Gerald Montague

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Mark Jansen

Sest, Inc., Brook Park, Ohio

Andrew Provenza

Glenn Research Center, Cleveland, Ohio

Ralph Jansen and Ben Ebihara

University of Toledo, Toledo, Ohio

Alan Palazzolo

Texas A&M University, College Station, Texas

Room Temperature Characterization of a

Magnetic Bearing for Turbomachinery

NASA/TM—2002-211904

October 2002

ARL–TR–2858

U.S. ARMY

RESEARCH LABORATORY
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Gerald Montague

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Mark Jansen

Sest, Inc., Brook Park, Ohio

Andrew Provenza

Glenn Research Center, Cleveland, Ohio

Ralph Jansen and Ben Ebihara

University of Toledo, Toledo, Ohio

Alan Palazzolo

Texas A&M University, College Station, Texas

Room Temperature Characterization of a

Magnetic Bearing for Turbomachinery

NASA/TM—2002-211904

October 2002

National Aeronautics and

Space Administration

Glenn Research Center

ARL–TR–2858

U.S. ARMY

RESEARCH LABORATORY
Acknowledgments

This research was funded by NASA Glenn Research Center’s Smart Efficient Components program (SEC), which is

managed by Bob Corrigan. The authors would like to thank Peter Kascak (University of Toledo) for his assistance

in the power measurement, John Poles (NASA Glenn) for his significant electrical contributions, and Tim Czaruk

(QSS Group, Inc.) for test support. The authors would also like to recognize Gerald Brown and Albert Kascak

(NASA Glenn) for their expert guidance and review. Finally, we praise Texas A&M students Randy Tucker, Jason

Preuss, and Andrew Hunt who designed and constructed the high temperature coils used in this bearing.

Available from

NASA Center for Aerospace Information

7121 Standard Drive

Hanover, MD 21076

National Technical Information Service

5285 Port Royal Road

Springfield, VA 22100

The Aerospace Propulsion and Power Program at

NASA Glenn Research Center sponsored this work.

Available electronically at http://gltrs.grc.nasa.gov
NASA/TM—2002-211904 1

Room Temperature Characterization of a 

Magnetic Bearing for Turbomachinery 

Gerald Montague 

U.S. Army Research Laboratory 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Mark Jansen 

Sest, Inc. 

Brook Park, Ohio 44142 

Andrew Provenza 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Ralph Jansen and Ben Ebihara 

University of Toledo 

Toledo, Ohio 43606 

Alan Palazzolo 

Texas A&M University 

College Station, Texas 77843 

ABSTRACT 

Open loop, experimental force and power 

measurements of a three-axis, radial, 

heteropolar magnetic bearing at room 

temperature for rotor speeds up to 20,000 RPM 

are presented in this paper. The bearing, NASA 

Glenn Research Center’s and Texas A&M’s 3rd

generation high temperature magnetic bearing, 

was designed to operate in a 1000 °F (540 °C) 

environment and was primarily optimized for 

maximum load capacity. The experimentally 

measured force produced by one C-core of this 

bearing was 630 lb. (2.8 kN) at 16 A, while a 

load of 650 lbs (2.89 kN) was predicted at 16 A 

using 1D circuit analysis. The maximum 

predicted radial load for one of the three axes is 

1,440 lbs (6.41 kN) at room temperature. The 

maximum measured load of an axis was 

1050 lbs. (4.73 kN). Results of test under 

rotating conditions showed that rotor speed has 

a negligible effect on the bearing’s load 

capacity. A single C-core required 

approximately 70 W of power to generate 300 lb 

(1.34 kN) of magnetic force. The room 

temperature data presented was measured 

after three thermal cycles up to 1000 °F 

(540 °C), totaling six hours at elevated 

temperatures. 

INTRODUCTION 

The gas turbine industry has a continued 

interest in improving engine performance and 

reducing net operating and maintenance costs. 

These goals are being realized because of 

advancements in aeroelasticity, materials, and 

computational tools such as CFD and engine 

simulations. These advancements aid in 

increasing engine thrust-to-weight ratios, 

specific fuel consumption, pressure ratios, and 

overall reliability through engine operation at 

higher rotational speeds and higher 

temperatures with greater efficiency. 

Rolling element bearings and squeeze film 

dampers are currently used to support gas 

turbine engine rotors. The bearings presently in 

operation are limited in temperature (<500 °F, 

270 °C) and speed (DN <2 million), and require 
NASA/TM—2002-211904 2

both cooling air and a lubrication system. 

Rolling element bearings in gas turbines are 

being pushed to their limits and new 

technologies must be found in order to take full 

advantage of other engine advancements. 

Magnetic bearings are well suited to operate at 

elevated temperature and higher rotational 

speeds and are a promising solution to these 

problems. Magnetic bearing technology is 

being developed worldwide and is considered 

an enabling technology for new engine designs. 

Turbine and compressor spools can be 

designed with higher operating temperatures 

and significantly larger, stiffer, highly damped 

rotors, which spin at higher rotational speeds. 

Magnetic bearings will allow compressor, 

turbine, combustor, and nozzle advancements 

to reach their full performance potential. 

Beams and Black performed some of the first 

research on magnetic suspension of rotating 

bodies [1] in 1937. An ultracentrifuge rotor, 

which was supported axially by compressed air, 

was replaced with a magnetic bearing. The 

vertical rotor position was so stable that no 

motion was detected with a 100X microscope 

and the reduction in power losses enabled 

operation at very high speeds. 

Some of the first work on magnetic properties of 

soft materials at extreme temperatures was 

done at Westinghouse Electric in 1967 [2]. 

Information was gathered about eight selected 

materials for high temperature electric devices. 

More recently, Fingers from the U.S. Air Force 

Research Laboratory supported the more 

electric aircraft effort [3]. The major objective of 

this work was to replace mechanical systems 

with electrical devices in order to increase 

military aircraft reliability and maintainability 

while reducing ground support equipment.

The first high temperature magnetic bearing 

fabricated and tested at NASA Glenn was 

homopolar design with an electromagnetic bias 

coil. The second generation was a heteropolar 

design that minimized pole face area to enable 

maximum volume for high temperature coil 

design. 

Data presented in this paper characterizes the 

3rd generation high temperature, high load 

magnetic bearing developed at the NASA Glenn 

Research Center at room temperature. Bearing 

force as a function of current and speeds to 

20,000 RPM is presented. In addition, bearing 

power consumption measurements are made at 

several current levels. 

TEST FACILITY 

FACILITY DESCRIPTION 

The high temperature magnetic bearing test 

facility is shown in Figure 1. The structural 

support can accommodate thrust and radial 

bearings up to 9.0 in (22.84 cm) diameter with a 

maximum axial loading of 5,000 lb (22.25 kN) 

and a maximum radial loading of 2,500 lb 

(11.13 kN). The test facility has already been 

used in several different configurations. 

Figure 1 – High temperature magnetic bearing test facility 

at NASA Glenn Research Center. 

The facility was previously used to measure 

linearized force coefficients (position and 

current stiffness and damping) of magnetic 

bearings versus speed, bias current, excitation 

frequency, and eccentricity. The second 

configuration was used to demonstrate stable 

rotor levitation and fault tolerance of a radial 

magnetic bearing at high temperature (1000 o

F) 

up to 15,000 RPM [4, 5]. 

The present configuration can be seen in 

Figure 2. The magnetic bearing is located at 

the center of gravity of the 2.97 in (75 mm) 

diameter rotor. A 0.022 in (0.559 mm) radial air 

gap exists between the stator poles and the 

rotor. The rotor has interchangeable sleeves 

on each end that interface it with the support 
NASA/TM—2002-211904 3

Figure 2 – Third generation high temperature magnetic 

bearing facility. 

bearings, which currently are high-speed, 

grease packed, duplex ball bearings. For this 

configuration, the rotor was fitted with zero 

clearance sleeves so it is supported on the ball 

bearings at both ends. This was done so that 

static forces exerted by the magnetic bearing 

could be easily measured outside the hot 

section. The outboard sleeve can be replaced 

with a positive clearance sleeve so that the 

magnetic bearing can support the rotor. An air 

turbine drives the rotor. 

The magnetic bearing stator is an isolated Ccore, 12-pole, heteropolar design and is 

described in detail in Reference 5. The rotor 

diameter is 2.98 in. (75 mm) and the stator 

width is 3.4 in. (76.1 mm). Each C-core is 

wrapped with two coils and each coil has 52 

turns of specially insulated wire. 

Power to the magnetic bearing is provided 

through tri-state pulse width modulated (PWM) 

amplifiers that are passively filtered to remove 

high frequency amplifier noise that results from 

amplifier switching. 

Heat is supplied to the bearing through three, 

3 kW band heaters wrapped around the stator. 

Ground Fault Circuit Interrupts (GFCI) have 

been incorporated into the heating system for 

personnel safety. 

SENSORS 

The test rig is equipped with several types of 

sensors. High load, high bandwidth, 

piezoelectric load cells support the inboard and 

Figure 3 – Six load cells supporting the outboard duplex 

ball bearing. 

outboard rolling element bearings (Figure 3). 

The load cells are resting on machined flats of 

the bearing housing. This load cell mounting 

design eliminates axial constraints of the ball 

bearing. 

These load cells are capable of measuring 

loads up to 1,688 lb (7.5 kN) at 200 kHz and 

have a maximum operating temperature of 

385 o

F (196 oC). Two load cells are aligned 

along each of the three magnetic bearing axes 

on both support bearings for a total of 12 load 

cells. Each load cell was set with a preload of 

~250 lb (1.1 kN). 

High temperature eddy current displacement 

probes are used just outside the stator on both 

sides of the magnetic bearing to monitor rotor 

position. Each side has four probes (X+, X-, Y+, 

Y-) for a total of eight probes. The probes are 

capable of 30 μin (1.18 μm) accuracy and are 

temperature compensated for the 1000°F 

(538 °C) environment. 

Temperatures are recorded at several locations 

within the test rig. Both sides of the stator have 

three thermocouples mounted to them. Each 

set of support ball bearings is also equipped 

with a thermocouple. A handheld infrared 

thermometer can also be used to gather 

additional thermal information about the shaft 

during rotation. 

Other sensors include: an eddy current 

displacement probe used to measure rotor RPM 

and phase, current sensors to measure C-core 
NASA/TM—2002-211904 4

currents, and temperature and redundant RPM 

sensors in the air turbine. 

DATA ACQUISITION 

A LabVIEW-based data acquisition system is 

used in this facility. This system is capable of 

capturing data at rates up to 15,000 samples 

per second per channel. Data is recorded for 

the drive end and outboard load cells, support 

bearing temperatures, stator temperatures, X 

and Y axis displacement probes, once around 

probe, and bearing coil currents. Data is 

displayed in real time and can be saved directly 

into a spreadsheet format. 

EXPERIMENTAL 

For the tests reported in this paper, the shaft 

was mounted on zero clearance ball bearing 

supports so the load exerted by the magnetic 

bearing C-cores on the rotor could be 

measured. 

ROTOR BOUNCE MODE 

For a rigid rotor supported on two springs of 

stiffness K: ωn =√(2K/m) = 24,143 RPM = 2522 

rad/s., with m = 17/386, K = 138,934 lb./in. The 

mode was verified two different ways. First, a 

swept sine input was fed into the PWM 

amplifiers and a dynamic signal analyzer 

recorded the DC current to load cell output 

transfer function. Since the load cells and ball 

bearings provide little damping to the rotor, the 

first rotor bending critical was easily measured 

to be 407 Hz. Second, a rap test on the rotor 

produced a first resonance at 386 Hz. These 

results are in 95 percent agreement. 

This resonance did not limit the maximum 

testing speed to 20,000 RPM. Rather, 

overheating of the unsealed, grease packed ball 

bearings at 20,000 RPM was the limiting factor. 

ROTOR IMBALANCE FORCES 

Rotor imbalance vs. rotational speed was 

determined so that this force could be 

accounted for in the DC load tests being 

performed at speed. Synchronous component 

amplitude and phase for all twelve-load cells 

were recorded with the magnetic bearing power 

off up to 20,000 RPM. The maximum 

transmitted unbalance force measured by a 

load cell was 52 lb (23.5 kg) at 20,000 RPM. 

However, due to the load cell preload and the 

vector summation for load calculation, the rotor 

imbalance force cancels itself during data 

acquisition of the magnetic axis load 

measurements. 

FORCE CHARACTERIZATION 

THEORETICAL FORCE PREDICTIONS 

For a general magnetic actuator, the force 

produced by magnetic flux in an air gap of area 

Ag is given by 

2

g 2G

NI A

2

1 F 



 



 = αμο (1) 

In this equation, μo is the magnetic permeability 

of evacuated space, NI is the ampere-turns of 

the coil(s) driving the flux through the circuit, 

and 2G represents the total effective magnetic 

air gap. α is a force de-rating factor, which 

accounts for leakage, fringing, non-uniformity in 

flux density across the pole faces, and finite 

path permeability [6]. 

The actuator measurements documented in this 

paper are based on the performance of a 

complete c-core circuit. Consequently, the force 

produced by a single c-core is twice that 

obtained using equation 1, since there are two 

air gaps between it and the rotor. 

Using μo = 4π x 10–7

 N/A2, N = 52*2 turns, Ag = 

2.0 in.2

 (12.9 cm2), G = 0.022 in. (0.559 mm), 

α = 0.8, and converting to Imperial units, yields: 

F( )I 2 53 I lbf 2 = . × (2) 

At 10 A, the force produced by C-core #1 is 253 

lbf (1.1 kN). Since the total air gap area is 4 in.2

(25.8 cm2

), the effective magnetic pressure 

produced by one C-core at 10 A is 63.3 psi 

(434 kPa). 

Flux density vs. applied current data taken for 

similar heteropolar magnetic bearings made of 

Hyperco 50 laminations have shown that 2 
NASA/TM—2002-211904 5

Tesla is achievable in the air gap. Using the 

equation 

2G

B sat

sat

μ ΝΙ = ο (3) 

and solving for Isat with Bsat = 2 T yields 16.9 A. 

The maximum force capacity of a C-core then 

from equation (2) becomes 720 lbf (3.2 kN), 

which corresponds to an effective magnetic 

pressure of 180 psi (1,241 kPa). 

The saturation flux density of 2 Teslas used 

here for predicting the saturation current is 

selected somewhat arbitrarily. Typically Bsat is 

selected as the point on the static B-H material 

data curve where linearity is lost and the curve 

heals over towards a final slope of μo. 2 Teslas 

may not be that point for this material. One way 

to determine Bsat is to insert a hall-effect probe in 

the gap of the c-core and measure flux density 

as a function of c-core current. This test will be 

performed for a future paper. 

For this 12 pole, isolated C-core bearing 

oriented as shown in Figure 4, the maximum 

radial force can be produced only in the X 

direction. A maximum positive force will be 

produced when C-cores #2,3,4 are driven with 

16.9 A, and C-cores #1,5,6 are not used at all. 

Notice that all six C-cores can contribute to the 

production of an X force, but only four can 

contribute to a Y force. In the X-direction, the 

maximum radial force that can be produced by 

this bearing is 1,440 lbf (6.4kN). 

Although this 3rd generation high temperature 

magnetic bearing was optimized for load 

capacity, the stator poles do not completely 

cover the rotor. To determine the effective 

maximum radial pressure, the maximum force 

must be divided by the net hemispheric surface 

area of the rotor. The outer diameter of the rotor 

is 2.97 in. (75.4 mm) and the stator poles are 

3.4 in. (86.4 mm) long, making the net surface 

area 16 in.2

 (103 cm2). Thus, the net effective 

maximum pressure produced by this bearing is 

90 psi (620 kPa). 

MAXIMUM FORCE CAPABILITY OF A SINGLE 

CORE

The first test run was to determine the 

maximum load capacity of a single core and the 

current level where saturation is reached. This 

was done by attaching the coil directly to a high 

power DC supply (100V, 40A) and measuring 

load as a function of current. Load remains 

constant as current increases due to flux 

saturation in the lamination material. The #1 Ccore was tested and the results are shown in 

Figure 5. 

Figure 4 – Total load on each magnetic axis is a sum of the 

force from three C-cores. 

0

100

200

300

400

500

600

0 5 10 15 20 25

Current

Load (lb)

Figure 5 – Maximum load and saturation level for C-core #1 

using the DC power supply. 
NASA/TM—2002-211904 6

DC POWER SUPPLY VERSUS PULSE WIDTH 

MODULATION POWER SUPPLY 

Power to the magnetic bearing is normally 

supplied by a pulse width modulation (PWM) 

power supply. This type of power supply is 

more efficient than a direct DC supply. For this 

facility, the practical output limit of the PWM 

supplies is 15 Amps. Since the saturation level 

of the bearing occurs at around 21 Amps, the 

DC supply was used. In order to assure that 

accurate data could be attained with the PWM 

supplies, a comparison was done between the 

PWM loads and the loads attained with two 

different DC supplies. There is almost exact 

correlation between all three. The results are 

shown in Figure 6. 

0

100

200

300

400

500

600

700

0 5 10 15 20 25

Current

Load (lb)

DCSupply 1 DCSupply 2 PWMSupply

Figure 6 – Comparison of loads obtained with PWM 

supplies and two DC power supplies (C-core #1). 

INDIVIDUAL C-CORE LOAD CAPACITY

The next series of tests run were to determine 

the load capacity of a single C-core as a 

function of current. Connecting a constant 

current supply to the power amplifiers and 

measuring the load at steady-state conditions 

achieved this. 

Each C-core was measured for load capacity at 

0 RPM and compared with one another. These 

results are shown in Figure 7 and summarized 

in Table 1. Discrepancy between the C-cores 

will be discussed later in this paper. Once each 

C-core load capacity was determined, the tests 

0

100

200

300

400

500

600

700

0 5 10 15

Current

Load (lb)

C-Core 1 C-Core 2 C-Core 3

C-Core 4 C-Core 5 C-Core 6

Figure 7 – Load as a function of current for individual 

C-cores. 

Table 1 – Summary of single C-core load (lb) capacity up to 

15 amps at zero RPM. 

C-Core 6 Amp Load 10 Amp Load 15 Amp 

Load 

1 71 203 431 

2 56 167 374 

3 49 140 363 

4 72 212 473 

5 101 291 597 

6 90 320 630 

to determine the load capacity along a magnetic 

axis and at rotor speeds to 20,000 RPM were 

performed. 

LOAD CAPACITY OF AN AXIS 

The second series of tests run were to 

determine the load capacity on an axis. The 

maximum load on an axis is a function of three 

adjacent C-cores. The two C-cores on either 

side of the main axis have a contribution to the 

total load in that axis (Figure 4). Static tests 

were performed to check if all the C-cores had 

similar load capacity. 

The total force along each magnetic axis is the 

sum of the forces from three C-cores (Figure 4). 

For example, in the #3 C-core direction, the total 

force is the sum of the #3 C-core, and the 

x-components from the #2 and #4 C-core. The 

y-components of force from the #2 and #4 coil 
NASA/TM—2002-211904 7

 

0

200

400

600

800

1000

1200

0 5 10 15

Current

Load (lb)

C-Cores 6-1-2 C-Cores 2-3-4

Figure 8 – Load as a function of current for two magnetic 

axes at zero RPM.

Table 2 – Resultant load (lb) along two magnetic axis 

capacity up to 15 amps at zero RPM. 

Magnetic 

Axis 

6 Amp 

Load 

10 Amp 

Load 

15 Amp 

Load 

6-1-2 151 460 1056 

2-3-4 128 369 853 

are equal and opposite and therefore cancel 

each other. 

The results of load vs. current at 0 RPM for 

each magnetic axis are shown in Figure 8 and 

summarized in Table 2. 

To confirm the load of the magnetic axis was 

applied in the same direction as the center Ccore of that axis, the angle at which the load 

was applied was also monitored. The angles 

matched the theoretical value (i.e. 120° for the 

#6-1-2 axis) within a few degrees. 

The discrepancy in the load capacity of the Ccores and also of the magnetic axis could be 

related to a non-uniform air gap between the 

rotor and stator. A dial indicator was used to 

measure 0.0005-0.001” rotor lamination run-out. 

Plastic feeler gages were used to verify the nonuniform air gap. A minimum gap of 0.015” was 

measured. The eight displacement sensors 

monitored position during all the tests to ensure 

the rotor did not bow or change position relative 

to the support stand. However, the stator was 

Table 3 – Coil load vs. average load at zero RPM. 

Amp Ave. 

Load #1 #2 #3 #4 #5 #6 

2 7 1 -1 -2 -1 2 3 

4 28 2 -3 -8 -3 5 7 

6 73 -2 -17 -24 -1 28 17 

8 136 -7 -33 -47 -9 48 49 

10 222 -19 -55 -82 -10 69 98 

12 329 -36 -83 -91 -16 103 122 

14 430 -52 -91 -114 -12 114 156 

15 478 -47 -104 -115 -5 119 152 

not monitored and its position may have 

changed slightly due to slipping, thus changing 

the air gap between the coils. 

A reduction in air gap would result in an 

increase in the measured force. If this were the 

case, the air gap between the 180° opposite 

C-core should increase and correspondingly the 

force measured would be lower. Table 3 shows 

the average value of all six single core 

measurements and their deviation from the 

average. The trend of opposite C-cores have 

approximately equal and opposite deviations 

from the average is observed. (1 vs. 4, 2 vs. 5, 

3 vs. 6). This trend is also observed from 

Figure 7. 

Another explanation for the force differences 

could be a slight difference in PWM amplifier 

gain. However, all six PWM amplifiers were 

tuned identically. Also, the load cell signal 

conditioners were set-up identically and 

statically tested in the rig with a mechanical 

force applicator. 

During build-up of the facility, the air gap was 

maintained during rotor alignment by inserting 

plastic shims between the rotor and stator. 

When the rig was fully assembled, accurate 

measurement of the air gap was nearly 

impossible. In the future, a high temperature 

displacement sensor could be used to monitor 

the stator position. 

SPEED DEPENDENCE OF LOAD CAPACITY 

Data was also taken for C-core #1 at speeds up 

to 20,000 RPM and various currents. This data 

is shown graphically in Figure 9 and 

summarized in Table 4. 
NASA/TM—2002-211904 8

Figure 9 shows no force reduction due to the 

rotating laminated rotor at room temperature 

with a uniform air gap using open loop control. 

The authors recognize that there are frictional 

forces generated within the rolling element 

bearings that are not accounted for in these 

results. However, the bearings provide a good 

load path from the rotating shaft. The option of 

measuring forces at the stator (Figure 10) has 

been tried prior to this publication. The problem 

with this method was that the load cells have to 

be in line with the stator mechanical support and 

an actuator system is required. This permits 

stator deflection and position slips during 

dynamic tests. 

Since there is little dependence of force on rotor 

speed, it can be inferred that rotor surface 

speed does not cause significant eddy current 

or hysteretic effects. Both may still have 

significant effects on actuator bandwidth.

0

100

200

300

400

500

2A 4A 6A 8A 10A 12A 14A 15A

Current (Amps)

Load (lb)

0 RPM 5K RPM 10K RPM 15K RPM 20K RPM

Figure 9 – C-core #1 load as a function of rotor speed for 

bearing currents up to 15 amps. 

Table 4 – Summary of load (lb) for C-core #1 up to 20,000 

RPM. 

Shaft 

RPM 6 Amp Load 10 Amp 

Load 

15 Amp 

Load 

0 71 188 459 

5,000 61 197 459 

10,000 64 189 468 

15,000 63 204 462 

20,000 63 201 455 

POWER CHARACTERIZATION 

POWER CONSUMPTION OF SINGLE CORE 

The power required by the bearing cores as a 

function of magnetic force, current was also 

determined. At room temperature, the average 

resistance (R) for a C-core (2 coils) is 

0.48 Ohms and inductance (L) is 18mH. The 

real power used by the C-core is a function of 

the output voltage of the pulse width modulation 

amplifier (22 kHz switching frequency) and the 

instantaneous current. Figure 11 shows the 

typical power system for one C-core. 

The high switching frequency of the PWM 

amplifier made it difficult to measure the power 

required for this magnetic bearing. Normal 60 

cycle RMS power measurements cannot be 

applied. 

Due to the high cycle frequency signal, a highspeed scope (1 million samples/sec.) was used 

to record the instantaneous voltage across the 

coil and the instantaneous current through the 

coil. The isolated, high impedance scope 

collected 41,000 samples (41 ms) for the single 

C-core tests. Average power for that time was 

Figure 10 – Load measurements at the stator using 

piezoelectric actuators. 

Rectifier H Bridge

AMC PWM

130V DC

AC

PWM C-CORE UTILITY

POWER

120V AC Multifrequency AC

Figure 11 – Power system for one C-Core. 
NASA/TM—2002-211904 9

calculated using a zero power factor. When 

data was collected for one magnetic axis 

(3 C-cores), the number of samples was 

reduced to 10,000 samples (10 ms) in order to 

have a manageable size data file. 

POWER MEASUREMENTS 

Average power was calculated for individual 

C-cores and the #6-1-2 magnetic axis at zero 

RPM. The power for the magnetic axis is the 

sum of the power from the three C-cores in that 

axis. All the power calculations are shown in 

Figure 12 and summarized in Table 5. 

0

100

200

300

400

0 400 800 1200

Load (lb)

Power (W)

C-core #1 C-core #2 C-core #5

Axis 6-1-2 Axis 2-3-4

Figure 12 – Power consumption using PWM supplies at 

zero RPM.

Table 5 – Summary of power (W) using PWM power 

supplies and load (lb) as a function of input current at 

0 RPM. 

# 2 

Amp 

4 

Amp 

5 

Amp 

8 

Amp 

10 

Amp 

15 

Amp 

#1 4 W 

7 lb N/A 19 W 53 lb N/A 59 W 203lb N/A 

#2 N/A 8 W 

26 lb N/A 33 W 102lb N/A 113W 385lb 

#5 9 W 

9 lb N/A 18 W 77 lb N/A 66 W 291lb N/A 

#2-3-4 N/A 24 W 

48 lb N/A 100W 233lb N/A 380W 1048lb 

#6-1-2 8 W 

18 lb N/A 36 W 105lb N/A 142W 460lb N/A 

CONCLUDING REMARKS 

Although the shaft is not doing any work and the 

rotor is not levitated, magnetic force and 

electrical power for a high load, high speed, 

room temperature radial magnetic bearings for 

turbomachinery were measured. These 

baseline measurements are important in order 

to provide industry performance characteristics 

for actual hardware and in assessing issues 

such as non-uniform air gaps. This high 

temperature magnetic bearing can apply over 

400 lbs./axis (1.78 kN/axis) using less power 

than three household light bulbs. Loads up to 

1050 lb (4.7 kN) per magnetic axis were 

achieved. No measurable force reduction was 

observed due to rotation up to 20,000 RPM. 

The discrepancy between the theoretically 

predicted force and the measured is most likely 

the result of a non-uniform air gap (un-centered 

rotor), and or a degradation of lamination 

magnetic properties due to thermal cycling. 

Further work will address these issues and will 

include force and power measurements at high 

temperature. High temperature, high-speed 

levitation tests will also be conducted without 

the use of rolling element bearings. 

Characterization tests similar to those 

presented in this paper are underway at 

1000 °F (540 °C) and will be presented in a later 

paper. 

REFERENCES 

1.) Beams, J.W., Black, S.A., “Electrically-Driven, 

Magnetically-Supported, Vacuum-Type 

Ultracentrifuge,” Vol. 10, pp.59–63, Feb. 1939. 

2.) Kueser, P.E., Pavlovic, D.M., Lane, D.H., Clark, 

J.J., and Spewock, M. “Properties of Magnetic 

Materials for use in High-Temperature Space 

Power Systems,” NASA SP–3043, 1967. 

3.) Fingers, R.T., “Creep Behavior of Thin 

Laminates of Iron Cobalt Alloys for Use In 

Switch Reluctance Motors and Generators,” Air 

Force Research Laboratory AFRL-PR-WR-TR199–2053, 1999. 

4.) Minihan, T., Palazzolo, A., Montague, G.T., 

Kascak, A.F., Provenza, A., “Fail Safe, High 

Temperature Magnetic Bearings,” IGTI ASME 

Turbo Expo, Amsterdam June 3–6, 2002. 

5.) Choi, B., and Montague, G.T., “Fail-safe 

Operation of a High Temperature Magnetic 

Bearing for Turbomachinery,” 6th International 

Symposium on Magnetic Suspension 

Technology. 

6.) Jansen, R.H., Ebihara, B.E., Montague, G.T., 

Provenza, A.J., Jansen, M.J., Palazzolo, A., 

Tucker, R., Preuss, J., Hunt, A., “Design and 

Construction of a High Temperature, High Load 

Radial Magnetic Bearing for Turbomachinery,”

NASA TM to be published, 2002. 

7.) Introduction to Magnetic Bearings, University of 

Virginia Short Course, presented at the Third 

International Symposium on Magnetic Bearings, 

Alexandria, Virginia, 1993. 
This publication is available from the NASA Center for AeroSpace Information, 301–621–0390.

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15

Room Temperature Characterization of a Magnetic Bearing for Turbomachinery

Gerald Montague, Mark Jansen, Andrew Provenza, Ralph Jansen, Ben Ebihara,

and Alan Palazzolo

Turbomachinery; Magnetic bearing; C-core; Power factor

Unclassified -Unlimited

Subject Categories: 07, 09 and 15 Distribution: Nonstandard

WU–708–28–13–00

1L161102AF20

Gerald Montague, U.S. Army Research Laboratory, NASA Glenn Research Center; Mark Jansen, Sest, Inc., 18000 Jefferson Park Road,

Suite 104, Middleburg Heights, Ohio 44130; Andrew Provenza, NASA Glenn Research Center; Ralph Jansen and Ben Ebihara, University of Toledo, 2801 W. Bancroft Street, Toledo, Ohio 43606–3328; and Alan Palazzolo, Texas A&M University, College Station, Texas

77843. Responsible person, Gerald Montague, organization code 0300, 216–433–6252.

Open loop, experimental force and power measurements of a three-axis, radial, heteropolar magnetic bearing at room

temperature for rotor speeds up to 20,000 RPM are presented in this paper. The bearing, NASA Glenn Research Center’s

and Texas A&M’s third generation high temperature magnetic bearing, was designed to operate in a 1000 °F (540 °C)

environment and was primarily optimized for maximum load capacity. The experimentally measured force produced by one

C-core of this bearing was 630 lb. (2.8 kN) at 16 A, while a load of 650 lbs (2.89 kN) was predicted at 16 A using 1D

circuit analysis. The maximum predicted radial load for one of the three axes is 1,440 lbs (6.41 kN) at room temperature.

The maximum measured load of an axis was 1050 lbs. (4.73 kN). Results of test under rotating conditions showed that rotor

speed has a negligible effect on the bearing’s load capacity. A single C-core required approximately 70 W of power to

generate 300 lb (1.34 kN) of magnetic force. The room temperature data presented was measured after three thermal cycles

up to 1000 °F (540 °C), totaling six hours at elevated temperatures.