Andrew J. Provenza

Glenn Research Center, Cleveland, Ohio

Kirsten P. Duffy

University of Toledo, Toledo, Ohio 

Experimental Methodology for Determining 

Turbomachinery Blade Damping Using Magnetic 

Bearing Excitation and Non-Contacting 

Optical Measurements

NASA/TM—2010-216770

October 2010

GT2010–22093
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Andrew J. Provenza

Glenn Research Center, Cleveland, Ohio

Kirsten P. Duffy

University of Toledo, Toledo, Ohio 

Experimental Methodology for Determining 

Turbomachinery Blade Damping Using Magnetic 

Bearing Excitation and Non-Contacting 

Optical Measurements

NASA/TM—2010-216770

October 2010

GT2010–22093

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Prepared for the

Turbo Expo 2010

sponsored by the American Society of Mechanical Engineers (ASME)

Glasgow, Scotland, United Kingdom, June 14–18, 2010
Acknowledgments

The authors would like to thank Efrain Patino for his assistance in manufacturing hardware for damping tests described herein.

Available from

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

Trade names and trademarks are used in this report for identifi cation 

only. Their usage does not constitute an offi cial endorsement, 

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Space Administration.

This work was sponsored by the Fundamental Aeronautics Program 

at the NASA Glenn Research Center.

Level of Review: This material has been technically reviewed by technical management. 
NASA/TM—2010-216770 1

Experimental Methodology for Determining Turbomachinery 

Blade Damping Using Magnetic Bearing Excitation 

and Non-Contacting Optical Measurements

Andrew J. Provenza

National Aeronautics and Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Kirsten P. Duffy

University of Toledo

Toledo, Ohio 43606

Abstract

Experiments to determine the effects of turbomachinery 

fan blade damping concepts such as passively shunted 

piezoelectric materials on blade response are ongoing at the 

NASA Glenn Research Center. A vertical rotor is suspended 

and excited with active magnetic bearings (AMBs) usually in 

a vacuum chamber to eliminate aerodynamic forces.

Electromagnetic rotor excitation is superimposed onto rotor 

PD-controlled support and can be fixed to either a stationary 

or rotating frame of reference. The rotor speed is controlled 

with an air turbine system. Blade vibrations are measured 

using optical probes as part of a Non-Contacting Stress 

Measurement System (NSMS). Damping is calculated from 

these measurements. It can be difficult to get accurate 

damping measurements using this experimental setup and 

some of the details of how to obtain quality results are 

seemingly nontrivial. The intent of this paper is to present 

those details. 

1.0 Introduction

NASA’s Fundamental Aeronautics Program is interested in 

pursuing research in technology areas which facilitate the 

development of lighter, quieter and more efficient fans for 

propulsion applications. One such application is commercial 

fixed wing, powered flight. High performance fan blades 

designed to achieve such goals will be subjected to higher 

levels of aerodynamic excitations which will undoubtedly 

lead to more serious and complex vibration problems (Ref. 1). 

There is concern over the levels of vibratory stresses, the 

reduction of high cycle fatigue (HCF), and aeroelastic 

instabilities. Furthermore, blade material and mechanical 

damping become issues of concern when material is removed 

from the blade design and when advanced stage design 

increases the number of vibration modes within the engine 

operating speed range (Ref. 1). 

Historically, NASA Glenn Research Center has used the 

Dynamic Spin Rig (DSR) to characterize the vibrations of fan 

blades and bladed structures and to study damping. In the past 

several years, this facility has been used to assess the effects 

of advanced blade damping concepts on flat plates and 

subscale fan blades. Damping concepts, such as viscoelastic 

material patches (Ref. 2), impact dampers (Ref. 3), plasmasprayed damping coatings (Ref. 4), and high-damping hightemperature shape memory alloy materials (Ref. 5) are 

typically investigated in this facility. Recent experiments 

explore the possibility of using piezoelectric materials in the 

form of patches to dampen problematic blade vibration modes

(Refs. 6 and 7). 

To facilitate the effective testing of damping concepts, an

active magnetic bearing (AMB) support system was designed 

for the DSR with the intent of using it to excite blade

vibrations (Refs. 8 to 10). Modifications to the DSR bearing

electronics and controls since last publication, including some 

used by Jansen, et al. (Ref. 11) and Dever, et al. (Ref. 12) will 

be discussed. Maturation of AMB techniques to excite the 

blades along with those historically used to obtain damping 

measurements are also presented.

2.0 Dynamic Spin Rig Facility

The DSR can be used for a variety of turbomachinery 

studies which include:

• Determining the effects of centrifugal loading on blade 

structural properties and behavior. 

• Determining the feasibility of new damping concepts. 

• Determining the effects of rotation on candidate damping 

system components utilized in the rotating frame.

• Determining structural damping coefficients versus

rotational speed, which can be used in aeroelastic finite 

element analysis codes. (Structural damping data 

determined experimentally can be used analytically to 

determine blade and vane response to unsteady loading).

2.1 Test Facility

Test articles, up to 80 cm diameter, usually consisting of a 

rigid hub and two blades, are mounted on a vertical rotor 

simply connected to an air turbine with a pin in slot. The rotor 

is supported on three AMBs (Fig. 1). An axial magnetic 
NASA/TM—2010-216770 2

bearing carries the weight of the rotating system. The two 

other magnetic bearings provide radial support. Grease-packed 

ball bearings provide back-up rotor support in the event of an 

AMB failure or instability which results in rotor excursions 

beyond a 0.25 mm radius. All rotating and support hardware 

are attached to the underside of the vacuum tank lid. Together, 

the lid and test hardware are lowered into a 1.22 m diameter, 

2.54 cm thick steel vacuum chamber for testing (Fig. 2). The 

tank has a 7.62 cm thick safety ring at the top to contain 

hardware in the event of a failure while spinning. A vacuum 

level of 70 Pa is typical.

Blade excitation is achieved using the same radial AMBs 

which maintain rotor support and control. Details of this 

excitation system will be discussed in Section 5. 

2.2 Other Excitation Methods

Several excitation methods have been used in the past, and 

are still available for use. These include a permanent magnet 

eddy current excitation system, air jet excitation, and electromechanical shakers. The intent of this work is to use AMBs to 

replace these other methods.

Permanent magnets have been used to excite titanium 

blades at fixed engine orders. The excitation frequency is 

fixed to the engine order defined by the number of magnet 
NASA/TM—2010-216770 3

circuits. To investigate other engine orders requires a 

mechanical reconfiguration of the magnets. A ring, holding

these magnet circuits arranged uniformly around the 

circumference, is lifted using precision linear actuators 

towards the blades to excite them (Fig. 3). Eddy currents are 

generated within the blades to combat the change in magnetic 

flux encountered as they pass. This produces a force on the 

blades and on the rigidly supported magnets. The excitation 

force magnitude is a function of the conductivity of the blade 

material, the magnitude and direction of the magnetic field, 

and the speed of the blade relative to the stationary magnet 

circuit. Excitation force increases as the magnets are lifted 

closer to the blades because the reduction in air gap between 

the blade and circuit increases the density of the magnetic 

field. Often, to achieve blade resonant responses large enough 

to be effectively measured, highly electrically conductive 

coatings such as silver are added to the excitation regions of 

the blades. These coatings however, can alter blade damping. 

The main drawback of this method is blade heating. In an

evacuated tank, there is negligible convective and radiant 

cooling. The blades rely on conduction to dissipate the 

localized heating. It is possible to melt the tips of blades using 

this method (Ref. 13).

Air jets placed in the vacuum chamber have been used. In 

previous tests, two jets 180° apart were used to excite blade 

vibrations. Disadvantages of this system include the 

reestablishment of aerodynamic loads and rig vacuum pump 

inability to maintain a vacuum pressure lower than 300 Pa.

Electromechanical shakers have also been used to excite 

blade vibration in a stationary frame. Axial oscillatory forces 

or moments are applied to the shaft through soft axial 

mechanical bearing supports (Ref. 14). Blade motion is 

induced by shaft motion and blade vibration amplitude varies 

with blade circumferential position. As with magnetic bearing 

excitation, blade natural frequencies can be excited 

independent of rotor speed. The soft mounting of the rotor 

supports, which facilitates shaft motion, introduces high levels 

of damping which contributes in some unknown amount to 

damping measurements of the test blade article response 

(Ref. 14).

The magnitudes of blade response to this form of excitation 

were often too low making it difficult to extract relevant 

information from strain gage data (Ref. 15). Problems like 

these led to the design of the current AMB excitation system.

Oil-jet spray excitation is a method not used at GRC, but 

has been used successfully by others. The U.S. Navy rotor 

spin facility uses multiple oil-jets with externally-controllable 

valves to excite various Campbell Diagram crossings up to 

64E without stopping to reconfigure (Ref. 13). When using 

oil-jets, NSMS measurements are made with eddy current 

probes as opposed to light probes which are affected by the oil 

mist in the vacuum chamber.

3.0 Magnetic Bearings and Control

Figure 1 shows the location of the magnetic bearings, eddy 

current displacement sensors, speed pickups, and auxiliary 

bearings. The thrust bearing, located at the top of the rotor 

system, is constructed of non-laminated components made 

from a soft magnetic alloy called Hyperco. Since the flux 

carrying components are not laminated, this thrust bearing has

a low bandwidth, on the order of 50 Hz. This bearing was not 

intended to provide excitation to the rotor, but to simply 

support the rotor to eliminate possible frictional heating due to 

physical contact and to eliminate the potential for axial 

support effects on blade vibration damping. Axial bearing 

control derivative, which provides AMB control damping, is 

minimized to reduce its potential effects on same. The radial 

bearings are 4 pole with laminated cores and a force output 

capacity of approximately 1 kN each. AMB cores, like 

transformers, are laminated to increase bandwidth by 

hindering the development of large time constant eddy 

currents. This has the side benefit of reducing power loss and 

thus heat generation as well.

Prior magnetic bearing control algorithms were coded in 

Turbo Pascal or C and shared resources on the resident 

computer CPU. Control loop times, which were much faster 

than DSP-based systems, were unfortunately not constant. 

This made prescribing rpm tracking excitation from within the 

controller difficult. As a result, an extra input channel was 

needed to facilitate use of an external function generator. A 

dSPACE DS1005 system is currently used to establish control 

and facilitate robust excitation of the rotor. Compiled 

controllers run on a dedicated PowerPC chip. The AMB 

controller is coded in Simulink and also C inside of SFunctions. Fixed loop times on the order of 10 μsec are typical 

using this approach.
NASA/TM—2010-216770 4

3.1 Controller Tuning

A conventional linearized independent PD control scheme 

is implemented. Figure 4 shows the dSPACE ControlDesk

interface for manipulating the control parameters. For each 

AMB, the bias current, rotor position, stiffness (P), and 

damping (D) can be set here for each axis of operation. For 

example, each radial bearing has two axes and the PD values 

are independently set. Typically, the bias is set higher than the 

excitation amplitude to maximize excitation level and to avoid 

excitation harmonics. Stiffness and damping values are set to 

provide a stable rotor support and to minimize rotor 

excursions during excitation. Damping values are also 

minimized to reduce the possible effects of the controller in 

damping blade vibrations. All control parameters are kept 

consistent to avoid possible effects of AMB control on 

damping measurements.

3.2 Notch Filtering

Notch filtering of feedback control signals has proven to be 

an effective way to prevent the excitation of fundamental rotor 

system modes by sensor and other control system noise 

sources (Ref. 12). In this application, the shaft which floats in 

its support, is often lightly damped. Noise in the feedback 

control system tends to excite the rotor’s first and/or second 

bending modes. Shaft ringing depends heavily on the construct 

of the rotor system. In a passive shunt experiment where 

several inductor coils were stacked with rubber disk separators 

and housed at the bottom of the rotor, no notch filters were 

required. This heavy mass at the bottom of the rotor shifted 

these frequencies lower and provided sufficient internal 

damping. To circumvent problems of this nature, fixed infinite 

notches centered about these frequencies are used to remove 

signal content at resonance frequencies from control feedback

sensor signals. Figure 4 shows the user controls for these

notches. Although provisions for a moving notch are coded, 

only fixed notches are currently used. The location of the 

notch is set by inputting the center frequency. The width of the 

notch, can also be adjusted.

In simple PD control, the derivative of rotor position is used 

to generate damping. This derivative serves to amplify noise 

present in the eddy current position sensors. This may be the 

culprit in bending excitation. Alternative control approaches, 

such as that described in (Ref. 16), could improve AMB 

control and simplify it by eliminating the need for fixed 

notches located at rotor resonances. 

3.3 PWM Switching Amplifier Filtering

The eddy current displacement sensors measure the rotor 

shaft excursions near each radial AMB in x and y directions.

Eddy current sensor noise is influenced by high frequency flux 

generation around the bearing’s electromagnetic coils. Pulse 

Width Modulation Amplifier (PWM) switching effects on 

eddy current sensors are discussed in detail in (Ref. 11). PWM 

output edge filters were installed in the DSR to reduce highfrequency magnetic field content broadcast by the AMBs and 

picked up by the position sensors. 

3.4 Imbalance Compensation

The control software includes a manual imbalance 

compensator for both radial bearings. A thorough description 

of feed forward imbalance compensation can be found in 

(Ref. 12). A force vector synchronous with the rotor speed is 

generated by first prescribing its magnitude and then searching 

for the proper phase angle (referenced to a 1/rev) that 

minimizes the radial imbalance magnitude, at the radial 

bearing proximity probe locations, calculated using a digital 

vector filter. Force vector magnitude is then optimized to 

reduce the orbit further. Rotor orbits at the sensor locations are 

typically kept under 5 μm during operation. Although the shaft

has been spin balanced outside of the DSR, test articles and 

damping hardware of varying mass and uniformity are 

interchanged frequently during testing. The rotor system, as a 

result, is often out of balance and either needs to be removed 

completely from the rig and rebalanced, balanced in-situ using 

trial and correction weights, or balanced directly during testing 

using AMB feed forward compensation. The first two options 

can be time consuming and extend test program duration. The 

third saves time and has proven to be successful. Recent tests 

have had a rotational speed limit of 4,000 rpm, and 

compensation for any imbalance has been successful.

Minimizing rotor orbits during rotation and excitation 

should improve NSMS blade displacement measurements, 

facilitate a greater transmission of excitation force to the blade 

hub, and reduce the contribution of other damping 

mechanisms from the blade vibration measurements.
NASA/TM—2010-216770 5

3.5 Excitation Limitations

The spin rig radial magnetic bearings were salvaged from a 

homopolar bearing used in prior research. These stators each 

have a force production capability of about 1 kN. Because the 

radial bearings operate in a vacuum, stator coil temperatures 

must be continuously monitored to assure that the wire 

insulation does not melt, producing shorts which could cause 

instability while spinning and possible permanent bearing and 

test rig damage. The radial bearing coils do heat up during 

testing. The higher the excitation amplitude, the faster they 

heat up. AMB coil temperatures have posed a problem in spin 

rig testing. When coil temperatures reach 80 °C, excitation is 

turned off and the bearings are allowed to cool. In addition, 

the AMBs have a frequency bandwidth of approximately 

1 kHz, limiting the blade modes that can be excited.

4.0 NSMS Measurement System

The current spin rig setup, which utilizes full magnetic 

suspension, does not have a slip ring or telemetry system 

installed. No capability to monitor blade responses using strain 

gages exists. In this configuration, blade responses are 

determined solely using a non-contacting technique using 

lasers. IFOSYS optical laser spot probes are used to provide 

blade tip time of arrival information relative to an optical 1/rev 

located at the rotor mid-span close to the test article axial 

plane of rotation. Multiple probes are aligned radially with the 

center of rotation and normally equally spaced around a 

portion of the circumference of the stage. As the blades rotate 

by, tip timing data is sent to a non-contacting stress 

measurement system where the data is converted into 

displacements. From the blade displacements, blade response 

and damping can be determined. Hood Technologies NSMS 

Analyze software can be used to calculate a damping quality 

factor (Q). Time history data of blade ringdown after the 

excitation is removed can be exported and damping calculated 

from the vibration decay.

The title non-contacting stress measurement system is 

really a misnomer—at least in the way it is used in this 

application. Herein, response at the blade tips is what is 

required. In other applications, tip response is correlated with 

blade strain acquired simultaneously using strain gages 

attached to the blade. Predicted blade stresses associated with 

an experimental blade strain field are then coupled to tip 

responses which allows for the sole use of a NSMS system 

and elimination of the need for strain gages over the entire 

duration of a test.

5.0 Blade Base Excitation Techniques

Blades are excited by direct excitation of the hub. The 

radial magnetic bearings serve a dual purpose in both 

maintaining rotor position and control and also providing 

blade excitation. Single frequency sinusoidal excitation is 

most commonly used, however, random or white noise is 

available and has been used to excite many modes at once. 

Since two bearings are used to generate excitation, various 

combinations of both forces and moments can be generated. 

Operating in a cylindrical or bounce mode, both bearings 

produce excitation force in phase. In a conical or tilt mode, the 

forces are out of phase and produce a resulting moment. Prior 

experience indicates that bounce excitation is better at exciting 

blade bending modes and tilt is better for torsional modes. 

However, this conclusion is for vertical flat plate testing and 

may not apply to twisted or pitched blades.

Hub-based excitation using radial AMB’s as described 

herein is used for rotor configurations which include only two 

blade specimens (Fig. 1). The current focus is on individual 

blade damping. All fundamental modes for this configuration 

are observable and controllable at the hub (Ref. 17). If they are 

controllable, then they are excitable. However, certain blade 

modes in rotor configurations with more than three blades are 

not excitable using this method. Any circumferential modal 

pattern which produces no net inertial force or moment on the 

hub will not be excitable. Axial excitation using AMBs could 

be added to excite modes where a net axial force is required 

(Ref. 17).

5.1 Fixed Excitation

Excitation can be fixed to a stationary or rotating frame of 

reference. In the fixed frame, the excitation that the blades 

experience, is related to the excitation frequency fe and the 

shaft rotation speed fr in the following way: fb = fe ± fr, 

resulting in two distinct excitation frequencies on the blades.

This reduces the force level that the blades experience at the 

frequency desired. For this reason, the rotating excitation 

method was developed.

In fixed excitation (see Fig. 5), Fexcite is stationary and,

Fexcite = A sin(2πfet) (1)

where A represents excitation force amplitude. Fnormal, which 

is the component of fixed excitation (shown perpendicular to 

the blade row) that excites blade vibration, can be expressed 

as,

Fnormal = Fexcite sin(Ωt) (2)

Substituting (1) into (2) with Ω=2πfr, 

Fnormal = A sin(2πfet) sin(2πfrt) (3)

Using a trigonometric identity,

Fnormal = A/2[sin(2πfet – 2πfrt) + sin(2πfet + 2πfrt)] (4)

Since fe = N*fr where N is an integer representing the desired 

engine order (EO) excitation, 
NASA/TM—2010-216770 6

Fnormal = A/2[sin[2π (N – 1)fr t] + sin[2π(N + 1)fr t]] (5)

Note that in order to achieve the proper excitation frequency 

in this mode of operation, fe must be offset. N must be 

prescribed one EO above or below that desired. The blades 

will still be excited at two frequencies. Also note that the 

initial amplitude, A, is shared by the two sine functions.

5.2 Rotating Excitation

A unique feature of the AMB system is the ability to rotate 

the excitation with the blades. In this case, Fnormal is computed 

directly, thus the excitation frequency fe = fb. There is no 

compensation for rotor speed. In the rotating frame, the blades 

will experience a single excitation frequency. A radial or 

transverse excitation force which follows the rotation of the 

rotor system is generated in the stationary frame of reference. 

This force vector is locked synchronously to the rotational 

speed and is phase controllable. An optimum excitation phase 

angle that produces the highest blade vibration amplitude for a 

given rotational speed and resonant frequency is currently 

determined manually. 

Typically, a small to moderate excitation amplitude is 

prescribed while the rotor is spinning in the rpm range of 

interest for the proposed experiment at a speed where an 

engine order line crosses a fundamental resonance of the test 

article (Fig. 9). First bending is typically used because it is the 

easiest to excite and produces a relatively large tip 

displacement. The excitation vector phase angle is altered until 

the maximum blade response is achieved. The manual 

procedure described was implemented for bounce and tilt 

excitation. This method could easily be automated if strain 

gage or NSMS computation feedback signals were available to 

the controller.

In fixed excitation, an excitation force generated at a 

specific frequency in the stationary frame produces an 

excitation force with two frequencies symmetric about Nfe

offset by fr. The inverse is also true and will be shown below: 

An excitation force with a single frequency generated in the 

rotating frame is produced by a force with two distinct 

frequencies symmetric about Nfe and offset by fr.

In Figures 6 and 7, orthogonal excitation current command 

components produced by the magnetic bearing controller are 

plotted for representative even and odd engine orders, 

respectively. These patterns are generated to produce a zero 

median sinusoidal excitation which is synchronized with rotor 

speed. In the 6-times-the-rotational-speed (6EO) example 

(Fig. 6), 12 lobes are visible. This is expected as for every sine 

wave there is a plus and a minus lobe. However, in the 5 EO 

example (Fig. 7), only 5 lobes are visible. In this case, there 

are actually 10 lobes—the pattern for the second half of a 

revolution has simply been repeated and sits directly on top of 

the first.

Figure 8 gives reassurance that the proper excitation is 

being produced in the rotating frame. This Fast Fourier 

Transform (FFT) magnitude and envelope plot provides the 
NASA/TM—2010-216770 7

frequency content of the magnetic bearing control current 

(generated in the stationary frame) measured at the output of 

the PWM amplifier associated with a lower bearing control 

axis. Two side band peaks are shown straddling the rotating 16 

EO excitation prescribed. In this example, the rotor speed (fr) 

was 2783 rpm (46.4 Hz). Excitation frequency, fb, was 16 

times that or 742.5 Hz. Peak 1 is 696 Hz and Peak 2 is 789 Hz, 

which are exactly 16 EO ± fr.

The red curve in Figure 8 shows FFT peak history 

information, over a specified rotor speed range, extracted from 

the entire data file. It has been included to qualitatively show the 

effects of magnetic bearing control and excitation system 

bandwidth on rotating excitation system amplitude. The 

negative sloping lines above the two excitation frequency peaks 

show that for a fixed, prescribed excitation amplitude, actual 

applied current amplitude will decrease with rotational speed. 

Over a set test rpm range, this effect can be accommodated to

preserve a fixed excitation either manually or in software. This 

feature has not yet been adopted in the DSR. 

The phase angle between rotor position and synchronous 

excitation also changes with rpm. As explained, excitation 

force angle with respect to the 1/rev signal is set to maximize 

response. The optimum is determined manually at a single 

rotor speed and excitation frequency. As the excitation 

frequency changes, so too does this angle. As the angle 

changes the force vector effectiveness in producing a response 

will change. This effect also could be compensated for 

manually or by using an automated scheme. 
NASA/TM—2010-216770 8

5.3 Engine Order Jumps

A technique being called engine order jumps was developed 

to improve the acquisition of blade response data in a timely 

manner. Magnetic bearing excitation makes this rotating force 

technique possible. Engine order jumps, while using rotating 

forced excitation, cannot be easily implemented with other 

previously mentioned techniques (magnetic ring, air jets). It 

should be possible to use piezoelectrics to achieve similar 

capability, and ways of making this possible in a rotating 

frame are being explored. 

In this technique, turbine acceleration and deceleration are 

set as low as practically possible which is approximately 

±3 rpm/sec in this facility. At speeds midway between 

locations where an EO line will intersect the target blade 

resonant frequency line (Fig. 10), the integer excitation EO 

multiplier value is incremented or decremented opposite the 

direction of rotation speed change. 

The excitation tracks the EO line as rpm changes causing 

the blade to vibrate at the resonance of interest, then continues 

on until commanded to jump to the next line. Figure 11 shows 

an example of blade response versus time generated using this 

technique during a slow acceleration from 2000 to 3800 rpm. 

5.4 RPM Hovering

Another technique developed to get useable blade response 

data is called RPM Hovering. When blade modes are lightly 

damped, 3 rpm/sec can be too fast for the Engine Order Jump 

Technique to produce blade responses which represent a full 

steady state response (Ref. 18). In RPM Hovering, rpm dwells 

with very small rpm excursions up and down are used to 

trigger a response at a slower ramp rate. RPM ramp rates 

lower than the 3 rpm/sec are achievable when the difference 

between the actual rotor speed and commanded rotor speed are 

on the order of 10 rpm. When the amplitude of a synchronous 

blade response jumps on the Hood Technology Acquire

Software Display (Fig. 12), the AMB excitation is 

immediately turned off. The ringdown of the blade is captured 

and used to provide an alternative measure of damping. Figure 

13 shows an example of multiple ring downs recorded during 

an RPM Hovering test. In this example rpm hovered around 

1975 rpm.

6.0 Damping Calculations

Response peaks from rpm sweeps during engine order 

jumps, such as shown in Figure 11, are used to determine the 

level of damping. The Analyze software directly determines 

the quality factor Q and resonance frequency from the engine 

order response data. Q levels from damped and undamped 

blades are compared to determine the effectiveness of a 

particular damping technique. It is necessary to assume a 

single-degree-of-freedom system for this calculation, so the 

modes under investigation should be well-spaced. For 
NASA/TM—2010-216770 9

example, the plates tested in this study were designed to have 

no frequency veerings within the speed range of this test (e.g.,

third bending and second torsion modes in Fig. 9).

If the blade damping is very low, the rpm ramp rate may be 

too high to capture a good response peak. In this case, ringing 

will occur in the response. Therefore, the rpm hovering 

technique was used. The ringdowns shown in Figure 13 can be 

used to determine resonance frequency and damping. This is 

not a typical ringdown, because the NSMS measurement 

technique yields highly aliased data. There may be several 

blade oscillations between data points. However, since the 

excitation frequency is known at each data point, a fit can still 

be made to this ringdown, as shown in Figure 14. The “blade 

1” data is the NSMS measurement with a time between data 

points of 0.0198 s. The “fit” data is the least-squares singledegree-of-freedom fit with time between data points of 

0.0005 s. The “aliased fit” is the same least-squares singledegree-of-freedom fit with data points only at the NSMS ∆t of 

0.0198 s. The Q of 822 calculated is very low, and would have 

required an rpm sweep rate of less than 1 rpm/sec.

Comparisons between ringdown and frequency response data 

have been made, with fairly good correlation.

This method has been applied to linear damping treatments 

only. Extending it to a nonlinear damping system such as 

impact damping (Ref. 3) may be problematic. Impact damping 

is a nonlinear function of displacement amplitude rather than a 

constant value. The aliased fit calculates displacement 

amplitude versus time, and a single damping value for the 

specified time period. It would be difficult to accurately 

calculate both amplitude versus time and damping versus time, 

although the technique could yield qualitative information 

(e.g., vibration envelope shape).

The next set of tests planned for the DSR will include a slip 

ring, allowing strain gage and accelerometer measurements to 

be made. It is expected that more accurate damping 

calculations can be made with this instrumentation; however, 

the tests described in this paper do show that measuring 

damping with NSMS alone can be accomplished. 

7.0 Concluding Remarks

The DSR, which uses a five-axis AMB system to support 

and excite rotating test articles to 80 cm in diameter and 

20,000 rpm in a vacuum, is being used for structural 

turbomachinery blade vibration damping experiments at GRC. 

Details of excitation and speed control techniques developed 

to obtain quality damping measurement were discussed. Two 

such techniques, Engine Order Jumps and RPM Hovering,

were used in recent damping tests of passively-shunted 

piezoelectric patches glued to rotating flat plate specimens. 

Blade tip timing data is obtained using NSMS probes and 

converted to displacement and damping quality factor Q using 

commercially available software or exported and postprocessed by other means. Comparisons between ringdown 

and frequency response data have been made, with fairly good 

correlation. Possible excitation control improvements to 

preserve amplitude and phase during changes in rotational 

speed have been briefly discussed and yet to be implemented 

in the DSR. 

References

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Cycle Fatigue of Turbomachinery Components—Industry 

Perspective,” AIAA–1997–3365, 1997.

2. Kosmatka, J.B., Appuhn, G., “Design and Fabrication of 

Integrally Damped Composite Fan Blades,” SPIE Conference on 

Passive Damping and Isolation, Newport Beach, CA, SPIE Vol. 

3672, Mar. 1999.

3. Duffy, K., Bagley, R., Mehmed, O., “On a Self-Tuning Impact 

Vibration Damper for Rotating Turbomachinery,” 36th Joint 
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Propulsion Conference, Huntsville, AL, AIAA–2000–3100, July 

17–19, 2000.

4. Duffy, K., Miller, R., Nguyen, Q., Fox, D., ”Damping Coatings 

Research at NASA Glenn Research Center,” Propulsion—Safety 

and Affordable Readiness (P-SAR) Conference, San Diego, CA, 

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Diego, CA, Mar. 10–13 2008, Proc. SPIE, Vol. 6929, 69291C 

(2008).

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Resonances Using Piezoelectric Transducers,” Propulsion—

Safety and Affordable Readiness (P-SAR) Conference, Myrtle 

Beach, SC, Mar. 18–20, 2008.

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Piezoelectric Damping of Centrifugally Loaded Plates,”

Proceeding of the 17th AIAA Adaptive Structures Conference, 

Palm Springs, CA, May 4–7. 2009, AIAA–2009–2524.

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for Spin Vibration Testing,” Physics and Process Modeling 

(PPM) and Other Propulsion R&T, Volume 2; NASA CP–

10193, Apr. 1997.

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Suspension and Excitation System for Spin Vibration Testing of 

Turbomachinery Blades,” 39th AIAA Structures, Structural 

Dynamics and Materials Conference, Long Beach, CA, Apr. 20–

23, 1998.

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pp. 226–237.

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Switching Frequency Effects on Eddy Current Sensors for 

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International Energy Conversion Engineering Conference, 

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Speed Flywheel System,” 1st AIAA International Energy 

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Experimental Methodology for Determining Turbomachinery Blade Damping Using 

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14. ABSTRACT

Experiments to determine the effects of turbomachinery fan blade damping concepts such as passively shunted piezoelectric materials on 

blade response are ongoing at the NASA Glenn Research Center. A vertical rotor is suspended and excited with active magnetic bearings 

(AMBs) usually in a vacuum chamber to eliminate aerodynamic forces. Electromagnetic rotor excitation is superimposed onto rotor PDcontrolled support and can be fixed to either a stationary or rotating frame of reference. The rotor speed is controlled with an air turbine 

system. Blade vibrations are measured using optical probes as part of a Non-Contacting Stress Measurement System (NSMS). Damping is 

calculated from these measurements. It can be difficult to get accurate damping measurements using this experimental setup and some of the 

details of how to obtain quality results are seemingly nontrivial. The intent of this paper is to present those details. 

15. SUBJECT TERMS

Turbomachinery; Vibration; Piezoelectricity; Magnetics bearings; Excitation 

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