Gerald Montague

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

Mark Jansen

University of Toledo, Toledo, Ohio

Andrew Provenza

Glenn Research Center, Cleveland, Ohio

Alan Palazzolo

Texas A&M University, College Station, Texas

Ralph Jansen and Ben Ebihara

University of Toledo, Toledo, Ohio

Experimental High Temperature

Characterization of a Magnetic

Bearing for Turbomachinery

NASA/TM—2003-212183

March 2003

ARL–TR–2929

U.S. ARMY

RESEARCH LABORATORY
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Gerald Montague

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

Mark Jansen

University of Toledo, Toledo, Ohio

Andrew Provenza

Glenn Research Center, Cleveland, Ohio

Alan Palazzolo

Texas A&M University, College Station, Texas

Ralph Jansen and Ben Ebihara

University of Toledo, Toledo, Ohio

Experimental High Temperature

Characterization of a Magnetic

Bearing for Turbomachinery

NASA/TM—2003-212183

March 2003

National Aeronautics and

Space Administration

Glenn Research Center

Prepared for the

59th Annual Forum and Technology Display

sponsored by the American Helicopter Society

Phoenix, Arizona, May 6–8, 2003

ARL–TR–2929

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, Julie Roden (Analex, Corp.) for her thermal expansion analysis, 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.

Final recognition is for Texas A&M students Randy Tucker, Jason Preuss, and Andrew Hunt who designed and

constructed the high temperature coils used in this magnetic 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

This report is a formal draft or working

paper, intended to solicit comments and

ideas from a technical peer group.

Available electronically at http://gltrs.grc.nasa.gov

The Propulsion and Power Program at

NASA Glenn Research Center sponsored this work.
NASA/TM—2003-212183 1 

Experimental High Temperature Characterization of a 

Magnetic Bearing for Turbomachinery 

Gerald Montague 

U.S. Army Research Laboratory 

Glenn Research Center 

Cleveland, Ohio 44135 

Mark Jansen 

University of Toledo 

Toledo, Ohio 43606 

Andrew Provenza 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Alan Palazzolo 

Texas A&M University 

College Station, Texas 77843 

Ralph Jansen and Ben Ebihara 

University of Toledo 

Toledo, Ohio 43606 

ABSTRACT 

Open loop, experimental force and power 

measurements of a radial, redundant-axis, 

magnetic bearing at temperatures to 1000 °F 

(538 °C) and rotor speeds to 15,000 RPM along 

with theoretical temperature and force models 

are presented in this paper. The experimentally 

measured force produced by a single C-core 

using 22A was 600 lb. (2.67 kN) at room 

temperature and 380 lb. (1.69 kN) at 1000 °F 

(538 °C). These values were compared with 

force predictions based on a 1D magnetic circuit 

analysis and a thermal analysis of gap growth 

as a function of temperature. Tests under 

rotating conditions showed that rotor speed has 

a negligible effect on the bearing’s load 

capacity. One C-core required approximately 

340 W of power to generate 190 lb. (8.45 kN) of 

magnetic force at 1000 °F (538 °C); however 

the magnetic air gap was much larger than at 

room temperature. The data presented is after 

the bearing had already operated six thermal 

cycles and eleven total (not consecutive) hours 

at 1000 °F (538 °C). 

INTRODUCTION 

The gas turbine industry has a continued vested 

interest in improving engine performance and 

reducing net operating and maintenance costs. 

These goals are being realized because of 

advancements in aeroelasticity, materials, 

computational tools such as CFD, and engine 

simulations. These advancements also aid in 

increasing engine thrust-to-weight ratios, 

pressure ratios, specific fuel consumption, and 

overall reliability through higher efficiency 

engine operation at higher rotational speeds 

and higher temperatures. 

Rolling element bearings and squeeze film 

dampers are currently used to support gas 

turbine engine rotors. These types of bearings 

are limited in temperature (<260 °C) and speed 

(< 2.5 million DN) and require both cooling air 

and a lubrication system. Rolling element 

bearings in gas turbines are being pushed to 

their limits and new bearing technologies must 

be developed to take full advantage of other 

aforementioned advancements. 
NASA/TM—2003-212183 2 

Magnetic bearings are well suited to operate at 

elevated temperature, higher rotational speeds, 

and extreme altitudes (thin air atmosphere) and 

are a promising solution to these current 

limitations. 

Magnetic bearing technology is being developed 

worldwide and is considered an enabling 

technology for new, hotter engine designs. Xu, 

Wang, and Schweitzer developed two high 

temperature magnetic bearing test rigs, a one 

degree of freedom and a five degree of freedom 

rig to demonstrate operation of a magnetic 

bearing at high temperature [1]. A 1,500 hour 

test of a magnetic bearing for a 176 lb (80 kg), 

12,000 RPM blower rotor operating at 750 °F 

(399 °C) was completed by Ohsawa et al. [2]. 

High temperature magnetic properties of 

candidate magnetic materials were investigated 

in the sixties by Kueser et al., [3] and recently 

by Kondoleon and Kelleher [4]. This paper 

advances the wafer test specimen results from 

[3] and [4] to a laminated c-core magnet. 

Magnetic bearings in gas turbine engines would 

eliminate lubrication analysis, leaks, spills, 

contamination, and unnecessary maintenance 

due to faulty chip detection. The magnetic 

bearing could provide health monitoring and 

adaptively modify the rotor support to actively 

respond to transients such as hard aircraft 

landings and sudden imbalances. 

Magnetic bearings will enable engine designers 

to take full advantage of other technological 

advancements in turbine engine components by 

allowing rotor components to spin at higher 

speeds and at higher temperatures. As a result, 

turbine and compressor spools can be designed 

with higher operating temperatures and with 

significantly larger, faster, stiffer, highly damped 

rotors. Mekhiche developed a high temperature 

magnetic bearing system that is described from 

design through testing in [5]. This paper 

performs similar test using a different approach 

and shows detailed results and interpretation. 

The 3rd generation high temperature, high load 

magnetic bearing, developed at the NASA 

Glenn Research Center, was first characterized 

at room temperature (R.T.) up to 20,000 RPM. 

It is capable of producing over 1,000 lb (4.5 kN) 

of loading per magnetic axis at R.T. and at 

speed. R.T. data for this bearing is presented in 

[6]. 

Data presented in this paper characterizes the 

3rd generation bearing at temperatures up to 

1,000 °F (537 °C). Bearing force as a function 

of current, temperature, and speeds to 15,000 

RPM is shown. In addition, bearing power 

consumption measurements taken at several 

applied current and temperatures are 

presented. 

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 (22.8 cm) diameter with a 

maximum axial loading of 5,000 lb (22.3 kN) 

and a maximum radial loading of 2,500 lb (11.1 

kN). The test facility has been configured a 

number of different ways [6, 7]. 

Figure 1 – High temperature magnetic bearing test facility 

at NASA Glenn Research Center. 

The current configuration can be seen in Figure 

2. The magnetic bearing is located at the center 

of gravity of the 2.98 in (75 mm) diameter rotor 

that weighs 17 lbs (7.7 kg). The 8.0 in (203 

mm) stator weighs 52 lbs (23.5 kg). A 0.022 in 

(0.56 mm) radial air gap exists between the 

stator poles and the rotor at R.T. The rotor has 

interchangeable sleeves on each end that 

interface it with the support bearings, which, for 

these tests, are high-speed, grease packed, 

duplex ball bearings. For this configuration, the 

rotor was fitted with zero clearance sleeves so it 
NASA/TM—2003-212183 3 

was supported on the ball bearings at both 

ends. This was done so that forces exerted by 

the magnetic bearing could be easily and 

directly measured outside the hot section at the 

support bearing locations. 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. 

Figure 2 – Third generation high temperature magnetic 

bearing facility. 

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

described in detail in [8]. The stator has a 

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

wrapped with two coils and each coil has 52 

turns of specially insulated and potted wire. 

The coil packing factor is 0.67. 

Power to the magnetic bearing is provided 

through six tri-state, pulse width modulated 

(PWM) amplifiers. These components are 

passively filtered to remove high frequency 

amplifier noise that results from amplifier 

switching and to reduce EMI emissions from the 

bearing coils that interfere the eddy current 

position sensor’s signal. 

Heat is supplied to the bearing through three, 3 

kW band heaters wrapped around the stator OD 

(Figure 2). Ground Fault Circuit Interrupts 

(GFCI) has been incorporated into the heating 

system and the PWM amplifier circuits to 

protect the hardware and for safeguarding 

personnel. The facility is fully described in [8]. 

Figure 3 shows the facility at 1,000 °F (538 °C) 

with the stator glowing “orange” hot. 

Figure 3 - Magnetic bearing at 1000 °F (540 °C). 

SENSORS 

The facility is equipped with several different 

types of sensors that monitor load, temperature, 

rotor positions, speed, and electric current. 

High load, high bandwidth, piezoelectric load 

cells with accuracy of ±0.3 lb (±1.5 N), support 

the inboard and outboard rolling element 

bearings (Figure 4). 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 °F (196 °C). Two load cells 

are aligned along each of the three magnetic 

bearing axes on both support ball bearings for a 

total of 12 load cells. Each load cell was set 

with a preload of ~250 lb (1.1 kN). This preload 

value supplied symmetric structural rotor 

support, put the critical speed out of the 

operating range, and established a large 

dynamic range. The preload was electronically 

zeroed during testing. 

Bearing Sections Glowing 
NASA/TM—2003-212183 4 

Figure 4 – Six load cells supporting the outboard duplex 

ball bearing. 

Temperatures are recorded at several locations 

within the test rig. Each side face of the stator 

has three thermocouples mounted on it. Each 

set of support ball bearings is also equipped 

with a thermocouple. A handheld infrared 

thermometer is also used periodically to gather 

additional thermal information about the shaft 

temperatures during rotation. 

Commercially-available, 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 

(0.0012 mm) accuracy and are temperature 

compensated for the 1,000 °F (538 °C) 

environment. 

Other sensors include: an eddy current 

displacement probe used to measure rotor RPM 

and phase, magneto-resistive current sensors 

to measure C-core currents, and 

temperature/redundant RPM sensors in the air 

turbine. 

DATA ACQUISITION 

The facility data acquisition system is capable of 

capturing data at rates up to 15,000 samples 

per second per channel. Data is recorded for 

inboard and outboard bearing loads, support 

bearing and stator temperatures, X and Y axis 

displacements, rotor speed, and bearing C-core 

currents. Data is displayed in real time and can 

be saved directly into 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 outside of the hot section of the test 

facility. 

The maximum load capacity for the bearing was 

determined by calculating the vector sum of the 

forces produced by any three C-cores along the 

middle core’s centroid axis. The primary force 

component was generated by the C-core inline 

with that axis. Each C-core adjacent to the 

primary C-core also contributed a component of 

force in the direction of the axis (Figure 5). For 

example, the total load in the C-core #3 

direction is the sum of the F2 and the cosine of 

F1 and F3. The sine components of the F1 and 

F3 cancel each other out. This axis is referred 

to as the #2-3-4 magnetic axis. 

 

Figure 5 – Total load on each magnetic axis is a vector sum 

of the force from three C-cores. 

MAXIMUM FORCE PRODUCE BY A SINGLE 

CORE AS A FUNCTION OF TEMPERATURE 

Recorded data sets of load capacity vs. 

temperature indicate the maximum load 

capacity of a single core and the current level 

where saturation is reached. For these tests, Ccore #1 was attached directly to a high power 

DC supply (100 V, 40 A) and load was 

measured while current was increased. This 
NASA/TM—2003-212183 5 

was done at several different temperatures. 

“Saturation level” is defined as the point at 

which the rotor and stator laminations approach 

a permeability equivalent to that of air. This is 

the point where lamination material can not 

carry any additional magnetic flux. 

Results are shown in Figure 6. Clearly, the data 

shows that force starts to become nonlinear and 

saturation sets in at around 17 A for each 

temperature set. This data also shows that the 

load capacity decreases by 36% at 22 A 

between R.T. and 1000 °F (538 °C). In [7], 

Minihan, et al., showed a force capacity 

degradation of about 33% between R.T. and 

1,000 °F (538 °C) for the 2nd generation high 

temperature magnetic bearing. In both cases 

the heat source was on the outside of the stator 

creating a thermal gradient between the rotor 

and stator. By the conclusion of this paper, the 

data will show the reduction in force is due to an 

increase in the air gap due to thermal expansion 

and not due to lamination material properties. 

0

100

200

300

400

500

600

0 5 10 15 20 25

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 6 – Maximum load and saturation level for C-core #1 

as a function of temperature. 

INDIVIDUAL C-CORE LOAD CAPACITY AS A 

FUNCTION OF SPEED AND TEMPERATURE 

Typical load capacity of a single C-core as a 

function of current, speed, and temperature is 

presented here. C-core #2 load capacity data 

from 0 to 15,000 RPM and from 80 °F (27 °C) to 

1,000 °F (538 °C) are shown in Figures 7–10. 

0

50

100

150

200

250

300

350

400

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 7 – Load as a function of current for C-core #2 at 

zero RPM. 

0

50

100

150

200

250

300

350

400

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 8 – Load as a function of current for C-core #2 at 

5,000 RPM. 
NASA/TM—2003-212183 6 

0

50

100

150

200

250

300

350

400

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 9 – Load as a function of current for C-core #2 at 

10,000 RPM. 

0

50

100

150

200

250

300

350

400

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 10 – Load as a function of current for C-core #2 at 

15,000 RPM. 

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 12) has been tried prior to this 

publication. The problem with this method was 

that the load cells had to be in line with the 

stator mechanical support and an actuator 

system was required. This support 

configuration permitted stator deflection and 

position slip during dynamic tests. 

0

20

40

60

80

100

120

140

160

180

200

0 5 10 15

Current

Load (lb)

1000°F, 0 rpm 1000°F, 15000 rpm

F

igure 11 – Load vs. Speed comparison 

As the speed increases, rotor eddy current and 

hysteresis losses increase. However, Figure 11 

shows these losses do not affect the static load 

production capability of this bearing at elevated 

temperatures up to 15,000 rpm. These losses 

still have significant effects on actuator 

bandwidth. 

For this particular C-core, a decrease in load 

capacity of 53% is evident at, 15 A, between 

R.T. and 1,000 °F (538 °C). This is a 

considerable amount higher than previous 

results have shown. Since force is proportional 

to the inverse square of the total air gap, small 

air gap changes can greatly affect bearing 

output for any current values. It would appear 

that Figures 7–10 would have shown a more 

favorable result if they contained data shown for 

C-core #1. But, it is highly likely that the test rig 

reached an equilibrium temperature when data 

was recorded for Figure 6. 
NASA/TM—2003-212183 7 

These tests were performed immediately when 

the stator had reached the desired test point 

temperature. Band heaters heat the back iron 

of the stator and the heat travels to all other 

components primarily through conduction and 

radiation. Consequently, the rotor takes longer 

to heat and therefore the air gap is larger than 

at room temperature. In fact, air gap size is a 

direct function of the difference in temperature 

between the rotor and the stator. 

Maintaining a uniform air gap for these tests is 

difficult. Concentricity within the rig has also 

been a challenge. 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 not monitored and its position 

may have changed slightly due to slipping, thus 

changing the air gap between the coils. 

Figure 12 – Load measurements at the stator using 

piezoelectric actuators. 

LOAD CAPACITY OF AN AXIS AS A 

FUNCTION OF SPEED AND TEMPERATURE 

The previous tests were expanded upon to 

determine the load capacity of an axis. The 

maximum load of 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). 

The results of load vs. current and temperature 

for zero rpm to 15,000 rpm using magnetic axis 

2-3-4 are shown in Figures 13–16. 

0

200

400

600

800

1000

1200

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 13 – Load as a function of current and temperature 

for the #2-3-4 magnetic axes at zero RPM. 

0

200

400

600

800

1000

1200

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 14 – Load as a function of current and temperature 

for the #2-3-4 magnetic axes at 5,000 RPM. 
NASA/TM—2003-212183 8 

0

200

400

600

800

1000

1200

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 15 – Load as a function of current and temperature 

for the #2-3-4 magnetic axes at 10,000 RPM. 

0

200

400

600

800

1000

1200

0 5 10 15

Current

Load (lb)

80°F 250°F 500°F 750°F 1000°F

Figure 16 – Load as a function of current and temperature 

for the #2-3-4 magnetic axes at 15,000 RPM. 

Once again, Figures 13–16, clearly show that 

speed does not have an effect on the load 

capacity of the magnetic axis at any 

temperature — at least to 15,000 RPM. 

To confirm that the applied magnetic force was 

in same direction as the center C-core of that 

axis, the angle at which the load was applied 

was monitored. This was done by taking the 

load cell data in real-time and plotting the load 

vector direction versus true horizontal. The 

stator load angles matched the theoretical 

values (i.e. 0° for the #2-3-4 axis) to within a few 

degrees. 

POWER CHARACTERIZATION 

POWER CONSUMPTION OF SINGLE CORE 

The power required by the bearing C-cores as a 

function of magnetic force, current, temperature, 

and speed was also determined. At R.T., the 

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

0.48 ohms and inductance (L) is 18 mH. 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 17 shows the 

typical power system for one C-core. 

Rectifier H Bridge

AMC PWM

130V DC

AC

PWM C-CORE UTILITY

POWER

120V AC Multifrequency AC

Figure 17 – Power system for one C-Core. 

The high switching frequency (22 kHz) of the 

PWM amplifier makes it difficult to measure the 

power required to produce bearing forces. 

Normal 60 cycle RMS power measurements 

cannot be applied. A high-speed scope (1 

million samples/sec) was used to record the 

instantaneous voltage across the coil as well 

the instantaneous current through the coil. The 

isolated, high impedance scope collected 

10,000 samples (10 ms). Average power for 

that time was calculated using a power factor 

equal to zero. 
NASA/TM—2003-212183 9 

POWER MEASUREMENTS 

Average power as a function of temperature 

was calculated for C-core #2 and the #2-3-4 

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 18 and 19. 

0

50

100

150

200

250

300

350

400

0 100 200 300 400

Load (lb)

Power (W)

80°F 250°F 500°F 750°F 1000°F

Figure 18 – Power consumption of C-core #2 at zero RPM. 

0

250

500

750

1000

0 400 800 1200

Load (lb)

Power (W)

80°F 250°F 500°F 750°F 1000°F

Figure 19 – Power consumption of magnetic axis #2-3-4 at 

zero RPM. 

The largest power requirement was about 1 kW 

at 1,000 °F (538 °F) to produce 517 lb (2.3 kN) 

on an axis. Power required to produce a 

specific load increases dramatically with 

temperature for a number of reasons. Resistive 

power losses increase since coil resistance 

increases. A small degradation of the flux 

carrying capacity of the Hiperco laminations at 

temperature also contributes to an increase [3]. 

The other factor here stems from the increase in 

the air gap dimension. 

THERMAL ANALYSIS 

INCREASE IN AIR GAP DUE TO THERMAL 

EXPANSION 

A major contribution to the decrease in load and 

the increase in power as temperature rises is 

the expanding gap between the rotor and the 

stator. Since the heaters are mounted directly 

on the outside of the stator, it heats at a much 

higher rate then the rotor, which relies only on 

convection and radiation for heating. Thus, the 

rotor will not have expanded as much as the 

stator, and as a result, the gap will be larger 

than nominal. 

In hindsight, it would seem advantageous to 

heat the rotor as well as the stator to alleviate 

thermal mismatch problems. With the current rig 

setup, however, this is not possible. Stresses 

generated from thermal expansion of the zero 

clearance bearing sleeves might cause the ball 

bearings to seize. In addition, the grease used 

to lubricate these bearings is only rated to 

275 °F (135 °C). In a future build, the ball 

bearings will be replaced with high temperature, 

hydrostatic bearings to allow for consideration 

of other heating methods and allow for testing at 

higher speeds, since the ball bearings heat up 

within minutes to their temperature limits. 

In order to better understand the growth of the 

gap, a linear thermal expansion analysis was 

conducted. The gap at any temperature has a 

minimum and maximum value just due to the 

tolerance stack up of the rotor and stator 

bearing components. The actual gap, which is 

somewhere between the max and min, will 

change with thermal mismatch. Thermal maps 

of the magnetic bearing section of the rotor 

were generated using temperatures measured 
NASA/TM—2003-212183 10 

at several different points on the rig using a 

handheld infrared thermometer. Figure 20 

shows that, even after 2.5 hours, the average 

gap dimension is greater than it is at room 

temperature. It is quite possible that the rotor 

would never reach 1,000 °F (538 °C) using the 

current heating method. The thermal map in 

Fig. 21 shows a temperature gradient on the 

rotor. This was due to different heating rates of 

each of the three band heaters (bands not in 

100% contact with stator) and also because the 

outer containment shell was removed for these 

tests allowing for convective cooling of the 

duplex ball bearing. Figure 22 shows that the 

rotor did reach a uniform temperature of about 

780 °F (415 °C) after 2.5 hours of steady 

heating once the stator poles reached 1,000 °F 

(538 °C). 

 

0.020

0.022

0.024

0.026

0.028

0.030

0.032

0.034

0.036

0.038

0.040

Room

Temp

500°F 1000°F 21⁄2 Hours

at 1000°F

Gap (in)

Min

Max

Figure 20 – Theoretical gap between stator and rotor as a 

function of temperature using linear method. 

Figure 21 – Thermal map of rotor and stator when rig first 

reaches 1000 °F 

Figure 22 - Thermal map of rotor and stator after rig at 

1000 °F for two hours 

In this model, the average radial gap at room 

temperature is 0.0244 in. (0.620 mm). This gap 

increases to 0.0312 in. (0.792 mm) at 1,000 °F 

(538 °C). Since the magnetic force is a function 

of the inverse square of the distance, it is clear 

why the force at elevated temperature drops 

more severely than shown in [7] for a fixed 

supply current. 

THEORETICAL FORCE PREDICTION 

For a general magnetic actuator, the force 

produced by the magnetic flux in the air gap of 

area Ag is given by:

2

2 2

1

























+

=

r

iron

o g L G

NI F A

μ

αμ (1) 

In this equation, uo is the magnetic permeability 

of free space and NI is the ampere-turns of the 

coil(s) driving flux through the circuit. 2G 

represents the total effective magnetic air gap 

and α is a force de-rating factor, which 

accounts for leakage, fringing, non-uniformity in 

flux density across the pole face and finite path 

permeability [9]. Liron represents the average 

iron path component of the C-core circuit; ur

 is 

the relative permeability of the iron laminations. 
NASA/TM—2003-212183 11 

The actuator measurements in this paper are 

based on the performance of a complete C-core 

circuit. Therefore, the forced produced by a 

single C-core is almost twice that obtained using 

Equation 1, since there are two air gaps 

between it and the rotor. Each of the two C-core 

poles is 15° off axis, therefore, the force 

produced by a single C-core is 1.93 times that 

obtained using Eq. 1. 

Using α = 0.8, uo = 4π x10-7 N/A2

, N = 52*2 

turns, Ag = 2.0 in2 (12.9 cm2), G = 0.022 in 

(0.559 mm), Liron = 5.7 in (145 mm), ur

 = 500. 

Figure 23 shows a comparison between the 

predicted force using the maximum gaps from 

the thermal analysis and the actual loads versus 

the measured current. 

0

50

100

150

200

250

300

350

400

450

500

0 5 10 15

Current (Amps)

Load (lb)

80°F Calc 500°F Calc

1000°F Calc @1000°F 2 hrs Calc

80°F Data 500°F Data

1000°F Data @1000°F 2 hrs Data

Figure 23 – Theoretical force prediction versus actual 

measured values for C-core #1 

There is good agreement between the 

calculated load and the actual measurements 

when the variation in gap is considered. 

CONCLUDING REMARKS 

Magnetic applied force and electric power for a 

high load, high speed, high temperature radial 

magnetic bearing for turbomachinery was 

measured. Even though the shaft is not doing 

any work and the rotor is not levitated, these 

tests provide baseline industry performance 

characteristics for actual hardware. Load 

capacity and power consumption of both a 

single C-core and a magnetic axis were 

measured from 80 °F (27 °C) to 538 °C. 

Thermal analysis of the facility to investigate 

gap growth due to non-uniform heating was 

performed. Theoretical force predictions were 

compared to actual results with good 

correlation. No appreciable decrease in static 

force production capability was observed due to 

rotation up to 15,000 rpm at any temperature. 

These initial results suggest that the reduction in 

static force capacity at 1000 °F (538 °F) due to 

lamination material properties may not be 

significant. This work also gives practical 

insight into the range in force capacity that can 

occur in actual hardware if changes in the rotorstator gap due to misalignment or thermal 

effects are not addressed. These factors and 

the stator mounting technique have to be 

considered in designing magnetic bearings for 

high temperature applications. 

Future work will include high temperature, highspeed levitation tests without the use of rolling 

element bearings and high temperature force 

measurements of a magnetic thrust bearing. 

REFERENCES 

1) Xu, L., Wang, L., and Schweitzer, G., 

“Development of Magnetic Bearings for High 

Temperature Suspension,” Seventh 

International Symp. on Magnetic Bearings, ETH 

Zurich, August 23–25, 2000. 

2) Ohsawa, M., Yoshiolo, K., Ninomiya, H., et al., 

“High Temperature Blower for a Molten 

Carbonate Fuel Cell Supported by Magnetic 

Bearings,” 1998, Proceedings of the 6th ISMB. 

3) Kueser, P., et al., “Properties of Magnetic 

Materials for Use in High-Temperature Space 

Power Systems,” NASA SP–3043, 1967. 

4) Kondoleon, A., Kelleher, W., “Soft Magnetic 

Alloys for High Temperature Radial Magnetic 

Bearings,” 7th Inter. Symp. On Magnetic 

Bearings, pp. 111–116, ETH Zurich, 

Switzerland, August 23–25, 2000. 

5) Mekhiche, M., Nichols, S., Oleksy, J., Young, J., 

Kiley, J., Havenhill, D., “50K RPM, 1100 °F 

Magnetic Bearings for Jet Turbine Engines,”

Seventh International Symp. on Magnetic 

Bearings, ETH Zurich, Switzerland, August 23–

25, 2000. 
NASA/TM—2003-212183 12 

6) Montague, G.T., Jansen, M.J., Provenza, A.P., 

Jansen, R.H., Ebihara, B., Palazzolo, A., “Room 

Temperature Characterization of a Magnetic 

Bearing for Turbomachinery,” NASA/TM—2002-

211904, Sept, 2002. 

7) Minihan, T., Palazzolo, A., Provenza, A., 

Montague, G.T., Kascak, A.F., “Fail Safe, High 

Temperature Magnetic Bearings,” IGTI ASME 

Turbo Expo, Amsterdam June 3–6, 2002. 

8) 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. 

9) 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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March 2003

NASA TM—2003-212183

ARL–TR–2929

E–13784

18

Experimental High Temperature Characterization of a Magnetic Bearing

for Turbomachinery

Gerald Montague, Mark Jansen, Andrew Provenza, Alan Palazzolo, Ralph Jansen,

and Ben Ebihara

Turbomachinery; Magnetic bearing; C-core; Power factor

Unclassified -Unlimited

Subject Categories: 07 and 37 Distribution: Nonstandard

WBS–22–708–28–12

1L161102AF20

Prepared for the 59th Annual Forum and Technology Display sponsored by the American Helicopter Society, Phoenix,

Arizona, May 6–8, 2003. Gerald Montague, U.S. Army Research Laboratory, NASA Glenn Research Center; Mark

Jansen, Ralph Jansen, and Ben Ebihara, University of Toledo, Toledo, Ohio 43606; Andrew Provenza, NASA Glenn

Research Center; 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 radial, redundant-axis, magnetic bearing at temperatures to

1000 °F (538 °C) and rotor speeds to 15 000 RPM along with theoretical temperature and force models are presented in

this paper. The experimentally measured force produced by a single C-core using 22A was 600 lb. (2.67 kN) at room

temperature and 380 lb. (1.69 kN) at 1000 °F (538 °C). These values were compared with force predictions based on a

1D magnetic circuit analysis and a thermal analysis of gap growth as a function of temperature. Tests under rotating

conditions showed that rotor speed has a negligible effect on the bearing’s load capacity. One C-core required approximately 340 W of power to generate 190 lb. (8.45 kN) of magnetic force at 1000 °F (538 °C); however the magnetic air

gap was much larger than at room temperature. The data presented is after the bearing had already operated six thermal

cycles and eleven total (not consecutive) hours at 1000 °F (538 °C).