285, 1825 Yanagi, Y., 211 Yanagise, N., 531 Yanai, M., 263, 307 Yanaka, S., 783 Yang, J., 625 Yang, Z.-Q., 1475 Yano, Y., 855, 859 Yao, H., 217, 221,233, 497 Yasohama, K., 1763, 1779 Yasuda, M., 1431 Yasuda, T., 1643 Yasukawa, Y., 135, 795, 1009 Yasunaga, T., 1553 Yasuoka, H., 1537 Yatsuda, T., 1505 Yazaki, T., 243 Yazawa, T., 1029, 1113 Ye, J., 1635 Yin, Z.-Z., 639 Yokogawa, K., 1915, 1919, 1923 Yokoyama, S., 1133 Yoneda, E. S., 913, 917, 921 Yonenaga, Y., 747 Yoshida, K., 131, 135,811, 1731, 1771 Yoshida, N., 991, 1155, 1413, 1583, 1597, 1847, 1855, 1859 Yoshida, S., 457, 657 Yoshida, T., 739 Yoshikawa, K., 319, 481 Yoshikawa, M., 1095, 1293 Yoshimura, N., 319 Yoshinaga, S., 203 Yoshino, Y., 1907 Yoshitomi, J., 933, 991, 1003, 1867 Yoshizaki, R., 1385, 1435 Yoshizawa, S., 701 Yotsuya, T., 1201, 1591 Yu, J.-P., 247, 315, 407, 677 Yumura, H., 1855, 1859
A u t h o r Index Yuri, T., 1899 Yuyama, J., 267, 1083 Yuyama, M., 1735, 1783 Zadro, K., 1557 Zeng, D., 1475 Zeng, Z. J., 1037 Zhang, C., 1475
Zhang, C.-Q., 271 Zhang, L.-A., 343, 347, 613 Zhang, P., 555 Zhang, P.-X., 1447 Zhang, Z.-Y., 343 Zhao, L., 247, 275, 315 Zheng, J.-Y., 247, 275, 315, 677 Zheng, X. G., 1451, 1639
Zhou, G.-L., 229 Zhou, L., 1373, 1417, 1421 Zhou, L.-A., 1447 Zhou, S.-L., 625, 1635 Zhou, Y., 271 Zhu, S.-W., 291 Zhu, W., 1021
I9
