kpm crack detection part3

7/31/2019 KPM Crack Detection Part3

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BLADE AND SHAFT CRACK DETECTION USINGTORSIONAL VIBRATION MEASUREMENTS

PART 3: FIELD APPLICATION DEMONSTRATIONS

Ken Maynard 1 and Martin Trethewey2

1 Applied Research Laboratory, The Pennsylvania State University2 Dept. of Mechanical Engineering, The Pennsylvania State University

State College, PA 16801, USA

Abstract: The primary goal of the this paper is to summarize field demonstrations of the feasibility ofdetecting changes in blade and shaft natural frequencies (such as those associated with a blade or shaft

crack) on operating machinery using non-contact, non- intrusive measurement methods. Laboratorydemonstration of feasibility and some special signal processing issues were addressed in Parts 1 and 2

[1, 2]. Part 3 primarily addresses the results of application of this non-intrusive torsional vibrationsensing to: a large wind tunnel fan; a jet engine high-pressure disk; a hydro station turbine; and to alarge coal- fired power plant induced-draft (ID) fan motors. During the operation of rotating equipment,

torsional natural frequencies are excited by turbulence, friction, and other random forces. Laboratorytesting was conducted to affirm the potential of this method for diagnostics and prognostics of blade and

shafting systems. Field installation at the NASA Ames National Full-Scale Aerodynamic Facility(NFAC) reaffirmed the ability to detect both shaft and blade modes. Installation on a high-pressure(HP) disk in a jet engine test cell at General Electric Aircraft Engines demonstrated that the fundamental

mode of the turbine blades was clearly visible during operation. Field installation at a hydro powerstation demonstrated that the first few shaft natural frequencies were visible, and correlated well with

finite element results. Finally, field installation on the ID fan motors also showed the first few shafttorsional modes. These field tests have resulted in high confidence in the feasibility of the application ofthis technique for diagnosing and tracking shaft and blade cracks in operating machinery.

BACKGROUND

The detection of blade and shaft natural frequencies in the torsional domain requires that the signalresulting from excitation of the rotating elements by turbulence and other random processes is

measurable. If measurable, these natural frequencies may be tracked to determine any shifting due tocracking or other phenomena affecting torsional natural frequencies. Difficulties associated with

harvesting the potentially very small signals associated with blade and shaft vibration in the torsionaldomain could render detection infeasible. Thus, transduction and data acquisition must be optimized fordynamic range and signal to noise ratio [1, 4, 5]. An overview of the laboratory results may be found in

[1, 2].


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The advantage of using shaft torsional natural frequency tracking over shaft lateral natural frequencytracking for detecting cracks in direct-drive machine shafts is twofold:

A shift in natural frequency for a lateral mode may be caused by anything which changesthe boundary conditions between the rotating and stationary elements: seal rubs, changes in

bearing film stiffness due to small temperature changes, thermal growth, misalignment, etc.So, if a shaft experiences a shift in lateral natural frequency, it would be difficult to

pinpoint the cause as a cracked shaft. However, none of these boundary conditionsinfluence the torsional natural frequencies. So, one may say that a shift in naturalfrequency in a torsional mode of the shaft must involve changes in the rotating element

itself, such a crack, or perhaps a coupling degradation.

Similarly, finite element modeling of the rotor is simplified when analyzing for torsional

natural frequencies. The aforementioned boundary conditions, which are so difficult tocharacterize in rotor translational modes, are near non-existent in the torsional domain for

many rotor systems. This means that characterization of the torsional rotordyanamics ismore straightforward, and therefore likely to better facilitate diagnostics.

Detection of the small torsional vibration signals associated with blade and shaft natural

frequencies is complicated by transducer imperfections and by machine speed changes. The useof resampling methods has been shown to facilitate the detection of the sha ft natural frequencies

by: (1) correcting for torsional transduction difficulties [2, 4] resulting from harmonic tapeimperfections (printing error and overlap error); and (2) correcting errors as the machineundergoes gradual speed fluctuation [2, 5, 6]. In addition, correction for more dramatic speed

changes was addressed in [6]. These corrections made laboratory testing quite feasible.

Transducer setup and methodology

The transducer used to detect the torsional vibration of the shaft included a shaft encoded with

black and white stripes, an infrared fiber optic probe, an analog incremental demodulator or

digital demodulator and an A/D converter. The implementation of the technique underlaboratory conditions was previously presented in [4, 5]. Figure 1 shows a schematic of thetransducer system using analog demodulation.

Figure 1: Schematic of transducer setup for torsional vibration measurement

Figure 2 shows a spectrum of the torsional vibration of a desktop rotor, with the natural

frequency of the blade cluster and of the single detuned blade (rogue blade) evident [5, 6].

Equally spaced black and

white stripe code

Fiber optic cableFiber optic probe

Power Supply

9.12

Analog

demodulator

A-D converter

Shaft experiencing

torsional vibration


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Figure 2: Torsional Spectrum of Laboratory Rotor with One Detuned Blade

IMPLEMENTATION FOR BLADE CRACKING

NASA Ames National Full-Scale Aerodynamic Complex (NFAC) Fans

The NFAC facility is the largest wind tunnel in the world, able to test a full-scale 737 in thelargest section (see Figure 3). The fans used to drive the wind tunnel, shown in Figure 4, haveexperienced some blade cracking in the past, and the repairs made continued inspection

unpractical. NASA decided to use modal impact testing of the blades to track the natural

frequencies of the blades to detect shifts that might be associated with cracking. However, it wasdetermined that the frequency might shift as much as ten percent simply by rotating the rotor360o and retesting the same blade. Since this was on the order of the expected shift due tocracks, other methods were considered. A feasibility study was conducted using torsional

vibration on an operating fan.

0 5 0 1 0 0 1 5 0 2 0 0 2 5 0 3 0 0 3 5 0- 1 0 0

- 9 0

- 8 0

- 7 0

- 6 0

- 5 0

- 4 0

First Shaft

Torsional

Mode

Cluster of

BladeModes

Rogue

Blade


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Figure 3: National Full-Scale Aerodynamic Complex (NFAC) at NASA Ames

Figure 4: NFAC Fans

The shaft was encoded with zebra tape, as shown in Figure 5, and testing was performed during

operation of the fans.


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Figure 5: Zebra Tape Installation on the NFAC Fan Shaft

The results are shown in Figure 6. The first two peaks correspond to overall shaft torsional

frequencies, based on simple dynamic models developed by the vendor in the 70s. The thirdpeak, near 13.5 Hz, corresponds to the blade group. Of particular interest is the tight packing of

the fifteen blade modes, especially knowing that the impact testing on a single blade could varyby 1 Hz or more during a single test session. We believe that the variable pitch mechanism(VPM) does not provide a repeatable boundary support for the blades when the fan is not

operating, but provides uniform and repeatable results during operation, when the VPM ispreloaded.

Figure 6: Averaged Torsional Spectrum of NFAC Fan

Subsequent to this testing, impact testing was performed on a blade with the oil pump operatingto attempt to allow torsional coupling. The resulting natural frequency was found to be about

14.6 Hz, as shown in Figure 7.

This testing demonstrated the ability of this measurement system to detect blade natural

frequencies during operation of the NFAC fans.

0.000001

0.00001

0.0001

0 2 4 6 8 10 12 14 16

Frequency (Hz)

Volts

Shaft TorsionalNatural

FrequenciesBlade Natural

Frequencies


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Figure 7: Results of Modal Testing of Single Blade

Application to aging vehicular turbines

In vehicular turbines, the loss of turbine blades has resulted in accidents and fatalities. With theaging of commercial and military fleets, fatigue cracking and failure of discs and blades becomes

a more pressing issue. For example, some of the engines currently used in about four thousand737, A319, A320, A321, A340, and DC8-70 commercial aircraft are approaching twenty years of

age. Although these engines experienced little blade and disk cracking early in their lives,cracking has been detected during regularly scheduled inspections on aging engines. There is

some concern that this cracking may accelerate as the engines age, and may require increasedinspections. The possibility of a crack precipitating and resulting in blade or disk failure inbetween scheduled inspections increases with age, and an in situ system could provide the

user/maintainer with blade/disk health information.

Similarly, military aircraft engines are exhibiting symptoms of aging, which include blade anddisk failure. From 1982 to 1998, 55% of USAF engine caused mishaps were related to high-

cycle fatigue (HCF) of engine parts [11], predominately blades and disks at the blade attachment[12]. In addition, about 87% of the risk management inspections were related to this HCF.

Thus, a large portion of the maintenance budget is associated with HCF of blades and disks at theblade attachment.

Instrumentation of an aircraft jet turbine

The HP turbine of a commercial jet engine was instrumented to determine the feasibility ofdetecting the blade natural frequencies during operation using torsional vibration measurement.The testing was performed in a warm air test facility, under load at about 9400 RPM. Figure 8

shows the fully assembled commercial engine and the HP disk as installed in the warm air testfacility.

14.609

0 10 20 30 40 50 60

Frequency (Hz)

H(f)

100

10

1

0.1


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Figure 8: a) Fully Assembled Jet Engine; b) HP Disk in Warm Air Test Facility

Several individual blades were tested by the engine manufacturer to determine the natural

frequencies. These blades were fixtured to simulate the boundary conditions during operation.Figure 9 shows a typical spectrum of the results of the modal testing, which is summarized for

the three specimen blades in Table 1.

Figure 9: Typical Spectrum of Individual Fixtured Blade


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Table 1: Results of Modal Testing of Sample Blades

The instrumentation included a 200-stripe zebra tape and fiber optic probes. In addition, the

output from a 60-tooth speed encoder was used as backup. The zebra tape did not survive thetest, and the speed signal from the 60-tooth gear was used.

Figure 10: a) Zebra Tape Installation; b) 60-Tooth Speed Gear

The use of the 60-tooth wheel limitedthe frequency range of the data to one-half the number ofteeth times the speed of the shaft, or about 4500 Hz. Figure 11 shows the torsional spectrum of

the HP turbine rotor. Note that torsional natural frequencies of the shafting system shouldchange very little with speed, and coupled blade and shaft torsional modes increase very slightly

with speed due to stiffening by axial force. So, we look for frequencies that are not shifting withspeed. Note that the first peak at about 2400 Hz is close to the first mode from the blade modaltesting.

Earlier experimental and analytical work, however, indicates that blade natural frequencies can

be significantly altered by coupling with shaft torsional modes [5,6]. It was shown that for asmall desktop rotor with much larger blade to rotor mass ratio the difference might be more than

30%, and may be higher or lower than rig impact testing. Although intuition might imply thatfor relatively small blades coupling would be less of a factor, the effects of torsional couplingshould be clarified using a simple dynamic torsional model of the engine.

Mode 1 Mode 2 Mode 3 Mode 4

Blade F12 2450 3943 6038 8409

Blade C14 2413 3853 6091 8469

Blade B7 2434 3875 5994 8364

Average 2432 3890 6041 8414

Std. Dev. 19 47 49 53


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Figure 11: Torsional Spectrum of HP Turbine

IMPLEMENTATION FOR SHAFT CRACKING

Application to a Hydro Turbine:

The hydro plant consists of five 3 MW electric turbine generators sets. The plant was originallybuilt in about 1910, but it has recently been redesigned to eliminate an underwater, wooden

(lignum vitae) bearing and improve efficiency. The layout of a unit is shown in Figure 12 andFigure 13.

-60

-55

-50

-45

-40

-35

-30

-25

-20

-15

-10

2200 2400 2600 2800 3000 3200 3400 3600 3800 4000 4200

Frequency, Hz

Gain

9000 rpm, 60 tooth

9450 rpm, 60 tooth

Blade F12, Mode 1

Blade C14, Mode 1

Blade B7, Mode 1


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Figure 12: Hydro turbine-generator set

Figure 13: Disassembled Hydro-Turbine

In the last five years, three of the newly designed turbine rotors have experienced severecracking. Instrumentation and analysis was performed on one of the units that had notexperienced cracking to demonstrate the feasibility of detecting shaft natural frequencies. Figure

14 shows the optic probe, tachometer, and encoded tape placement.


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Figure 14: Optic probe, encoded tape, and tachometer placement

The data was analyzed using the double resampling technique [3,6] to eliminate the adverse

effects of the presence of running speed and its harmonics on frequency identification. Theresults of four test runs are shown in Figure 15. Note the peaks at about 16 Hz and 41 Hz. These

correspond well to the results from a simple finite element model (16 Hz and 40 Hz).

1.E-04

1.E-03

1.E-02

0 5 10 15 20 25 30 35 40 45 50

Frequency (Hz)

Amplitude

Run 4

Run 5

Run 7

Run 11

Figure 15: Torsional spectrum of hydro unit shaft motion

Fiber Optic

Probe

Infrared

Tachometer

Encoded tape


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The frequencies below 5 Hz are somewhat enigmatic. Since the operating speed of the unit is300 RPM, or 5 Hz, it was at first assumed that these frequencies correspond to fluid whirl, which

generally occurs at speeds between 0.42 and 0.48 times operating speed [10]. However, the shaftlateral vibration data exhibited none of the signs of whirl. In addition, the three closely spaced

subsynchronous peaks were stable and repeatable from run to run, as seen in Figure 16. Such

stability and repeatability for three closely spaced frequencies does not correspond to the whirlphenomenon. In addition, similar spectral components have since been observed on hydro units

at other sites. So, we hypothesize that these subsynchronous frequencies corresponds to therigid body torsional mode on torsional springs corresponding to the bearing film stiffness in

shear. Further investigation will be necessary to confirm this and to clarify the significance ofthese spectral components.

1.E-04

1.E-03

1.E-02

0 1 2 3 4 5

Frequency (Hz)

Amplitude

Run 1

Run 2

Run 4

Run 5

Run 7

Run 9

Run 10

Run 11

Figure 16: Subsynchronous torsional spectrum for hydro unit shaft

Several issues arose during the on-site data acquisition and analysis. Figure 17 shows some of

the data of Figure 1 along with runs that had significant distortion due to tape errors. When thetape was changed, or even the axial location of the transducer was changed on the same tape, the

spurious frequencies shifted. These spurious frequencies seem to be related to the encoded tape,and often interfered with the identification of shaft natural frequencies.


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1.E-04

1.E-03

1.E-02

1.E-01

0 5 10 15 20 25 30 35 40 45 50

Frequency (Hz)

Amplitude

Run 1

Run 2

Run 4

Run 5

Run 7

Run 9

Run 10

Run 11

Figure 17: Torsional spectra showing encoded tape error spectral content

Application to an Induced-Draft Fan Motor

The motors on the fossil- fired induced draft fan were constructed using rectangular cross sectionwebs from the shaft to the rotor coil supports. The square end on the web was then welded to the

circular shaft without machining to match the contours. The result has been a number of failuresof the motors due to failure of the web welds. Two of these motors were instrumented to detectshaft natural frequencies and establish a baseline to track the changes that may be associa ted

with web weld failure. Figure 18 shows the fan motor.

Figure 18: ID fan motor: (a) Motor housing; (b) scaled with minivan


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Installation of the tape was more difficult on the ID fan than on the hydro unit due to the shaft

size and the tight quarters. Figure 19 shows the installation of the transducer system.

Figure 19: Tape, fiber optic probe, and tachometer installation on ID fan

In addition, the butt joint misalignment of the ends of the tape appeared to be exacerbated by

thermal growth of the shaft. It was observed that a space between the ends appeared after heatup of the unit. This underlap, in some cases, caused saturation and malfunction of the analog

demodulator. Figure 20 shows the results for one of the fans. The first mode appears to be about10 Hz. Once again, it was observed that changing tapes or changing the shaft axial position ofthe optical probe on the encode tape changed some of the spectral content above 20 Hz. It is

difficult to assess the remainder of the spectrum with high confidence due to the spectral contentof the tape. However, most likely the second and third modes are at about 16 Hz and 19 Hz.

Fiber Optic

Probe

Infrared

Tachometer

Encoded tape


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1.E-04

1.E-03

1.E-02

0 5 10 15 20 25 30 35 40

Frequency (Hz)

Amplitude

Run 4

Run 5

Figure 20: ID Fan motor torsional spectrum

SUMMARY AND CONCLUSIONS

The techniques developed for detecting torsional natural frequencies in the laboratory were

implemented on a large wind tunnel fan which has experienced blade cracking, a jet enginewhich has experience cracking in the disc near at blade root, a hydro-turbine shaft which has

experienced cracking, and a power plant induced draft fan motor shaft which has experiencedcracking. The goals of the implementation project were to demonstrate the feasibility of fieldapplication, and to establish a baseline for each class of machine. The data acquired clearly

demonstrated the feasibility of field implementation, and established baseline natural frequenciesfor specific blades and shafts.

However, interference from tape related spectral content was experienced. This interference wasnot experienced in the laboratory due to shaft size, access, and environmental differences. It is

believed that this spectral content is associated not with tape printing error or overlap, but wasintroduced by the installation. This may be overcome by using a more precisely fabricated

encoding device (such as the 60-tooth speed encoder at the jet engine test facility), or bydeveloping correction algorithms for the installation error.

Future work: Research to establish the size of detectable cracks in each class of machine isneeded. For instance, reactor coolant pumps (RCP) in nuclear power plants have experienced

severe cracking with no advance notice given by the crack detection instrumentation. Similarly,although it is known that blade cracking in a jet engine would affect blade natural frequency, it is


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unclear what effects are measurable in the torsional domain. In order to establish the torsionalvibration method as practical in field machines, it is important that we be able to estimate the

size of the crack at the detection threshold and the remaining useful life of the shaft/blade at thatdetection threshold. In addition, correction of the installation errors must be accomplished to

remove ambiguity and make the technology widely accessible. This work is currently underway.

Acknowledgement : This work described herein was supported by the Southern Company

through the Cooperative Research Agreement Torsional Vibration and Shaft Twist Measurementin Rotating Machinery (SCS Contract Number C-98-001172); NASA Ames Research Center

Grant Condition-Based Monitoring of Large-Scale Facilities, (Grant NAG 21182); byMultidisciplinary University Research Initiative for Integrated Predictive Diagnostics (GrantNumber N00014-95-1-0461) sponsored by the Office of Naval Research; and by NASA

Glenn/GE Aircraft Engines Turbine Disk Crack Detection Using Torsional Vibration: AFeasibility Study (GEAE Purchase Order 200-1X-14H45006). The content of the information

does not necessarily reflect the position or policy of the Government, and no officialendorsement should be inferred.

References:

1. Maynard, K. P., and Trethewey, M. W., Blade and Shaft Crack detection Using TorsionalVibration Measurements Part 1: Feasibility Studies, Noise and Vibration worldwide, Volume

31, No. 11, December 2000, pp. 9-15.

2. Maynard, K. P., and Trethewey, M. W., Blade and Shaft Crack detection Using TorsionalVibration Measurements Part 2: Resampling to Improve Effective Dynamic Range, Noise

and Vibration worldwide, Volume 32, No. 2, February 2001, pp. 23-26.

3. Vance, John M.,Rotordynamics of Turbomachinery, John Wiley & Sons, New York, 1988,

pp. 377ff.

4. Maynard, K. P., and Trethewey, M., On The Feasibility of Blade Crack Detection ThroughTorsional Vibration Measurements, Proceedings of the 53rd Meeting of the Society for

Machinery Failure Prevention Technology, Virginia Beach, Virginia, April 19-22, 1999, pp.451-459.

5. Maynard, K. P.; Lebold, M.; Groover, C.; Trethewey, M., Application of Double Resamplingto Shaft Torsional Vibration Measurement for the Detection of Blade Natural Frequencies,Proceedings of the 54th Meeting of the Society for Machinery Failure Prevention

Technology, Virginia Beach, VA, pp. 87-94.

6. Groover, Charles Leonard, Signal Component Removal Applied to the Order Content in

Rotating Machinery, Master of Science in Mechanical Engineering Thesis, Penn StateUniversity, August 2000.

7. McDonald, D, and Gribler, M., Digital Resampling: A Viable Alternative for Order Domain

Measurements of Rotating Machinery, Proceedings of the 9th Annual International ModalAnalysis Conference, Part 2, April 15-18, 1991, Florence, Italy, pp. 1270-1275.

8. Potter, R., A New Order Tracking Method for Rotating Machinery, Sound and Vibration,September 1990, pp. 30-35.


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9. Hernandez, W., Paul, D., and Vosburgh, F, On-Line Measurement and Tracking of TurbineTorsional Vibration Resonances using a New Encoder Based Rotational Vibration Method

(RVM), SAE Technical Paper 961306, Presented at the Aerospace Atlantic Conference,Dayton, OH, May 22-23, 1996.

10. Sawyer, John W., Ed., Sawyers Turbomachinery Maintenance Handbook, 1st Ed., Vol. II,

Turbomachinery International Publications, 1980, p. 7-34ff.

11. Davenport, Otha B., Operational Readiness and High Cycle Fatigue, Turbine Engine

Symposium, Dayton Convention Center, 14-17 September 1998.

12. Davenport, Otha B., Senior Executive Service, director of engineering, Propulsion Product

Group, Aeronautical Systems Center, Air Force Materiel Command, Wright-Patterson AirForce Base, personal conversation, 29 June 2001.

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