Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011 Leuven, Belgium, 4-6 July 2011 G. De Roeck, G. Degrande, G. Lombaert, G. M¨uller (eds.) ISBN 978-90-760-1931-4
3214
Experimental investigation on squeal noise in tramway sharp curves R. Corradi1, P. Crosio1 S. Manzoni1, G. Squicciarini1 Politecnico di Milano, Department of Mechanical Engineering, Via La Masa 1, 20133, Milan Italy e-mail:
[email protected],
[email protected],
[email protected], 1
ABSTRACT: The aim of the present article is to bring comprehensive experimental data useful for better understanding the mechanism of squeal noise in tramway sharp curves. Simultaneous measurements of rail vibration, wheel vibration and radiated noise were performed during the transit of a tramcar. The presented results refer to two different test configurations: the instrumented wheel was positioned at first on the inner side of the curve and then on the outer one, so that the influence of wheel-rail contact on the excitation of the wheel vibration can be investigated. During the tests, the influence of the wheel-rail friction coefficient was investigated too. KEY WORDS: squeal noise measurements, wheel vibrations, tramcar. 1
INTRODUCTION
Squeal noise has been a deeply studied problem for many years and extensive literature is now available. This kind of noise often occurs in brake discs ([1][2][3][4]) and railway wheels. Such a noise is characterised by one or few harmonics and this is the reason of its name. The loud sound radiation is caused by disc or wheel vibration excited by brake-disk or wheel-rail contact forces. Many papers have explained the basic principles of squeal for brake discs but less literature is available about railway wheels. Nevertheless, important results have already been reached by papers specifically dealing with wheel squeal noise ([5][6][7][8]). Squeal usually arises when a train negotiates a curve, especially a short radius one. Two main mechanisms may generate such a phenomenon: • squeal originates as the wheel velocity is made up of a rail tangential component and another one normal to it (named lateral creep velocity). The lateral creepage of the wheel causes a stick/slip phenomenon and consequent forces, normal to the wheel plane, are induced, which can excite the wheel bending modes, causing noise radiation ([5][6][7][7]); • squeal can also be originated due to contact on the wheel flange [5]. It is generally agreed that, in any case, squeal noise derives from the creep forces occurring in sharp curves which are due to frictional instability [9]. As already mentioned, squeal noise is mainly made of a single frequency component, thus perceived as a pure tone which is considered one of the most annoying kind of noise [10]. The authors have experienced that a loud squeal emission can take place also for speeds lower than 15 km/h [11] (in agreement with reference [12]). This causes a great annoyance to people living near the tracks and it is of great impact when city-trams are considered, as they run very close to buildings. This why researchers, vehicle manufacturers and operators are looking for solutions capable of reducing squeal noise. As
an example resilient wheels have been used on city trams instead of the traditional mono-block wheels. Nevertheless, this solution has not been able to completely solve the problem ([13][14][15][16]), like many others as well [13]. The only solutions which have demonstrated to be reliable are the adoption of lubricating systems [17] or the active vibration control on wheels ([14][18][19]) but these solutions are currently expensive. Therefore this paper is aimed at providing a contribution to the knowledge of the squeal noise phenomenon. The authors carried out an extensive experimental activity with city-trams. A curve where significant sound radiation usually takes place was identified, then the kind of tram providing the louder squeal noise was selected for the tests. Once the curve and the tram were identified, various tests with the instrumented tram were performed on the same curve. The next section describes the system (i.e. tram car, wheel, rail and so on) on which tests have been performed. Furthermore, details are given about the measurement set-up adopted for experiments with the running tram. Then Section 3 is intended to face the test results and to give the main conclusions on the considered squeal phenomenon. 2
MEASUREMENT SET-UP
The curve chosen for tests has a radius of 17.5 m and is equipped with grooved rails and ballasted tracks. The tested city tram is a modern articulated vehicle, with seven carbodies and resilient wheels. The experimental tests performed can be split into three main activities: • modal analysis of the wheel mounted on its bearings. Actually two different modal analyses have been carried out: one with the suspended wheel and the other with the wheel laid down on the rail (i.e. as in operating conditions); • line tests with the instrumented wheel mounted on both the inner and outer side of the leading axle of a fourwheel bogie;
3215
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
line tests with reduced friction coefficient.
• 2.1
Wheel modal analysis
The wheel modal analysis has been carried out twice: once with the wheel suspended and then with the wheel laid down on the rail. In both cases the wheel was mounted on the bogie frame through its own bearings (Figure 1).
pure radial, or axial, mode does not exist but each type of motion always occur together with the other one. For the first three modes the ratio between radial and axial vibration amplitude is respectively equal to 0.16, 0.63 and 2.15. Table 1. Identified vibration modes. Modes are described as out-of-plane or in-plane depending on the direction of the highest vibration amplitude.
Figure 1. Experimental Modal Analysis set-up.
2
(m/s /N) Coherence
Excitation has been given by means of a dynamometric impact hammer while acceleration responses have been collected in 53 points, both in axial and in radial direction, on the web and on the rim. The frequency resolution used for the modal analysis was 2 Hz. In Figure 2 a non-collocated frequency response function is depicted for an accelerometer mounted on the rim. The input is given on the rim too. 1 0.5 0 1.5
Damping ratio (%)
536
0.78
1273
0.75
1423
1.36
2229
0.49
2478
0.38
Description Out-of-plane vibration with two nodal diameters Out-of-plane vibration with three nodal diameters In-plane vibration with three nodal diameters Out-of-plane vibration with four nodal diameters In-plane vibration with four nodal diameters
In Figure 3 the mode shape of the third mode is shown (1423 Hz), it will prove to be the most important mode in the sound radiation. The dashed circle stands for the separation between the web and the rim, while the black dots represent the measuring points. Although this mode has its maximum amplitude in the radial direction, the figure represents the outof-plane motion.
5
1 0.5
Natural Frequency (Hz)
1
2
4 3
(deg)
0 180 90 0 -90 -180 0
500
1000 1500 2000 frequency (Hz)
2500
3000
Figure 2. Inertance measured on the wheel rim for an input on the rim. Table 1 summarises the modes mainly involved in the squeal noise phenomenon in the laid down configuration. The modal parameters of the modes identified in the suspended configuration are similar to the ones presented in the table, the natural frequencies difference ranges within few Hertz. It can be noticed that, due to the characteristics of the resilient wheel, the in-plane (radial) vibration is always coupled with the out-of-plane (axial) one and vice versa, this means that a
Figure 3. Natural mode at 1423 Hz, picture of the out-of-plane motion. 2.2
Line tests
The wheel tested at the previous step was then instrumented and mounted on the tram on the front axle of the second bogie. Two different test sessions were performed: the first with the wheel on right side and the second on the left. Since all tests were performed on the same curve (to the left), this allowed to make tests with the instrumented wheel rolling onto both the inner and the outer rail.
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
The transducers mounted on the wheel were six accelerometers (named WA) and their position and sensing directions are given in Figure 4 and in Table 2. Four accelerometers were located on the rim covering a quarter of the wheel and measuring in the axial direction, one was mounted on the web (axial direction) and finally one was measuring the radial acceleration of the rim. A radio telemetry system was used to bring signals from the rotating wheel to the acquisition system on board. In the picture showed in Figure 5 it is possible to distinguish the telemetry pick up: this component of the acquisition system receives the acceleration data from the antenna hidden under the tape covering the hub. Between the accelerometers and the antenna several signal conditioning modules are needed to firstly treat the signals, they are located in a cavity between the web and the hub as visible in Figure 6. WA1
WA2
WA3
WA6
WA4
WA5
Figure 4. Position and sensing direction of the wheel accelerometers. Table 2. Position and sensing directions of the wheel accelerometers Sensor name
Location
WA1 WA2 WA3 WA4 WA5 WA6
rim rim rim rim rim web
Measuring direction axial axial radial axial axial axial
Figure 5. Sensors and telemetry pick-up on the wheel.
3216
Figure 6. Conditioning modules and battery on the wheel. Other transducers were mounted on the tram as well and they are listed here below: • an eddy current proximitor used to measure the wheel rotation. This has been obtained gluing some aluminium blocks on the outer wheel surface as targets and mounting the transducer just in front of the wheel; • a servo-accelerometer on the tram to acquire its accelerations in transverse direction and two potentiometer displacement transducers between different tram carriages. This has allowed to understand when and which parts of the tram were in curve; • a webcam mounted on the bottom to monitor the wheel and the rail when running along the curve. All these sensors were used to acquire data on board but other transducers were adopted to collect data at the ground level. These are: • four microphones (named M), their location on the tested curve is shown in Figure 7. The microphones have been moved to the outer part of the curve (i.e. their position is mirrored to respect of the axis of the rail) when the instrumented wheel faced this side. Microphone M1 is located at 1.75 m from the rail were the instrumented wheel is running on. Microphones M2 and M3 are positioned at 6.75 m from the same rail, the arc length subtended to angle formed by the radii obtained joining microphones’ position with the curve centre is equal to 8 m. Finally microphone M4 is positioned along the bisecting of the above described angle at a distance of 11.8 m from the rail. The height of microphone M1 is equal to 0.5 m while for the other three is 1.2 m. • six accelerometers (named RA) placed on rails. Both the inner and the outer rails have been instrumented with three accelerometers: one on the flange, one on the grooved head and one on the rail web, as shown in Figure 8. Therefore, two different acquisition systems have been used, one for vehicle sensors and the other for ground sensors. The signals of the two systems have been synchronised acquiring the same radio channel and through cross-correlation operations. In all the cases the accelerometers were piezoaccelerometers in order to be able to measure high frequency components [20]. Those used on the wheel rim had a full scale of 20000 m/s2, while the others of 10000 m/s2 to prevent
3217
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
signal saturation. The sampling frequency has been 16667 Hz for both the acquisition systems.
5 1. 7
Vehicle
8 M1 h0.5 6.75 6.75
11.8
M3 h1.2
M2 h=1.2
M4 h=1.2
R17.5 Figure 7. Location and height of the accelerometers, instrumented wheel on the left side. Dimensions in meters.
RA2
RA1
RA3
Figure 8. Position and sensing direction of the rail’s accelerometers.
Looking at Figure 9.c it can be noticed that the spectrogram is characterized by several narrow band frequency packets centered around 530 Hz, 1300 Hz, 1500 Hz, 2500 Hz and 3700 Hz. Each packet has an evolution in time and it is evident that no leading packet exists along the whole record. It is instead interesting to note how the two contributions at 1300-1500 Hz and at 2500 Hz tend to alternate in dominating the wheel vibration. This comes clearly out of Figure 9 part b, in which the contributions of the two frequency bands 1.4-1.8 kHz and 2.2-2.8 kHz are compared to the corresponding global vibration level. The contribution of other packets is minor but not completely negligible. In Figure 10 a zoom of the time history showed in Figure 9 is presented. During the selected two seconds windiw the typical pulsating behavior of the squeal phenomenon is evident. Similar considerations can be done for the results presented in Figure 11, where the rim radial acceleration is analyzed. Again the signal is composed of several frequency packets but, in contrast with the previous case, the packets centered around 530 Hz and 1300 Hz are almost absent. Also in this signal the two alternatively dominating frequency packets are those centered at 1500 Hz and 2500 Hz, Figure 11.b/c clearly summarizes this concept. The overall level of the radial acceleration WA3 is greater than the axial one WA2. It is thus straightforward to conclude that the contribution of the 1423 Hz vibration mode, which corresponds to an almost in-plane (radial) vibration is mainly detected the wheel acceleration WA3. (a)
Acceleration
7500 5000 2500 0 -2500 -5000 -7500
Global level 1.4-1.8 kHz 2.2-2.8 kHz
(b) 200 180 160 140 10
15
20 Time (s)
25
30
(c) 30 Time (s)
A first set of results, from Figure 9 to Figure 12, is presented considering the data acquired while the tram was running along a left sided curve at 10 km/h, with the instrumented wheel mounted on the left bogie, front axle, inner side. In this case the flange back of the wheel goes in contact with the rail grooved head. A second set of results (from Figure 13 to Figure 15) refers to the configuration with the instrumented wheel facing the outward, in this condition there might be a contact between the flange and the running head. Each figure from Figure 9 to Figure 15 (except Figure 11) is composed of three diagrams. In the first one the time history of the signal acquired during the whole tram transit is reported, in the second one there are three different plots: the time evolution of the global level of the measured quantity and the time evolution of the levels restricted to two 1/3 octave frequency bands whose lower and upper limits are respectively equal to 1.4-1.8 kHz and 2.2-2.8 kHz. In the last diagram the spectrogram of the signal is reported. Both wheel acceleration and sound pressure levels in the second diagram of each figure are presented in dB, the reference values being equal to 10-6 m/s2 and to 2x10-5 Pa respectively. The sound pressure levels are the one defined as LZF in the standard [21] (fast time windowing with no frequency weighting) while the acceleration levels are calculated using a time window with a length of 0.125 s. Three sensors are considered as an example of all the collected data: sensor WA2 (rim axial acceleration), sensor WA3 (rim radial acceleration) and microphone M2.
(m/s 2)
MEASUREMENT RESULTS
dB re 10 -6 m/s 2
3
20 10
0
1000 2000 3000 Frequency (Hz)
4000
Figure 9. Left curve, V=10 km/h, front inner wheel. Sensor WA2 (wheel acceleration on the rim in axial direction).
3218
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
7500
2
2500 0 -2500 -5000 -7500 14
14.5
15 Time (s)
15.5
16
Figure 10. Zoom of the acceleration time history of sensor WA2 shown in Figure 9.
(a) 10 5 (Pa)
In Figure 12 the same results are plotted for the microphone M2. It is immediately clear that the signal is now dominated by one frequency component at about 1500 Hz, as evidenced by part b and c of the figure. Even if other contributions are visible in the spectrogram, the results plotted in part b clearly show that, when the squeal noise reaches its higher level, the dominating frequency band is the one centered around 1500 Hz.
(b) 200 180
100 95 90 85 10
15
25
30
20 Time (s)
25
30
30 Time (s)
15
30
20 10
20
0
20 Time (s)
(c)
140 10 (c)
Time (s)
(b) 105
160
10
Global level 1.4-1.8 kHz 2.2-2.8 kHz
-10
dB re 20x10
Global level 1.4-1.8 kHz 2.2-2.8 kHz
0 -5
-6
Pressure
7500 5000 2500 0 -2500 -5000 -7500
dB re 10 -6 m/s 2
Acceleration
2
(m/s )
(a)
Pa
Acceleration (m/s )
5000
a different behavior than in the previous case. Figure 13.c and Figure 14.c show that the packet centered at 1500 Hz is now dominant in the whole transit; only in two moments an event occurs that gives at the wheel a broad band excitation. This is the reason of the two horizontal lines in the spectrogram occurring around 13 s and 19 s. Only in this two moments the band centered at 1500 Hz is exceeded by the band centered at 2500 Hz, as depicted in Figure 13.b and in Figure 14.b. A possible reason for this is that there are two defects on the external rail of the curve, which generate wheel broad band excitation. To confirm this, other data collected on the same curve have the same horizontal lines in the spectrogram and the time spacing between them is the same. The overall radial and axial accelerations are now comparable, but both are lower than those measured when the wheel was mounted on the inner side of the curve (compare Figure 9.a and Figure 11.a with Figure 13.a and Figure 14.a). It is authors’ opinion that this is due to the different wheel-rail contact conditions. In fact, in the first case the contact occurs between the flange back and the rail grooved head, while in the latter case it occurs between the flange and the running head. Acceleration data acquired on the rails confirm this hypothesis.
1000 2000 3000 Frequency (Hz)
4000
Figure 11. Left curve, V=10 km/h, front inner wheel. Sensor WA3 (wheel acceleration on the rim in radial direction). Considering now the results with the instrumented wheel facing outer side of the curve the following considerations can be proposed. The axial and radial acceleration of the rim show
0
1000 2000 3000 Frequency (Hz)
4000
Figure 12. Left curve, V=10 km/h. Sensor M2 (microphone located at 6.75 m from the inner rail).
3219
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
(a) (m/s 2)
7500 5000 2500 0 -2500 -5000 -7500
180
m/s
2
dB re 10
160 140
10
15
20 Time (s)
200 180 160 140
10
25
15
20 Time (s)
25
(c)
(c) 25 Time (s)
25 Time (s)
Global level 1.4-1.8 kHz 2.2-2.8 Hz
(b)
-6
m/s 2 dB re 10
200
Acceleration
Global level 1.4-1.8 kHz 2.2-2.8 kHz
(b)
-6
Acceleration
(m/s 2)
(a) 7500 5000 2500 0 -2500 -5000 -7500
20 15
20 15 10
10
0
4000
4000
(a) 10 (Pa)
5 0 -5 Global level 1.4-1.8 kHz 2.2-2.8 kHz
-10 Pa
(b)
dB re 20x10
In Figure 15 microphone results are reported. The sound pressure levels are comparable with the ones obtained in the previous case (instrumented wheel facing the inner side of the curve) and again the signal is dominated by the band centered around 1500 Hz. In observing the microphone results it is important to point out that the instrument is sensitive to the noise radiated from each part of the tram. In particular, considering as an example the time interval in which the bogie with the instrumented wheel passes in front of the microphone, the transducer receives mainly the noise coming from the two nearest wheels and then the noise produced by all other wheels of the tram. It was observed that the vibration behavior of the wheel depends on its position on the bogie: if the contact between the wheel and the rail occurs between the wheel flange back and the rail grooved head there are several frequency bands, along with the 1.4-1.8 kHz one, involved in the vibration, while, if the contact occurs between the flange and the grooved running head, the dominant band is just the one centered around 1500 Hz. This is also the frequency band which always dominates the sound pressure signal. Recalling the experimental modal analysis results, in this frequency band there exist a mode (1423 Hz, see Figure 3).
1000 2000 3000 Frequency (Hz)
Figure 14. Left curve, V=10 km/h, front outer wheel. Sensor WA3 (wheel acceleration on the rim in radial direction).
105 100 95 90 85
10
15
20 Time (s)
25
(c) 25 Time (s)
Figure 13. Left curve, V=10 km/h, front outer wheel. Sensor WA2 (wheel acceleration on the rim in axial direction).
Pressure
1000 2000 3000 Frequency (Hz)
-6
0
20 15 10 0
1000 2000 3000 Frequency (Hz)
4000
Figure 15. Left curve, V=10 km/h. Sensor M2 (microphone located at 6.75 m from the outer rail).
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
Pressure, dB re 20x10
-6
Pa
The influence of the wheel-rail friction coefficient was also investigated in the test campaign, by comparing measurements taken in case of dry friction and wet track. In Figure 16 the two lines represent respectively the time evolution of the global sound pressure level (fast time windowing) and the same levels in the band 1.4-1.8 kHz, in case of dry contact. As already observed, this band dominates along the whole transit. In Figure 17 the same levels are reported for the test performed in wet conditions It can be observed that a water lubricated contact can completely prevent the squeal noise to occur. In fact not only the global level is lower but also the band evidenced as the mostly noisy in the dry condition has now a marginal effect, indicating thus that the tonal phenomenon is completely inhibited. 105 100 95 90 85 80 75 70 65 60 55 Global level 50 1.4-1.8 kHz 45 0 10 20 30 Time (s)
Pressure, dB re 20x10
-6
Pa
Figure 16. Left curve, V=10 km/h, sensor M2 dry friction. 105 100 95 90 85 80 75 70 65 60 55 50 45 0
REFERENCES [1]
[2]
[3]
[5]
[6]
[7]
[8]
5
10
15 20 Time (s)
25
Figure 17. Left curve, V=10 km/h, sensor M2 reduced friction coefficient (wet track). 4
rail. The wheel accelerations and the synchronous microphone measurements provide interesting information on the different behavior of the two wheels on a same axle. Typically the front outer wheel of a bogie goes in contact with the rail running head, while the flange back of the front inner one goes in contact with the rail grooved head. By comparing the vibrations measured in these two conditions it has been pointed out that the inner wheel exhibits the higher vibration amplitudes and that the nature of the contact is, in this case, capable of exciting a greater number of modes. The sound pressure signal, which is sensitive to the noise radiated from all wheels, shows comparable levels either if the microphone is located on the inner or on the outer side of the curve. Moreover the microphone signal is generally dominated by the frequency band containing the wheel third natural mode (1.4-1.8 kHz). It is the authors’ opinion that this is due to the fact that the contribution of this mode is always present in all wheel vibration signals. Finally, the performed tests confirmed that a water lubricated contact can completely prevent the squeal noise phenomenon to occur. This comes clearly out of the measured sound pressure signals, showing that the tonal noise around 1500Hz completely disappears in case of wet track.
[4]
Global level 1.4-1.8 kHz
[9] [10]
[11]
CONCLUSIONS
The results presented in this paper are intended to provide some new experimental data on tramcar squeal noise phenomenon. During the experimental campaign wheel accelerations, rail accelerations and consequent noise radiation were measured. Attention is focused on the effect of wheel-rail contact conditions in tramway sharp curves, equipped with grooved
3220
[12]
[13]
[14]
P. Liu, H. Zheng, C. Cai, Y.Y. Wang, C. Lu, K.H. Ang, G.R. Liu, Analysis of disc brake squeal using the complex eigenvalue method, Applied Acoustics, Vol. 68, Issue 6, 2007, Pages 603-615. J.-J. Sinou, Transient non-linear dynamic analysis of automotive disc brake squeal – On the need to consider both stability and non-linear analysis, Mechanics Research Communications, Vol. 37, Issue 1, 2010, Pages 96-105. A. Akay, O. Giannini, F. Massi, A. Sestieri, Disc brake squeal characterization through simplified test rigs, Mechanical Systems and Signal Processing, Vol. 23, Issue 8, 2009, Pages 2590-2607. F. Massi, L. Baillet, O. Giannini, A. Sestieri, Brake squeal: Linear and nonlinear numerical approaches, Mechanical Systems and Signal Processing, Vol. 21, Issue 6, 2007, Pages 2374-2393. A.D. Monk-Steel, D.J. Thompson, F. G. de Beer, M.H.A. Janssens, An investigation into the influence of longitudinal creepage on railway squeal noise due to lateral creepage, Journal of Sound and Vibration, Vol. 293, 2006, Pages766-776. M.A. Heckl, I.D. Abrahams, Curve squeal of train wheels, part 1: mathematical model for its generation, Journal of Sound and Vibration, Vol. 229, Issue 3, 2000, Pages 669-693. Heckl, M. A., 2000, “Curve squeal of train wheels, part 2: which wheel modes are prone to squeal?”, Journal of Sound and Vibration 229(3), 695-707. G. Xie et al., Introduction of falling friction coefficients into curvingcalculations for studying curve squeal noise, Vehicle System Dynamics Vol. 44, Issue 1, 2006, Pages 261-271. D. Thompson, C. Jones, A review of the modeling of wheel/rail noise generation, Journal of Sound and Vibration. 2000;231:519-536. Acoustics – Description, measurement and assessment of environmental noise – Part 1: Basic quantities and assessment procedures (ISO 19961:2003). R. Corradi, A. Facchinetti, S. Manzoni, M. Vanali, Effects of track parameters and environmental conditions on tramcar induced squeal noise, in Proceedings of the International Conference on Noise & Vibration Engineering ISMA 2006, Leuven (Belgium), September 1820, Pages3745-3760. N. Vincent, J.R. Koch, H. Chollet, J.Y. Guerder, Curve squeal of hurban rolling stock – Part 1: State of the art and field measurements, Journal of Sound and Vibration Vol. 293,2006, Pages 691-700. C.J.C. Jones, D.J. Thompson, Means of controlling rolling noise at source, in Noise and vibration from high-speed trains, V.V. Krylov, ed., Thomas Telford Publishing, London, UK, 163-183, 2001. A. Cigada, S. Manzoni, M. Redaelli, M. Vanali, Noise and vibrations of railway wheels: generation mechanisms and attenuation, in Railway
Proceedings of the 8th International Conference on Structural Dynamics, EURODYN 2011
[15] [16]
[17]
[18]
[19]
[20] [21]
transportation: policies, technology and perspectives, N.P. Scott, ed., Nova Science Publishers, New York, NY, USA, 237-278, 2009. A. Cigada, S. Manzoni, M. Vanali, Vibro-acoustic characterization of railway wheels, Applied Acoustics Vol. 69, 2008, Pages 530-545. A. Cigada, S. Manzoni, M. Vanali, Geometry effects on the vibroacoustic behaviour of railway resilient wheels, accepted in 2010 and to be published on the Journal of Vibration and Control (DOI: 10.1177/1077546310362452). D.T. Eadie, M. Santoro, W. Powell, Local control of noise and vibration with KELTRACKTM friction modifier and Protector® trackside application: an integrated solution, Journal of Sound and Vibration, Vol. 267, 2003, Pages 761–772. M.A. Heckl, X.Y. Huang, Curve squeal of train wheels, part 3: active control,. Journal of Sound and Vibration, Vol. 229, Issue 3, 2000, Pages709–735. A. Cigada, H. Fehren, S. Manzoni, M. Redaelli, M. Schedewitz, H. Siebald, Investigation on the attenuation of squeal noise from a resilient railway wheel by means of piezo-actuators, In proceedings of the International Conference on Noise and Vibration Engineering ISMA 2008, Leuven (Belgium), 15 - 17 September. E.O. Doebelin, Measurement systems, McGraw-Hill, 2003. CEI EN 61672-1, Electroacoustics – Sound level meters Part1: specifications, 2003-11
3221
