doctoral degree defense
DESCRIPTION
Doctoral Degree Defense. A MINIATURE REVERSE-BRAYTON CYCLE CRYOCOOLER AND ITS KEY COMPONENTS: HIGH EFFECTIVENESS HEAT RECUPERATOR AND MINIATURE CENTRIFUGAL COMPRESSOR. Defender: Lei Zhou Advisor: Dr. Louis C. Chow Dr. Jay Kapat - PowerPoint PPT PresentationTRANSCRIPT
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Doctoral Degree Defense
Defender: Lei ZhouAdvisor: Dr. Louis C. Chow
Dr. Jay KapatCommittee members: Dr. Louis C. Chow; Dr. Jay Kapat; Dr. Q.
Chen; Dr. R. Chen; Dr. Larry AndrewDepartment of MMAE
University of Central FloridaNOV 10th,2003
A MINIATURE REVERSE-BRAYTON CYCLE CRYOCOOLER AND ITS KEY COMPONENTS: HIGH EFFECTIVENESS HEAT RECUPERATOR
AND MINIATURE CENTRIFUGAL COMPRESSOR
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Research Background
• Applications– Oxygen/Nitrogen liquefaction– Infrared image sensor array– Electronic device cooling– Out space exploration– HTS (High temperature superconductor) cooling
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General refrigeration cycle
H eat re jection Q hto am bient at T h
H eat absorption
Q c watts at T c
Fluid expansionto reduce
tem perature
W ork done onprocess flu id
Power in v iacom pressoror drive unit
Q H X
H TS load at ~T c
ch
cCarnot TT
T
Carnotideal
1COP
)3.01.0(
COPCOP ideal
real
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General COPOPERATING
TEMPERATURE CARNOT
COP (Watt Input per Watt
Lifted)
"TYPICAL" COP FOR >100 WATT HEAT LOADS
(Watt Input at 300 K per
Watt Lifted at Top) 273 K 0.11 ~ 0.4
200 K 0.52 ~ 2 150 K 1.01 ~ 4 100 K 2.03 ~ 8-10 77 K 2.94 ~ 12-20 50 K 5.06 ~ 25-35 40 K 6.58 ~ 35-50 30 K 9.10 ~ 50-75
Treject = 303 K
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Major Cryogenic Technologies
• Stirling machine
• Pulse tube
• Gifford-McMahon
• RTBC (reverse Turbo-Brayton Cycle)Technology Pros Cons
S High efficiency, compact Vibration, unreliable
P Compact, reliable, no moving parts Efficiency lower than Stirling
G Simple, reliable Bulky, gas purity sensitivity, inefficient
R Compact, no vibration, efficient Moving part
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Cycle efficiency vs. compressor electrical power
0
5
10
15
20
25E
FF
ICIE
NC
Y
(% o
f C
arno
t)
10 10 10 10 100 1 2 3 4
COM PRESSO R INPUT POW ER (W )
TRW
TRW
LM
LM
TRW
NIS T
4-valve
Turbo-B ray ton
M ixed -gas JT
M ixed -gas JT80 K
90 K
G iffo rd -M cM ahon
0.5 W@ 80 K
LM
NIS T
TRW
TRW
LM
TRW
P u lse Tube (Stirling -type )
Stirling
G iffo rd -M cM ahon
P u lse Tube(G M -type )
C ryocoo lerE ffic iencyT = 80 K
1 W
10 W
50 W
100 W
5 W Activebuffer
Turbo-B rayton
85%
85%
85%
85%
Stirling
JPLLM
Courtesy of Ray Radebaugh, NIST-Boulder
Proposed miniature
RTBC
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Cycle efficiency vs. operating temperature
1 5 10 50 100 300
EF
FIC
IEN
CY
(%
Car
not)
TEM PE R ATU R E (K )Radebaugh 2001
C ryocoo ler E ffic iency
0
5
10
15
20
25
30
Turbo-B ray ton
M ixed -gas JT
P u lse Tube (Stirling -type )
Stirling
G iffo rd -M cM ahon
P u lse Tube(G M -type )
(<10 kW com pressor input)
Courtesy of Ray Radebaugh, NIST-Boulder
Proposed miniature
RTBC
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Cryocooler Applications and Operating Regions
Proposed miniature
RTBC
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RTBC concept
2
turbine
6
generator compressor
motor
Heat exchanger to Ambient
Heat Load
1
34
5Heat regenerator
COP=Heat removed from heat load end / Power input to
system
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RTBC Mollier Diagram
Work in
Work out
Heat exchange
Heat in
Heat out
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Miniature RTBC cryocooler
• Proposed cooling power: 20Watt at 77K
• Proposed COP: 0.08~0.1
• Miniature size– Miniature single stage mixed flow centrifugal
compressor– Micro channel heat recuperator– Integrated high efficiency motor/alternator– Advanced air-foil bearings
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Advantages of miniature RTBC cryocooler
• Portability
• Suitable for weight/size critical applications
• Simplicity
• Low maintenance
• Low cost
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Miniature cryocooler concept
m cm mmm
Micro scale Meso scale Macro scale
10mW W 0.1kW
Poor COP Good COP Best COP
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Thermal efficiency analysis of miniature reverse-Brayton cycle(1)
),,,,,,( 411 prdTEETPTfCOP tc
1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.8 3.0 3.20.00
0.05
0.10
0.15
0.20
0.25
0.30
0.35
DT1.5 DT1.7 DT1.9 DT2.1 DT2.3 DT2.5 DT3.0 DT4.0
CO
P
Pressure Ratio
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Thermal efficiency analysis of miniature reverse-Brayton cycle(2)
1.5 2.0 2.5 3.0 3.5 4.00.235
0.240
0.245
0.250
0.255
0.260
0.265
0.270
0.275
0.280
COP
CO
P
Delta T (K)
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Thermal efficiency analysis of miniature reverse-Brayton cycle(3)
0.60 0.65 0.70 0.75 0.80 0.85 0.90
0.10
0.15
0.20
0.25
0.30
COP
CO
P
Ec,Et
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Result of thermal efficiency analysis:system parameters
2
turbine
6
generator compressor
motor
Heat exhausted=261W
Cooling Load=20W
1
34
5Heat regenerator
COP=0.083
Eff=0.993
T1=64K
T5=300.2K
T6=76.0K
T2=74.4K
T3=299.5K
T4=440K
Pmotor=262W
Pressure ratio=1.75
Mass flow rate=2.81g/s
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Micro-channel heat recuperator
s
d
dw
L
Insulated surface
Thot,in
Tcold,in
Hot end Cold end
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Stacked multi-layer construction
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Physical model
Cold Neon
Hot Neon
d
d
wY
X
Z
0)]()([4
0)]}()([)]()([{
0)]()([4
2
2
xTxTdx
Tcm
xTxTxTxTdx
TkA
xTxTdx
Tcm
whhh
p
hhccw
sw
wccc
p
30PrRe
nk
dUCpPe
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Numerical Model (1-D)
Hot gas node
Cold gas node
Wall
Interface
HjHj+1
Cj Cj+1
Wj Wj+1WtMetal
Material
Insulation material
Hot fluid
Cold fluid
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Numerical simulation for single material
Fig.5 axial heat conduction in wall
0.000 0.005 0.010 0.015 0.02080
100
120
140
160
180
200
220
240
260
hot duct wall cold duct
Tem
pera
ture
(K)
Length (m)
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20 40 60 80 100 1202
3
4
5
6
7
8
9
10
dT
cold
end t
em
pera
ture
diffe
rence
dT
(K)
Length L (mm)
dt VS. Length (total temperature different =220K)
L
dU
L
m
kNuL
TCpmdT
n
2
4
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1-D Numerical for two material
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Comparison of heat conductivity
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Comparison of single material and two materials
Configuration 1+1 (1mm) 10+10 (10mm) 40+40 (40mm)
T(K)3 30 120
SiO20.377 0.848 0.853
Metal0.4005 0.5333 0.6198
Alternative Insulator/Metal
0.4003 0.937 0.991
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Conclusion of micro-channel heat recuperator design
• 1-D numerical simulation is suitable for the performance estimation of the micro-channel heat recuperator
• With proper parameter selection, the micro-channel heat recuperator can achieve 0.99 effectiveness at an acceptable pressure loss
• For the reason of manufacturing, this heat recuperator may be constructed as many thin layers stacked together. It provides the possibility of two materials (one have high heat conductivity and another have very low heat conductivity) stacked alternatively to provide 0.99 effectiveness.
• This simulation provides the guidance to select the material to manufacture the heat recuperator. LTCC may be a good candidate due to its low heat conductivity and high solidity after cured.
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Centrifugal compressor design• Advantages of single stage centrifugal compressor
– Simplicity: only 1 moving part– Reliability (better than reciprocating compressor)– Possible high efficiency– No vibration: high revolution speed (>>100 kRPM)– Compact
• Disadvantages:– Difficult design: complicated flow field– Relatively expensive: manufacturing rows of blades in
small size– Low compression ratio
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Testing Compressor specifications
• Working fluid: Nair
• Operating pressure: 1 bar
• Operating temperature: 300K
• Mass flow rate: 4.5 g/s
• Compression ratio: 1.7
• Bearing: conventional ball bearing
• Driver type: direct Motor
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Compressor Design flow chart
Basic layout design
Basic thermodynamics and sizing
Geometry design
1-D flow calculation
3-D CFD verification
Manufacturing and testing
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Basic Layout
Flow direction
•Radial IGV
•Mixed flow impeller
•Axial diffuser
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R-Z plane X-Y plane
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3-D geometry design ---- hub-shroud contour (R-Z plane)
IGV shroud curve
IGV hub curve
Impeller shroud curve
Impeller hub curve
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3-D geometry design ---- X-Y plane blade angle
0,0
Leading edge
Trailing edge
X-Y projection line of blade at
shroud/hub surface
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Geometry implementation in Pro/Engineer(1)
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Geometry implementation in Pro/Engineer(2)
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Geometry implementation in Pro/Engineer(3)
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Introduction of 2-zone model of impeller
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3-D view of IGV
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3-D view of diffuser
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Compressor assembly (1)
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Compressor assembly (2)
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3-D CFD geometry#
#: 3-D simulation results is provided by Xiaoyi Li
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3-D results
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3-D results
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CFD results
Flow Separation inertia force and centrifugal force
Suggestion reduce the length of IGV add deswirl vane
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Conclusion of compressor design
• 2-zone model is the most powerful 1-D design tool in centrifugal compressor design. With proper mathematics and interactive program codes, 3-D geometry can be designed and then implemented with pro/engineering software
• 3-D CFD simulation show the improvements should be done in next design. Mixed flow impeller with axial diffuser may have severe flow separation problem at the bending section. A deswirl vane is needed before this section
• Impeller may need to be refined with inducer to reduce entrance separation.
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Compressor Testing Run set up
Pressure and Temperature at Diffuser
Exit
Motor Case Temperature
Mass Flow Controller
Power InBearing Temperature
Power Out of Motor
Motor Bearing Temperature
Pressure and Temperature at Inlet
Motor Bearing Temperature
Bearing Temperature
Pressure and Temperature after Mixer
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Testing assembly (coupler improvement)Coupler design speed:
~30,000 RPM
Coupler with steel sleeve in test run
~97,000 RPM
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Testing assembly
EyeP, T sensor here
Impeller outT sensor here
Diffuser outP, T sensor here
To mass flow rate meter
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‘Blank Shaft’ Test•Motor efficiency = 40% to 70%
–90,000 rpm = 65% with load
• Loss per bearing = 105 Watts at 90,000 rpm
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Compressor test
• Curved Blade Impeller– 89,485 rpm, 3.13 g/sec, 2.70 psig
• Straight Blade Impeller– 93,984 rpm, 5.14 g/sec, 5.05 psig
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Compressor testPower versus Speed
0.0
200.0
400.0
600.0
800.0
1000.0
1200.0
1400.0
0 20000 40000 60000 80000 100000 120000
Speed (rpm)
Po
wer
(W
atts
)
Cast Impeller
StraightBladeImpeller Test1
StraightBladeImpeller Test2
StraightBladeImpeller w/DataAcquisition
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Compressor testGage Pressure Versus Speed
0.0
2.0
4.0
6.0
8.0
10.0
12.0
0 20000 40000 60000 80000 100000 120000 140000 160000
Speed (rpm)
Gag
e P
ress
ure
(p
sig
)
Cast Impeller
Straight BladeImpeller Test 1
Straight BladeImpeller Test 2
Straight BladeImpeller w/ DataAcquisition
Design Point
Theoretical points
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Compressor efficiency
• Actual output conditions:– 93,984 rpm
– 1.29 pressure ratio
– 61.2% isentropic efficiency
– 5.1 grams per second mass flow rate
Speed Impeller Work Mass Flow Raterpm Watts g/s50000 60 3.849 0.57160000 90 3.159 0.37670000 150 3.389 0.35580000 200 3.952 0.38190000 225 5.139 0.612
Compressor Efficiency
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Testing conclusion• Straight Blade Impeller more effective than Curved Blade Impeller
• In order to run at full speed, an integrated Motor/compressor design is needed
• Compressor was on way to design conditions
– Pressure ratio of 1.7 at Operating speed of 150,000 rpm
– Mass flow rate of 4-8 grams per second
• Reduce losses
– Improve alignment
• Implement laser aligning procedures
• Introduce rigid coupler
• Incorporate one shaft throughout the assembly
– Incorporate air foil bearing / air journal bearing
• Only if power consumption remains high
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CONCLUSIONS
• Miniature RTBC which can provide middle cooling power (1-20 Watt at 77K) may have high efficiency and small footprint.Its unique features including reliability, vibration free and low maintenance may have promising applications
• Its key components, including 0.99 effectiveness micro heat recuperator and meso-scale centrifugal compressor and related bearing technologies are key enabling technologies which can make it have good COP comparing to other competing cryogenic systems.
• The design of micro-scale heat recuperator and compressor is on the way to successful which provide solid evidence to the success of miniature RTBC technology