Performance of miniature shakers for vehicle component testing
Bart Peeters, Peter van der Linden, Christophe De Veuster LMS International Interleuvenlaan 68 B-3001 Leuven BELGIUM Tel: +32 16 38 42 00 Fax: +32 16 38 43 50 E-mail:
[email protected]
ABSTRACT Extremely compact electrodynamic shakers (diameter 40.5 mm, weight 0.15 kg) have been developed to allow dynamic excitation, in the 30 to 5000 Hz frequency range, at difficult accessible locations. Within this small space, these miniature shakers have integrated force and acceleration transducers. Engine/power train internal excitation was the initial target application for such shakers (engine combustion chambers, crankshaft bearings, etc.). But, this paper investigates the performance of such shakers for in-situ vehicle component testing. The tests are challenging because of the poor access to the component and the analyses are challenging because of the occurrence of strongly coupled and highly damped modes. Two cases are discussed: an in-vehicle rim deformation test and an in-vehicle dash-steering carrier test. The excitation by miniature shakers is compared with traditional excitation using an instrumented hammer or classical electromagnetic shakers. Relevant items for this comparison are the test set-up efficiency, data quality, and analysis accuracy. 1 INTRODUCTION The mechanical product industry relies increasingly on CAE-based virtual prototyping to optimise the functional performances of its designs, moving away from the test-analyse-fix approach on physical prototypes. But contrary to the belief that this would reduce the demands for testing, it has opened new application fields, resulting in new challenges and opportunities. Test data play a critical role on each level of the development process, in product benchmarking, target setting, model verification, load analysis, hybrid model building to product qualification and performance monitoring. What is clear though is that the requirements towards accuracy, test and analysis ease and execution speed are more stringent than ever before [1]. This paper verifies the performance of miniature shakers related to test efficiency and data quality. The E-MISHA electro-dynamic miniature shaker was developed by LMS Engineering Services to do highly accurate structural excitation in a minimum of space [2]. Their diameter is 40.5 mm and the overall weight 0.15 kg (Figure 1). The shaker mass-loading of the test object is kept very low at 10 gram, but the stiffness-loading is higher than most stinger-shaker combinations. The shakers have integrated force and acceleration transducers. The maximum force RMS value is 2.5 N in an effective frequency range from 30 to 5000 Hz. The
Figure 1: Miniature shaker, diameter 40.5 mm.
shakers are directly glued to the test structure, so there is no need for external shaker supports. Their were initially developed to provide dynamic excitation at difficult accessible (and small) locations such as engine combustion chambers and crankshaft bearings (Figure 2). The typical application is vibroacoustic transfer path analysis [3]. In a hybrid engineering approach, the miniature shakers are used to identify an experimental model of the engine. The test-based engine model is then combined with finite element (FE) models of the engine brackets and the subframe [4][5]. Miniature shakers are also sometimes used to excite car and truck bodies for experimental statistical Figure 2: Miniature shaker exciting an engine [5]. energy analysis (SEA). SEA is particularly suited to model high frequency noise and vibration problems where the modal overlap is high, and the modes are mostly localized. Experimental SEA modelling procedures derive internal and coupling loss factor information from a set of experimental data, combining input power measurements with response measurements on the individual SEA subsystems. Finally, miniature shakers are also used as excitation devices in experimental modal analysis (EMA) applications. Typical structures are engines, transmissions, small machinery and vehicle components in general. The performance of E-MISHA miniature shakers for in-situ vehicle component testing is precisely the topic of this paper. Two cases are discussed: • In-vehicle rim deformation test: a comparison is made between excitation by miniature shakers and by traditional shakers. • In-vehicle dash-steering carrier test: a comparison is made between excitation by miniature shakers and by an instrumented hammer. 2
IN-VEHICLE COMPONENT TESTING AND MODAL ANALYSIS
2.1 Component testing considerations Several reasons exist to perform vehicle component tests: • Acceptance tests for component stiffness and deformation modes (dashboard, rim, …); • Acceptance tests for component attachment (intake filter-box, steering-pump, …); • Correlation tests for FE model verification (axles, engine brackets, oil-sumps, …); • Product diagnosis and improvement tests (transmission flange, dashboard assembly, …). Depending on the type of component and the application, one opt for specific boundary conditions during the vibration test: free-free suspension (subframe, steering wheel, …); clamped on test rig (engine brackets, mirror, …); in-situ vehicle (dash-carrier, rim, tires, axles, …). This last type of test is discussed in this paper. They are challenging because of the poor access to the excitation and response locations. Also the data analysis is challenging because of the occurrence of strongly coupled and highly damped modes. It is expected that the use of miniature shakers have some advantages for this type of tests as compared to traditional shakers or hammers: • Fast instrumentation. The miniature shakers are simply glued to the structure without the need for shaker support frames. The shaker integrates a stinger and a force and acceleration sensor, which are selfaligning; so there is no need to carefully align the stinger and position external sensors. • Flexible excitation location choice. Due to the miniaturisation, they require much less space than a traditional shaker. If compared to a hammer, they need less visibility and no swing space. • Accurate measurements. Typical sensor-force alignment errors are prevented because the miniature shakers are self-aligning. Moreover, thanks to the flexibility, an optimal choice for the excitation location can be made increasing the data quality.
Figure 3: Rim deformation test. (Left) excitation by two miniature shakers; (Right) excitation by traditional shakers.
Disadvantages of using miniature shakers are: • The relatively low excitation force. Tests in quiet environments are well possible, even on large and heavy objects like engines and car-bodies. But tests in locations with a high background noise and background vibrations can be difficult. • The non-axial stiffness of the miniature shakers is higher then many normal long-stinger-shaker combinations. They should be connected to stiff locations of the test object. Good results are obtained at nodes of reinforcements and locations where mass is concentrated on the test objects. 2.2 Rim deformation test The purpose of this test was to identify the rigid body, flange deformation and spoke deformation modes of a rim of a car. The accelerations were measured in 3 directions at 25 locations – i.e. 75 degrees of freedom (DOFs) – while 2 shakers were simultaneously exciting the structure. The measurements were performed using LMS Scadas III hardware [6] and LMS Test.Lab Structural Testing software [7]. A first test was carried out using two miniature shakers, a second with traditional shakers (Figure 3). For practical reasons, it was impossible to use the same excitation points in both tests. Figure 4 compares the excitation power spectral densities (PSDs) and the multiple coherences of some typical DOFs. In Figure 5 a reciprocity check is performed: the frequency response
F F B B B B
/
1.00
PSD force 1 PSD force 2 Coherence WL:1:+Y/Multiple Coherence WL:2:+Y/Multiple Coherence WL:3:+Y/Multiple Coherence WL:4:+Y/Multiple
10.0e-6
0.00
20.00
10.0
Amplitude
/
( N2/Hz) Log
Amplitude
1.00
( N2/Hz) Log
10.0
Hz
1024.00
10.0e-6
0.00
20.00
Hz
Figure 4: Force PSDs and typical multiple coherences. (Left) miniature shakers; (Right) traditional shakers.
1024.00
200e-3
( g/N) Log
( g/N) Log
200e-3
FRF acc 2/force 1 FRF acc 1/force 2
FRF acc 2/force 1 FRF acc 1/force 2 100e-6
180.00
180.00 ° Phase
° Phase
100e-6
-180.00
-180.00
20.00
Hz
1024.00
20.00
Hz
1024.00
Figure 5: Theoretically reciprocal FRFs. (Left) miniature shakers; (Right) traditional shakers.
function (FRF) between the force at the 1st shaker and the acceleration at the 2nd shaker is compared to the force at the 2nd shaker and the acceleration at the 1st shaker. Theoretically both FRFs should be the same. Although the reciprocity in case of the traditional shakers is still acceptable (Figure 5, Right), it is clearly better in case of the miniature shakers (Figure 5, Left). Misalignment of stinger – accelerometer – force sensor at the driving point locations is responsible for non-reciprocal behaviour. Non-reciprocal data will affect the quality of the modal model that is derived from the FRFs. It is not shown here, but it was found that the quality of the driving point FRFs was good in both cases. This quality is verified by checking that resonances are alternated by anti-resonances and that the FRF phases are between 0 en 180° [8]. Again misalignment can be the cause of low driving-point quality. As there was some concern of impedance loading of the structure due to the miniature shakers, a test was performed with the 2 traditional shakers in 2 situations: no miniature shakers were present versus one of the miniature shakers was (passively) present. The same FRF for both situations is compared in Figure 6. Although a small effect can be observed, the mass – damping – stiffness loading of the miniature shaker is negligible if highimpedance locations are selected for attaching them.
( g/N) Log
50.0e-3
1.00e-3 20.00
Hz
1024.00
Figure 6: Verification of the impedance loading due to the miniature shaker. (Red/black) no miniature shaker was present; (Grey/green) one of the miniature shakers was passively present during the traditional shaker test.
F F B
1.00
FRF WL:5:+Y/force 1 FRF WL:5:+Y/force 2 Coherence WL:5:+Y/Multiple
500e-6
F F B
0.90
20.00
Hz
FRF WL:5:+Y/force1 FRF WL:5:+Y/force2 Coherence WL:5:+Y/Multiple
5.00e-3
1024.00
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( (m/s2)/N) Log
/
( g/N) Log
Amplitude
1.00
Amplitude
100e-3
0.90
20.00
Hz
1024.00
Figure 7: FRFs and multiple coherences. (Left) Burst random excitation. (Right) Periodic chirp excitation.
In combination with miniature shakers, it is interesting to use special excitation signals such as periodic chirp (a fast sine sweep over the frequency range of interest within 1 acquisition block) or a multisine with minimised crest factor [9]. A multisine consists of multiple sine waves of which the frequencies correspond to the spectral lines. The crest factor is the ratio of peak amplitude to RMS value. If compared to classical (burst) random signals, the same shaker can inject more energy in the system when using these special signals, precisely because of their low crest factor. Figure 7 compares burst random with periodic chirp excitation. The multiple coherence value of the periodic chirp data reveals a slight improvement in data quality (Please note that the coherence axis extends between 0.9 and 1). After having verified the data quality, the modal parameter estimation can be performed. Hereto, the new “PolyMax” method is used, which is the Test.Lab [7] implementation of the new fast-stabilising polyreference least-squares complex frequency-domain parameter estimation method [10][11]. The new method outperforms other commercially available methods when it comes to difficult cases involving strongly coupled and highly damped modes and when a lot of noise is present. The method and some industrial applications are discussed in another paper of these proceedings [11]. A very attractive feature of PolyMax is that it yields very clear stabilisation diagrams (see Figure 8 in which a comparison is made with the LSCE stabilisation diagram). Many modes are tire modes with basically a rigid body motion of the rim. An average correlation of 97% is achieved between the measured FRFs and the FRFs synthesized based on the identified modal parameters (the definition of FRF correlation can be found in [11]). Two typical mode shapes are shown in Figure 9. There was no difference in difficulty of the modal analysis process or in the accuracy of the results between miniature and traditional shaker data. Table 1 gives a detailed timing of the rim deformation test and data analysis. The fast attachment-alignment and the accuracy of the data accelerated the process instrumentation from start to analysed results by approximately 20 %, compared to classical shakers. There is a clear efficiency advantage when using the miniature shakers: only 81% of the time is required of the test with classical shakers.
Table 1: Timing of the rim deformation test and data analysis. Except for the last column, all values have hours [h] as units. Timing [h]
Comp.
System
Meas.
Shaker
Trials
Meas.
Data
Post-
Total
Relative
Prep.
set-up
locations
attach.
and
runs
verif.
proc.
time
time [-]
and
and
and
checks
and
and
testing
wireframe
alignment
Traditional shaker
0
1
2.5
3
1
2
1
3
13.5
100
E-MISHA
0
1
2.5
1
0.5
2
1
3
11
81
storing
analysis
Figure 8: Stabilisation diagram obtained by applying the traditional time-domain LSCE method (Top) [7][8] and the new frequency-domain PolyMax method (Bottom) [10][11] to rim FRFs in a band between 260 and 620 Hz. The background functions are the complex mode indicator functions.
Figure 9: Two typical rim deformation mode shapes. (Left) mode at 329 Hz. (Right) mode at 963 Hz.
2.3 Dash-steering carrier test The purpose of this test was to study the steering wheel vibrations in view of the driver’s comfort by identifying the resonance frequencies of the dash-steering carrier inside a car. The accelerations were measured in 3 directions at 15 locations – i.e. 45 degrees of freedom (DOFs). A first test was carried out using two E-MISHA miniature shakers, a second with an instrumented hammer applied at 2 different locations. One excitation point was on the steering column (Figure 10), the other was on the lateral beam. The access to the excitation and measurement locations was rather difficult. Because of this difficult access the hammer and shaker excitation location are not identical. Figure 11 compares the excitation power spectral densities (PSDs) and the multiple/ordinary coherences of some typical DOFs. The hammer force spectra look good, but the coherences are rather moderate.
Figure 10: Dash-steering carrier test. (Left) excitation by a miniature shakers; (Right) excitation by a hammer.
F F B B
F F B B
0.00
Hz
PSD Shak:1:+Z PSD Shak:2:+Z Coherence Beam:2:+Y/Shak:2:+Z Coherence Beam:10:+Y/Shak:2:+Z
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200.00
/
1.00
PSD force 1 PSD force 2 Coherence Beam:2:+Y/Multiple Coherence Beam:10:+Y/Multiple
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20.00
2.00e-3
Amplitude
/
Real
( N2/Hz) Log
1.00
( N2/Hz) Log
2.00
0.01
20.00
Hz
200.00
Figure 11: (Left) miniature shakers force PSDs and multiple coherences (20 averages); (Right) hammer impact PSDs and ordinary coherences (4 averages).
The direct FRFs are shown in Figure 12. The direct FRFs at the lateral beam show much less “dynamics” than the direct FRFs at the steering column. In Figure 13 a reciprocity check is performed: the frequency response function (FRF) between the force at the 1st excitation location and the acceleration at the 2nd excitation location is compared to the force at the 2nd excitation location and the acceleration at the 1st excitation location. Both FRFs should coincide. Although the reciprocity in case of the hammer is still acceptable (Figure 13, Right), it is clearly better in case of the miniature shakers (Figure 13, Left). The E-MISHA shakers have an excellent alignment of stinger – accelerometer – force sensor at the driving point locations, whereas impact testing with a hammer suffers from position and direction uncertainty, especially at difficult accessible locations. As in previous example (Section 2.2) the impedance loading of the structure due to the miniature shakers was verified. Again it can be concluded that the mass – damping – stiffness loading of the miniature shaker is negligible since they were attached to the structure at high-impedance locations.
100e-3
( g/N) Log
( g/N) Log
1.00
FRF Shak:1:+Z/Shak:1:+Z FRF Shak:2:+Z/Shak:2:+Z
FRF acc 1/force 1 FRF acc 2/force 2 1.00e-3
180.00
180.00
° Phase
° Phase
1.00e-3
-180.00
-180.00
20.00
Hz
200.00
20.00
Hz
Figure 12: Direct FRFs: force and acceleration measured at the same DOF. (Left) miniature shakers; (Right) hammer
200.00
20.0e-3
( g/N) Log
( g/N) Log
200e-3
FRF Shak:2:+Z/Shak:1:+Z FRF Shak:1:+Z/Shak:2:+Z
FRF acc 2/force 1 FRF acc 1/force 2 100e-6
180.00
180.00 ° Phase
° Phase
100e-6
-180.00
-180.00
20.00
Hz
200.00
20.00
Hz
200.00
Figure 13: Theoretically reciprocal FRFs. (Left) miniature shakers; (Right) hammer.
Figure 14: Stabilisation diagram obtained by applying the new frequency-domain parameter estimation method [10][11] to the dash-steering carrier FRFs in a band between 20 and 200 Hz.
Table 2: Eigenfrequencies and damping ratios obtained by applying the Polyreference LSCF method to E-MISHA data. Poles
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
Freq. [Hz]
32.5
55.4
67.4
69.7
101.6
102.4
115
121
123
131
144
154
164
171
174
Damp. [%]
3.9
3.6
2.7
3.9
4.9
1.1
3.2
2.4
4.3
4.2
4.5
3.4
1.7
1.0
2.8
After having verified the data quality, the modal parameter estimation can be performed. Again, the new PolyMax method was used [7][10][11]. A total of 15 modes could be extracted from the stabilisation diagram shown in Figure 14. The eigenfrequencies and damping ratios are listed in Table 2. Some typical measured and synthesised FRFs are shown in Figure 15. Considering that it concerns an in-situ vehicle test, the correspondence is excellent. Finally, Table 3 gives a detailed timing of the dash-steering carrier test and data analysis. Again the miniature shaker measurements reduced the entire test-analysis process by approximately 20%. Even with a very fast start of the instrumented hammer measurements, the required repeated hammering and the inaccuracy of the data take a lot of time, especially when access is limited. 3 CONCLUSIONS New technology has made it possible to make modal and general FRF testing more efficient and more accurate. Miniaturisation of the shaker-stinger system, the electro-magnetic drive and sensors has led to an integrated miniature shaker: LMS-MISHA. This innovative device will excite, small and large, stiff components and structures between 30 and 5000 Hz. It is glued onto test objects in seconds, self-aligning, and therefore providing good repeatability, reciprocity and direct FRFs. Compared to normal shaker-stinger excitation and instrumented hammer excitation, clear efficiency and accuracy advantages are possible. In combination with miniature shakers, it can be interesting to use excitation signals with a small crest factor, such as periodic chirp or an optimised multisine. Finally, the PolyMax method, a new frequency-domain parameter estimation method, dramatically facilitates the modal analysis process in “difficult” cases such as in-situ vehicle component testing.
2.00
( (m/s2)/N) Log
( (m/s2)/N) Log
2.00
FRF Beam:8:+X/forc:1:+Z FRF Beam:8:+X/forc:1:+Z FRF Beam:8:+X/forc:2:+Z FRF Beam:8:+X/forc:2:+Z
FRF acc:2:+Z/forc:1:+Z FRF acc:2:+Z/forc:1:+Z FRF acc:2:+Z/forc:2:+Z FRF acc:2:+Z/forc:2:+Z
200e-6
200e-6
180.00
° Phase
° Phase
180.00
-180.00
-180.00
20.00
Hz
200.00
20.00
Hz
200.00
Figure 15: Comparison of measured and synthesised FRFs. (Left) steering column DOF; (Right) lateral beam DOF.
Table 3: Timing of the dash-steering carrier test and analysis. Except for the last column, all values have hours [h] as units. Timing [h]
Comp.
System
Meas.
Shaker
Trials
Meas.
Data
Post-
Total
Relative
Prep.
set-up
locations
attach.
and
runs
verif.
proc.
time
time [-]
and
and
and
checks
and
and
testing
wireframe
alignment
Instrum. hammer
0
1
2.5
0
1.5
4
2
3
14
100
E-MISHA
0
1
2.5
1
0.5
2
1
3
11
79
storing
analysis
ACKNOWLEDGEMENTS This work was carried out in the frame of the EC-GROWTH research project GRD1-2001-40034 “AMPA” (Automatic Measurements Plausibility and Quality Assurance). The support of the EC is gratefully acknowledged. REFERENCES [1] VAN DER AUWERAER H. Requirements and opportunities for structural testing in view of hybrid and virtual modelling. In Proceedings of ISMA 2002, the International Conference on Noise and Vibration Engineering, Leuven, Belgium, September 2002. [2] LMS INTERNATIONAL. The E-MISHA, a Miniature Shaker with Integrated Sensors, Leuven, Belgium, www.lmsintl.com, 2003. [3] VAN HERBRUGGEN J., P. VAN DER LINDEN, H.-J. KNITTEL AND J. SCHNUR. Engine internal dynamic force identification and the combination with engine structural and vibro-acoustic transfer information. In Proceedings of the SAE Noise and Vibration Conference, SAE paper 2001-01-1596, Traverse City (MI), USA, 30 April – 3 May 2001. [4] SAKAI T., M. TERADA, S. ONO, N. KAMIMURA, L. GIELEN AND P. MAS. Development procedure for interior noise performance by virtual vehicle refinement, combining experimental and numerical component models. In Proceedings of the SAE Noise and Vibration Conference, SAE paper 2001-01-1538, Traverse City (MI), USA, 30 April – 3 May 2001. [5] SAKAI T. AND A. SAKAMOTO. Improvement of engine noise for the 2003 Accord using hybrid CAE technology. In Proceedings of the SAE Noise and Vibration Conference, SAE paper 2003-01-1427, Traverse City (MI), USA, 5–8 May 2003. [6] LMS INTERNATIONAL. LMS Scadas III – The Scalable Frontend, Leuven, Belgium, www.lmsintl.com, 2003. [7] LMS INTERNATIONAL. LMS Test.Lab – Structural Testing Rev 4B, Leuven, Belgium, www.lmsintl.com, 2003. [8] HEYLEN W., S. LAMMENS AND P. SAS. Modal Analysis Theory and Testing. Department of Mechanical Engineering, Katholieke Universiteit Leuven, Leuven, Belgium, 1995. [9] GUILLAUME P., P. VERBOVEN, S. VANLANDUIT AND E. PARLOO. Multisine excitations – new developments and applications in modal analysis. In Proceedings of IMAC 19, the International Modal Analysis Conference, Kissimmee (FL), USA, February 2001. [10] GUILLAUME P., P. VERBOVEN, S. VANLANDUIT, H. VAN DER AUWERAER AND B. PEETERS. A poly-reference implementation of the least-squares complex frequency-domain estimator. In Proceedings of IMAC 21, the International Modal Analysis Conference, Kissimmee (FL), USA, February 2003. [11] PEETERS B., P. GUILLAUME, H. VAN DER AUWERAER, B. CAUBERGHE, P. VERBOVEN AND J. LEURIDAN. Automotive and aerospace applications of a new fast-stabilising polyreference frequency-domain parameter estimation method. In Proceedings of IMAC 22, Dearborn (MI), USA, January 2004.
