TASK III A – REVERSE TURBO BRAYTON CYCLE CRYOCOOLER DEVELOPMENT
FOR LIQUID HYDROGEN SYSTEMS
University
Miniature Engineering Systems Group (University of Central Florida)
(http://www.mmae.ucf.edu/~kmkv/mini)
PI: Dr. Louis Chow -Project Director (MMAE)
Co-PIs: Dr. Jay Kapat -Project Co-director (MMAE),
Drs. Q. Chen (MMAE), J. Cho (MMAE), A. Wang (MMAE); L. An (AMPAC); C. Ham
(FSI); K. B. Sundaram (ECE), T. Wu (ECE).
Partners: Dr. Neelkanth Dhere (FSEC); Dr. Nagaraj Arakere (UF); Rini Technologies,
Inc.; Electrodynamics Associates, Inc.; Constellation Technology; NASA KSC: Mr. Bill
Notardonato and Mr. George Haddad; NASA GRC: Mr. Tom Tomsik.
Report of Work Accomplished: (October 01, 2003 – September 30, 2004)
Description
Spaceport operations of the near future are one of the prominent applications for
usage of large quantities of cryogenic propellants. Efficient storage and transfer of these
fluids is necessary for reducing the launch costs. Also, for future manned and
unmanned deep space missions and other missions to Mars, NASA is planning for
extended cryogenic propellant storage durations of the order of several months as
opposed to a few days or weeks. The goal of this task is one such development
comprising of configuring a reliable, compact, highly efficient in its class reverse turbo
Brayton cycle cryocooler, which is capable of extracting 50 W of heat at liquid hydrogen
temperature (~18 K).
Objective
To develop reliable, lightweight, efficient and affordable cryocoolers for
distributed cooling of liquid hydrogen systems for spaceport applications.
Benefit to NASA
All the previous attempts of cryocoolers by NASA for zero boil-off (ZBO) of
cryogenic propellants in space have cooling powers of less than 2W at liquid hydrogen
temperature. These versions of flight-like cryocoolers would be more appropriate for
cooling of sensors and super conducting materials in a spacecraft. The present
cryocooler under development with 50W of cooling power at liquid hydrogen
temperature would be ideal for ZBO of cryogenic propellants in NASA’s future robotic
missions to Mars and for other human space missions.
Task 3: Miniature Centrifugal Compressor Design Verification by Numerical
Simulation and Testing (with appropriate scaling)
(i)
Liquid Nitrogen/Hydrogen Cryogenic System Design
The cryogenic system is a liquid hydrogen cooler, which is designed to liquefy
hydrogen gas at atmospheric pressure and maintain the liquid hydrogen below 19 K.
According to the application requirements, the liquid hydrogen cryocooler should be
able to start up at room temperature and cool down the gaseous hydrogen continuously
to its boiling point (~19K) and keep it at this temperature. It includes two processes:
cool-down process and stable refrigerating process. A two-step RTBC (reverse turboBrayton cycle) cryocooler is proposed for this purpose and is shown in figure 1.
Figure 1. System Scheme of the Two-Step RTBC Cryocooler
The major advantages of the two-step RTBC cryocooler are: a) top step can be
separately started up to cool down the gaseous hydrogen and its container to liquid
nitrogen temperature, which is a relatively efficient cooling process that can save power
and have rapid cooling ability; b) top cycle can use neon as working fluid, which is better
than helium in the cryogenic system. Neon is relatively easy to compress by a dynamic
compressor and its larger molecules provide for easy sealing; c) separated two steps
also give versatility to system development and operation. The top and bottom steps
can be separately developed and freely combined. One top step machine can also drive
multiple bottom steps; d) bottom helium cycle works in cryogenic temperature, which
can reduce its compressor’s rotational speed for the same pressure rise.
The two-step scheme also has some major technical difficulties, which finally
forced us to switch to a single-step system. These drawbacks include: a) inter-cycle
heat exchange will be very troublesome because heat exchange between gases need
large surface area and possibly large pressure loss. The heat that must be exchanged
could be up to 1kW, which may include the heat generated by the motor; b) two different
working fluids cause inter-leakage problem; c) coupled cycle steps need complicated
control system in operation; d) the top step need to cool down the motor of bottom step
which is an extra cost of the two-step structure and will bring down its performance.
Figure 2 shows the thermodynamic scheme of the single-step RTBC cryocooler.
Figure 2. System Scheme of the One-Step RTBC Cryocooler
The conclusion is, although the two-step scheme is better for the cool-down
process, it shows lower overall coefficient of performance (COP), larger size and
inconvenient operation. After carefully evaluating the cost and development difficulties,
the one-step scheme proves to be advantageous. Table 1 shows the comparison of the
two types of schemes.
Table 1. Hydrogen Cryocooler Scheme Comparison
Cooling power (W)
Overall COP
Working fluid pressure ratio
(ii)
Two-step
20
0.0045
Neon 3.3
Helium 2.7
One step
20
0.005
Helium 2.8
Design, Simulation and Testing of the Miniature Helium Two-Stage Intercooled
Centrifugal Compressor
The major difficulty of developing a one–step RTBC hydrogen cryocooler is the
compressor. Helium is difficult to compress, especially at room temperature because of
its large specific heat (> 5000 J/g*K). The requirement of a small mass flow rate adds to
the difficulty. In order to avoid complicated multistage-compressor design, a high
rotational speed is a must. Even then, we still have to design a two-stage compressor
with intercooler to increase its performance. Using our current effort on single stage
neon centrifugal compressor, a helium compressor can come from the similarity
principle. The similarity principle can be expressed as:


2


 RT00
Pˆ
ND
ND
 m

prtt ,tt ,
 f
,
,
,


3
5
2
00 N D

p
D


RT
01
00







PerformancVariables
SimilarityVariables


This equation predicts the performance of a new compressor if it comes from
scaling of an existing compressor. It also predicts the performance of a compressor if
we change its inlet conditions or working fluid. Our current neon compressor shows
good performance in simulation and tests. It provides a prototype we can begin with to
start the helium compressor design. Table 2 compares the stage performance when we
switch gas from neon to air and helium.
Table 2. Stage Performance for Different Gases
Rotating speed (RPM)
Mass flow rate (g/s)
Inlet temperature(K)
Inlet pressure (atm)
Impeller diameter(cm)
Gas specific heat ratio
Gas constant (J/kg*K)
Compression power (W)
Pressure ratio
Neon
141k
8.9
300
1
4.8
1.67
416
950
1.7
Helium
313k
4.6
300
2
4.8
1.67
2079
2370
1.7
Air
108k
10.6
300
1
4.8
1.4
286
823
1.55
The good side of switching from neon to helium is that they have same specific
heat ratio, which means that it is easy to maintain the same pressure ratio. The helium
compressor is an intercooled two-stage compressor, which is designed to have a
pressure ratio of 2.8.
The hydrogen RTBC cryocooler with this helium compressor will have the
following performance characteristics as shown in Table 3.
Table 3. Helium compressor specifications
Stages
Intercooler power (W)
Mass flow rate (g/s)
Compression power (W)
Stage efficiency
Rotation speed (RPM)
Inlet condition
2
630
4.6
3780
0.67
313,000
300K, 2 atm
Due to the limitation of human resources, currently we will focus on the
development of the helium compressor. The solid model of the helium compressor
prototype is shown in Figure 3.
Figure 3. Two-Stage Compressor Solid Model
The first stage is exactly like the neon compressor we have developed, but the second
stage will have a slightly scaled down impeller and a traditional radial diffuser. The
major difficulty in finalizing this design is the bearing structure and rotordynamic issues.
The thermal-fluid design for first stage is partially verified through simulation and
experimental testing.
(iii)
Design, Simulation and Testing of the Single-Stage Compressor
We continued to test the single stage neon/air compressor numerically and
experimentally. This work is not only because we want to have a workable neon/air
compressor but also the knowledge we gain in developing such a compressor is
essential in the helium compressor design. As we have stated, the aerodynamic design
of the helium compressor is mostly identical to the neon/air compressor we are currently
developing.
The single stage compressor is mounted to a driven motor through a flexible
coupler as shown in figure 4. This two-shaft structure gives the test rig rotordynamic
problems, which we are still working on. Figure 5 gives the pressure rise performance of
the compressor before rotordynamic instability happens.
Figure 4. Compressor/Motor Test Rig
Figure 5. Compressor Pressure Rise Performance
In figure 5, we can see that the designed pressure rise is not achieved at the offdesign-points. This under-performance can also be expected at the design point. With
simulation analysis, we figured out that a flow separation at the first blade rows (figure
6). This separation severely reduces the diffuser performance and wastes most of the
kinetic energy of the flow from impeller. Table 4 shows the thermodynamic status at a
series of stations and component efficiencies. We can see that the old diffuser’s
efficiency is as low as 0.086, and that there is almost no pressure rise in the diffuser
(1.278 atm to 1.3 atm). Depending on the simulation results, we redesigned the diffuser
and ran the simulation again. Figure 6 shows the comparison between old and new
diffusers. The flow separation is eliminated. Table 5 shows that the diffuser efficiency is
improved to 0.748 and there is much more pressure recovered in diffuser (1.302 atm to
1.5 atm).
The experimental test of the new diffuser is scheduled. The rotordynamic
problems, which prevented us to go beyond 92,000 rpm, need be solved with analysis.
Details about the current state of testing of the motor for driving the compressor are
mentioned in the next task section.
Table 4. Simulation Data
Static Pressure (atm)
Total Pressure (atm)
Static Temperature (K)
Total Temperature (K)
Mass Flow rate (g/s)
N (RPM)
Efficiency
IGV-Efficiency
Impeller Efficiency
Diffuser Efficiency
P00S
P01S
P02S
P03S
P00O
P01O
P02O
P03O
T00S
T01S
T02S
T03S
T00O
T01O
T02O
T03O
Old Diffuser
0.955
0.862
1.278
1.3
1
0.978
1.564
1.302
297.55
287.38
329.9
249.85
298
297.88
349.4
350.01
15.97
108000
0.4487239
0.7634409
0.8300031
0.0859476
New Diffuser
0.994
0.879
1.307
1.5
1
0.975
1.595
1.501
297.45
289.17
330.84
351.08
298
297.86
350.43
350.5
15.7
108000
0.698388
0.782609
0.855529
0.748759
(00- inlet of IGV; 01-inlet of impeller; 02-outlet of impeller; 03-outlet of diffuser)
Figure 6. Old and New Diffuser Flow Fields using CFD Simulation
Task 4: Motor/Generator: Design, Fabrication and Testing
This part of the report summarizes the design, simulation, fabrication, control and
testing of the super-high speed permanent magnet synchronous motor (PMSM) with the
rating of 2000 W output shaft power at 200,000 rpm.
(i)
Motor Structure and Dimensions
The permanent magnet is centrally located inside the hollow shaft. The rotor was
assembled by heating the shaft and cooling the permanent magnet. For super-high
speed, 2-pole configuration is preferred since the required electrical frequency is the
lowest for the same speed. Low electrical frequency can reduce switching loss of the
controller. The number of turns per phase and the pitch factor were considered based
on the requirements such as dc supply voltage, shaft speed and desired back-EMF.
Figure 7. Cross- section of the Designed PMSM
Figure 8. Dimensions of the Hollow Shaft
The designed PMSM (figures 7, 8) consists of a hollow shaft with a magnet
inside it. The direction of flux flow is radial. The cross-section of the shaft with magnet
inside is as shown in fig. The non-uniform shape of the magnet prevented it from
slipping inside the shaft while rotating. Two plugs of same material as that of the shaft
were designed to support the magnet at the center from two sides. The bearings were
mounted on the plugs.
(ii)
Material Selection
The materials were carefully selected for cryogenic and super-high speed
application. Neodymium-iron-boron (Nd-Fe-B), which has the highest energy product
compared to other types of permanent magnets, is widely used in electrical machines
and other applications. However, when the temperature is below 140 K, Nd-Fe-B will
change from a uniaxial material to an easy-cone anisotropy material, which makes the
magnet to easily get demagnetized. Therefore, Nd-Fe-B is generally not considered for
applications below 140 K, although operations at 60~64 K has been reported.
Samarium Cobalt (SmCo) also has very high energy product. It is very stable at low
temperatures and has very low temperature coefficient of coercivity and remanence.
The Curie temperature and operating temperature of SmCo are also very high.
So while inserting the permanent magnet into the shaft, we can heat the shaft to very
high temperature without affecting the performance of SmCo. Therefore, for 77 K
application, SmCo was chosen as the permanent magnet in our design.
Copper was used for the winding. The copper dc loss is greatly reduced at low
temperature due to increased conductivity. However, the eddy current loss of the
winding due to proximity effect will increase because it is inversely proportional to the
conductivity. Thus, multi-strand Litz-wire has been used to reduce the eddy current loss.
The Litz-wire is constructed using 75 strands of AWG 36 (0.125 mm diameter), coated
with heavy 200°C polyesterimide and overcoated with polyamide-imide to meet IEC MW
35. After that, it is further wrapped with sofimide to withstand temperatures down to 77
K. The constructed Litz-wire was tested at 77 K, and the results show that it can survive
at 77 K without any insulation problems.
The motor stator was made of laminated silicon steel. The magnetic properties
of the silicon steel are relatively insensitive to the temperature changes. However, since
the conductivity of silicon steel will increase with the decrease of temperature, the eddy
current loss will increase when temperature decreases. To reduce eddy current loss,
thin laminations are required. 0.127 mm (0.005 in) non-oriented silicon steel that is
available in the market is used. The conventional epoxy adhesive (Magna-Tac E645) is
used to bond the laminations.
The selection of the shaft material was based on certain major factors: 1) The
shaft material should be non-magnetic and must have a high yield strength to withstand
the centrifugal stress developed in the shaft-magnet system while operating at 77 K and
200,000 rpm. 2) The assembly of the shaft and magnet should be done such that the
magnet is at interference fit while operating at 77K and 200,000 rpm. An unequal
expansion-contraction in shaft and magnet at 77 K as well as an expansion of the shaft
due to centrifugal force while rotating at 200,000 rpm may cause the magnet to be loose
fit and cause mechanical failure. 3) The shaft material should be machineable.
The materials taken into consideration were Titanium, Inconel and MP35N. All
these materials have high yield strength at 77 K. But density of Titanium is less. The coefficient of thermal expansion of Titanium is almost the same as that of magnet material
SmCo. Thermal conductivity of Titanium is more than other metals considered. The
material used for shaft was commercially available Titanium 6 Al -4 V. This material
possesses a yield strength of 1400 MPa at 77 K. It has similar co-efficient of thermal
expansion as that of magnet material (SmCo). This alloy has less density compared to
other materials that were taken into consideration and is fairly machineable.
(iii)
Simulation and Analysis of Motor Performances
The air gap flux densities in normal and tangential directions are shown in the
figure 9. The corresponding harmonics analysis of the flux density in normal direction
shows that the dominated harmonics is the 3rd harmonics, which is still very low
compared to that of slot structure.
The no-load back-EMF was simulated by using 2-D transient time-stepping solver with
motion (figure 10). The FFT analysis shows that the normalized 3 rd harmonics of back
EMF is less than half that of the normalized 3 rd harmonics of air gap flux density in
normal direction. This improvement is due to the short-pitch winding.
0.5
20
Normal
Tangent
15
10
0.3
Back EMF (V)
Flux density (T)
0.4
0.2
0.1
5
0
-5
-10
0.0
-15
-0.1
-20
0
20
40
60
80 100 120 140 160 180 200
Angle (Deg.)
-20
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
Time (ms)
Figure 9. Air Gap Flux Density in
Normal and Tangential Directions
Figure 10. No-load Back EMF
The torque was simulated by applying field-oriented current. When there is no
harmonics in the field oriented current, the simulated results show that the torque ripple
can almost be ignored.
Eddy current losses of the winding were simulated for both solid wire and multistrand Litz-wire cases. If solid wire is used instead of multi-strand Litz-wire, the eddy
current density can be larger than 2,000 A/mm 2 when the rotor is rotating at 200,000
rpm and the operating temperature is 77 K. Such a large eddy current is mainly due to
the super-high speed and extremely low copper resistivity at 77 K. So solid wire is not
practical for the cryogenic motor with slotless structure and rotating at super-high
speed. The eddy current is greatly reduced after using multi-strand Litz-wire. The
simulation and analysis results (table 5) show that the motor efficiency is about 89.9%
when running at room temperature, and 91.8% when running at 77 K.
Table 5. Simulated and Estimated Losses
Copper DC loss (W)
Copper AC loss (W)
Stator iron loss (W)
Bearing loss (W)
Windage loss (W)
Low pass filters loss (W)
Controller loss (W)
Total loss (W)
Efficiency
@ RT
71.3
1.4
4.2
@77K
7.5
12.8
9.85
30
6.3
11
100
224.2
177.5
89.9%
91.8%
(iv)
FEM Simulation of the Rotor
The rotor structure is subjected to various stresses under operation. The shaft
was tested for pure shear with Maximum Shear Stress theory. The maximum bending
stress is given by σmax = M/Z, where Z = section modulus. It was found that bending
stress developed was much less than ultimate tensile strength of shaft and magnet.
The rotating shaft is also subjected to stress due to centrifugal force while
operating at 200,000 rpm. Centrifugal stress analysis was done using FEM software at
the operating speed. The shaft was observed to expand more than the magnet and
would result in a clearance fit. A high risk of crumbling of the magnet while shuddering
inside the shaft prevails. Thus, interference fit was required. It was calculated that
heating the shaft to 400 0C and cooling the magnet to liquid nitrogen temperature would
produce the necessary clearance to insert the magnet in the shaft. On cooling, the
interference fit would be established. Residual stresses arising due to this process of
assembling were calculated. The magnet was tested for failure by compressive stress
due to interference fit between the shaft and magnet. The centrifugal stress and residual
compressive stress would contribute to the total stress developed while the shaft is in
operation. The maximum total stress observed was 852 MPa (figure 11). This value was
much less than the yield strength of Titanium grade that was used to build the shaft.
Figure 11. Stress in the Shaft and Magnet Assembly
Bearing
Plug
Shaft
Magnet
Plug
Bearing
Figure 12. Conceptual Design of the Rotor
The conceptual design for the rotor system can be seen in figure 12. The magnet
had to be assembled inside the shaft. The two plugs were used to house the bearings.
One side of the plug had the same section as that of the shaft. Flexible couplers could
have been used on a separate component to house the bearings, but the critical speed
would go down. The plugs were press fit on to the shaft and act like a rigid coupler.
Rotordynamic analysis was done prior to fabrication. Two models were
considered here.
Figure 13. Rotordynamic Analysis of Motor Shaft for Model 1
The first model is based on our initial conceptual design. Figure 13 is a
schematic of the FE rotordynamics model with 9 elements and 40 degrees of freedom.
The upper half of the model shows the mass distribution while the lower half shows the
stiffness distribution. The mass distribution shows the magnet counted for the mass, but
not for the stiffness. The above model has shown that it has a certain bearing span. The
bending critical speeds are quite sensitive to both bearing stiffness and bearing span.
Ball bearing stiffness can vary widely depending on preload, operating internal
clearance, etc. In this case the bearing stiffness can vary anywhere from 10000 lb/in to
200000 lb/in, resulting in fairly large variations in the 1st critical speed (~155,000 rpm to
~250,000 rpm). Ball bearings have very little damping and hence operation near critical
speeds results in very large transmitted forces. Generally, the 1st bending critical speed
is kept at least 30% higher than the highest operating speed.
Figure 14. Critical Speed Map and Strain Energy Distribution for the First Model
The 1st mode (bending mode) strain energy in the shaft represents nearly 94%
of the total energy, with 6% of the strain energy due to motion at the bearings (figure
14). However, even this motion at the bearings will transmit large forces. The curves in
the critical speed map represent the three bending modes. As the bearing stiffness
increases the critical speeds go higher.
Figure 15. Rotordynamic Analysis of Motor Shaft for Model 2
The second model shown in figure 15 represents a modification of first model in
which an attempt has been made to increase the first bending critical speed. The above
model has a slightly smaller bearing span (by 14 mm) from model 1. Note, however,
that this results in considerable changes in the critical speed locations. The first mode
increases considerably, to around 250k rpm.
Figure 16. Critical Speed Map and Strain Energy Distribution for the Second Model
For the 1st mode (bending mode) strain energy in the shaft represents about
86% of the total energy, with 14% of the strain energy due motion at the bearings (figure
16). Note that the first three critical speeds can vary a fair amount in the bearing
stiffness range. We are only concerned about the 1st critical speed since it is the one
near the maximum operating speed of 200,000 rpm. Bearing stiffness has to be around
100,000 lb/in to get the 1st mode above 200k rpm. The fabrication of the rotor of the
Permanent Magnet Synchronous Motor was based on the second model shown above.
(v)
Fabrication and Assembly
The components of the PMSM were fabricated and are shown in figures 17, 18
and 19.
Figure 17. Hollow Shaft and Magnet
Figure 18. Stator and Winding
Figure 19. Casing, Caps and the assembled Shaft
The rotor of the motor was fabricated out of Titanium 6Al-4V. The shaft was
machined by EDM process. It included two caps that house the bearings and also
support the magnet inside the shaft. These parts were machined partly by EDM and
regular machining process.
Titanium
Shaft
SmCo
Magnet
10 mm
70 X 70 mm
Ceramic
Base
Figure 20. Fixture to assemble the Shaft and Magnet
The assembly process involved heating the Titanium shaft with simultaneous
cooling of the magnet to achieve a specified clearance between them and shrink fit the
magnet inside the shaft. The two caps that hold the magnet and house the bearings
have interference fit with the main shaft. The bearings have the fit with the housing as
specified by the bearings manufacturer. The housing is air-cooled. The fixture in figure
20 was used to assemble the shaft and magnet. The heated shaft was placed on it as
shown. The magnet was dropped inside the shaft and was located by the projected part
from the fixture base.
(vi)
Controller Design
The controller is based on TI Digital Signal Processor TMSLF2407A. It includes
hardware and software parts. Hardware is mainly to provide suitable drive power to the
motor. Software is to produce the control signals based on space vector PWM to obtain
good performance. Several optimal approaches are considered to improve the
efficiency and performance of new controller.
1.
Hardware improvement: Various new chosen components are employed to
improve controller dynamic characteristics and provide higher efficiency. Firstly, power
dissipation in MOSFETs is a main concern of controller system efficiency. The total
power loss includes resistive power loss and power loss in switching. Resistive power
loss will increase a lot if MOSFET drain to rain on resistance (Rds (on)) is high when
load current is huge. New MOSFET, FDP047AN08A0, has an Rds (on) of 4.7m, which
is less than one-tenth of the prior one.
The new controller system performance will also be highly improved because of
much better dynamic and switching characteristics of new components. At first, increase
in the gate drive current to 2A, which is a ten fold increase compared to the prior one
will optimize the system dynamic and switching performance. Higher gate drive current
also can reduce conduction and switching losses quite a lot. Secondly, the total gate
charge of the new MOSFET is only 92nC, which is small enough to make sure that
transistors are switched on/off correctly. Thirdly, high-speed optocoupler HCPL 2631
(10M bit/s) replaced HCPL 2531 (1M bit/s), which is fast enough for our high switching
frequency need and will lower SVPWM signal distortion.
The drive chip is an important component in the controller system, which critically
affects the dynamic characteristics. Well-known IR 2110 is used for new 2000W
controller. The hardware circuit is shown in figure 21.
Figure 21. 2000 W PMSM Motor Controller Hardware Circuit Schematic
(Only one phase shown in this figure because of symmetry of the other two phases)
2.
Software improvement: Constant volts/Hertz method is employed to produce
controlling voltage responding to frequency because of its easiness to realize and
advantage at middle to high speed applications. The purpose of constant V/f control is
to generate required phase voltage Vs* which is maintained proportional to the
frequency so that the stator flux remains constant. This permits for nearly maximum
available torque per ampere of the stator current and fast transient response for the V/f
control. As a result, higher efficiency and better performance can be permitted. For
normal motor system, constant V/f control is a linear control because stator resistance is
small compared with motor inductance and can be neglected. However, because of the
compact size requirement of the 2000W PMSM motor, the stator resistance effect
cannot be neglected anymore. An optimal volts/Hertz control is proposed based on
system simulation and analysis. System simulation results provide optimal V/f control as
shown in figures 22 and 23.
Figure 22. Optimal V/f Control Simulation
Figure 23. Proposed Optimal V/f Curve for the Super High-Speed Motor
Test results show that the proposed optimal V/f control method improved system
performance and efficiency. In space vector PWM method, voltage modulation index, m,
is chosen less than 0.952 to avoid over-modulation. And also, a carefully chosen
switching frequency is helpful to realize higher speed and better performance.
(vii)
Motor testing
The free spin test was first performed (figure 24). The motor ran very smoothly
below 130,000 rpm. We expect the free spin loss at 200,000 rpm to be about 184 W if it
is well adjusted. And the projected efficiency at 200,000 rpm with 2000W output is
around 90%.
Figure 24. Motor Testing Curve
Currently, when the speed is above 140,000 rpm, the loss will increase greatly.
One reason of large loss is that the pre-load is not applied in this prototype. Preload can
help to avoid skidding of the bearing balls to reduce noise and vibration. There are two
basic methods of preloading: solid preloading and spring preloading. Solid preload can
be obtained by mechanically locking all of the rings in position while under a reasonable
axial load. The advantages of solid preload are simple structure and high stiffness. The
disadvantage is high variation in preload under temperature variation, and that the
preload can reduce with wear. Spring preload can be applied using coil springs. An
advantage of spring preload is that it maintains consistent preload with temperature
variation. The disadvantages are that the designs are more complex and normally have
lower stiffness.
If the applied preload is insufficient, fretting corrosion can occur. This happens as
a result of vibration causing the balls to resonate and abrade on the raceways.
Therefore, obtaining the correct preload is very important. The cap was modified to
consider spring preload (figure 25). Four springs were used inside the cap to provide
adjustable axial preload. The motor will be tested again with the modified cap.
Figure 25. Modified Cap considering Spring Preload
Overall Project Schedule/Tasks List:
Task 1. Design and Fabrication of Miniature Centrifugal Compressor
Task 2. Design of a High-speed, High-efficiency PMSM
Task 3. Miniature Centrifugal Compressor Design Verification by Numerical Simulation
and Testing (with appropriate scaling)
Task 4. Fabrication and Testing of PMSM
Task 5. Two-stage Centrifugal Compressor – Design and Fabrication
Task 6. 5.4 kW PMSM – Design and Fabrication
Task 7. Integration and Preliminary Testing of the Motor/Compressor Assembly
Task 8. Overall System Optimization – Systematic Testing of the Motor/Compressor
Assembly, Evaluation, Possible Design Changes
Start Date: July 01, 2002; Estimated End Date: September 30, 2006;
Estimated Duration: 51 months
Months -
1
2
3
4
5
6
7
8
9
10
11 12
13
14
15
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Task 1
Task 2
Task 3
Task 4
Task 5
Task 6
Task 7
Task 8
YEAR I
YEAR II
YEAR III
YEAR IV
TRL 2
BETWEEN TRL 2 AND TRL 3
BETWEEN TRL 3 AND TRL 4
TRL 4
Overall Project Milestones: Year II
M1 – 09/30/04: To verify the design of miniature centrifugal compressor by numerical
simulation and testing (with appropriate scaling) (completed)
M2 – 09/30/04: To fabricate and test the PMSM (completed)
Annual Report for Year II (completed)
Any Significant Changes to Milestone Chart: None
Problem Areas Affecting Cost Elements: None