SOME INTERESTING ESTING FEATURES OF TURBOCHARGER ROTOR DYNAMICS

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1 Colloquium DYNAMICS OF MACHINES 2013 Prague, February 5 6, 2013 CzechNC 1. I SOME INTERESTING ESTING FEATURES OF TURBOCHARGER ROTOR DYNAMICS Jiří Šimek Abstract: Turbochargers for combustion engines are simple machines, working with very high parameters, namely extremely high speeds. To check dynamic behaviour of the rotor it is necessary to measure rotor relative e vibration, because vibration on machine casing is damped by two oil films in series. In nstallation of relative sensors into TBCH is difficult, but measured results are very interesting. Some measured results s will be presented and confronted with outcomes of calculation. Behaviour of turbocharger rotors with floating ring bearings inspired the idea of active control of hydrodynamic bearings using piezoactuators. Key words: turbocharger rotors, floating ring bearings, rotor instability, relative e vibrations, subharmonic vibration, active control. INTRODUCTION Turbocharger r (TCH) rotors are supported in floating- ring bearings (Fig. 1, item 4). Floating rings could rotate or their rotation could be prevented, in which case only damping effect outer oil film is utilized. Nonrotating (stalled) floating rings could be integrated into one common mon bushing (double-bushing). bushing). These bearings are undemanding from manufacturing uring point of view and at the same time have good dynamic properties, resulting from high damping of the two oil films arranged in series. Fig. 1 Rotors of medium sized turbocharger with rotating bushings (left) and small turbocharger with nonrotating double-bushing bushing (right) Rotors ors supported in commonly monly used rotating bushings, exhibit in some cases instability of outer oil film. Vibration amplitudes of the rotor in this regime achieve practically the whole of bearing clearanc learance and rotor operation is enabled only due to strong nonlinearity of the oil film. Operation at the regime of instability is very dangerous and that is why nonrotating bushings, which do not have problems with instability, are used more frequently. To achieve rotor stability, nonrotating bushings should have special geometry of inner sliding surface. Theoretical analysis of TCH dynamics yields some unorthodox Ing. Jiří Šimek, CSc., TECHLAB s.r.o., Sokolovská 207, Praha 9, j.simek@techlab.cz

2 results and also measurement of relative vibration showed interesting features of rotor behaviour, which are presented in the paper 2. DYNAMIC ANALYSIS OF TURBOCHARGER ROTORS TCH rotors are specific by heavy impellers at both overhung ends. Very big gyroscopic moments of compressor and turbine impeller result in splitting critical speeds to branches with co-rotating and counter-rotating precession (see Fig. 2). Fig. 2 Campbell diagram of small (left) and medium TCH rotor with gyroscopic moments Critical speed with counter-rotating precession is not excited by unbalance and branch with co-rotating precession is lifted up, so that in some cases there is no bending critical speed within operating speed range. In example of Fig. 2 (right) maximum operating speed of TCH was rpm, while the 1 st bending critical speed was located around rpm. Another unorthodox feature of TCH rotors supported in rotating bushings is the fact that in most cases calculation yields two lowest eigenfrequencies as unstable, while in reality the rotor operates with small stable amplitudes of vibration. 1,2E+07 1,0E+07 8,0E+06 Bearing stiffness vers. speed Kxx Kxy Kyx Kyy 3,0E+07 2,5E+07 2,0E+07 Bearing stiffness vers. speed Kxx=Kyy Kxy Kyx stiffness (N/m) 6,0E+06 4,0E+06 2,0E+06 0,0E+00-2,0E+06-4,0E+06 stiffness (N/m) 1,5E+07 1,0E+07 5,0E+06 0,0E+00-5,0E+06-6,0E speed (rpm) -1,0E speed (rpm) Fig. 3 Stiffness coefficients vers. speed: rotating (left) and nonrotating (right) bushing

3 One of the reasons for this phenomenon can be looked for in bearing stiffness coefficients. While the main (principal) stiffness elements K xx, K yy return journal to its steady state (stable) position, the cross-coupling coefficients K xy, K yx promote journal orbiting around centre of the bearing thus causing instability. As can be seen from Fig. 3, lightly loaded cylindrical bearings, which is also the case of rotating bushings, have cross-coupling stiffness elements much higher than the principal ones (left diagram). On the other hand, lightly loaded profiled bearings, such as e.g. with three-lobbed geometry, have principal stiffness coefficient higher than cross-coupling ones (right diagram) and that is why stability analysis of rotors in profiled (noncircular) bearings indicates in most cases stable operation. Even if indicated instability complicates rotor-dynamic analysis, it can give some guidance as to how to improve rotor operation. It is illustrated by a case of TCH exhibiting very high subharmonic vibration component, which was even dominating in prevailing part of operation speed range. This TCH suffered from rotor breakdowns after very short period of operation even less than 50 hours. Rotor-dynamic analysis showed as usual two unstable eigenvalues. However, dynamic analysis also indicated, that by modification of bearing geometry, namely by increasing outer bearing clearance, stability reserve of the 3 rd eigenvalue could be increased by more than 80%. Test of TCH with modified bearing geometry showed that subharmonic vibration component and thus also overall rotor vibrations were substantially reduced (see Fig. 4) as was predicted by calculation. amplitude(um) Amplitudes of rotor 1 in compressor impeller location speed(rpm) subharmonic synchronous overall amplitude(um) Amplitudes ofrotor2in compressor impeller location speed(rpm subharmonic synchronous overall Fig. 4 Comparison of subharmonic vibration component measured at compressor impeller location before and after modification of TCH bearing geometry 3. EXAMPLES OF MEASURED TURBOCHARGER ROTOR VIBRATIONS The 1 st example presents the case of instability of outer oil film in relatively big TCH (output of the order of several hundreds of kw) with rotating floating bushings. Fig. 5 Vibration signal in time domain and frequency spectra (rotating bushings)

4 As can be seen from Fig. 5 (left diagram), rotor vibration measured at rpm next to both journal bearings achieved values up to 140 µm, which is close to whole bearing clearance. Top down are shown signals: rotor - compressor side (RC), rotor - turbine side (RT), bushing - compressor side (BC) and bushing - turbine side (BT). Frequency spectra at the right side of Fig. 5 show dominating subharmonic frequency component around 55 Hz, which is one half of bushing speed indicating instability of outer oil film. Vibration amplitudes adequate to those at rpm were measured in the whole operating speed range. By replacing rotating floating bushings with stalled bushings with three-lobbed inner geometry (Fig. 6), rotor instability was suppressed. Fig. 6 Bidirectional (left) and unidirectional (right) three-lobbed bearing geometry Fig. 7 Vibration signal in time domain and frequency spectra (stalled bushings) As is evident from Fig.7, TCH rotor in stalled bushings exhibited stable operation with maximum vibration amplitude up to 25 µm. Only synchronous vibration component 700 Hz, corresponding to speed of rpm, is visible in frequency spectra. Top down in Fig. 7 are the same signals as in Fig. 5. Operation of TCH rotor in stalled bushings was stable in the whole speed range, which was predicted also by dynamic analysis. As another interesting example of measured results we will present rotor response to external excitation, which is quite common in TCH operating on combustion engines. Fig. 8 Vibration signal in time domain and frequency spectra with harmonic external excitation

5 By connecting TCH to vibrator with counter-rotating unbalances it was possible to excite it by harmonic vibration with frequency of 25 Hz and acceleration equal to ±5 g (±50 m.s -2 ). Fig. 8 shows respective vibration signals in time domain and frequency spectra. Top down are signals: rotor - compressor side (CS), rotor - turbine side (RT), bushing - turbine side (BC), acceleration in radial direction at TCH casing. It is evident that TCH rotor and casing vibrate with dominant frequency of excitation, while frequency of operation speed 1,133 khz ( rpm) is hardly recognizable in frequency spectra. RMS (root mean square) values of rotor maximum vibration amplitudes did not exceed 17 µm, which corresponds to double-amplitude of 48 µm. With overall bearing clearance over 120 µm maximum relative rotor vibration amplitude constitutes less than 50% of bearing clearance. It is evident that TCH operation with such a level of external excitation can be considered quite safe. Quite interesting is also passage of TCH through the regime of surge, which is illustrated by Figs. 9 and 10. Fig 9 shows passage through surge of a big TCH, such as in Fig. 5. The upper and middle signals indicate rotor relative vibrations at compressor and turbine side respectively. Rotor excursions at compressor side (near the source of disturbance) achieve nearly 60 µm, while at turbine side they are only about 13 µm. The lower signal is measured oil film thickness in TCH thrust bearing. It can be seen, that change of oil film thickness amounts to nearly 90 µm, which is prevailing part of axial clearance. Fig. 9 Rotor vibration and oil film thickness Fig. 10 Rotor and bushing vibration and at surge of big TCH acceleration at TCH casing at surge Top down in Fig. 10 are vibration signals of the rotor at turbine and compressor side, bushing at turbine and compressor side during passage through the surge. While rotor at compressor side achieved maximum excursions more than 80 µm, at turbine side it was only about 30 µm. The interesting fact is, that rotor and bushing deviation at compressor side are in-phase, while at turbine side they are out-of-phase. Fig. 11 Influence of surge on axial force

6 Fig. 11 is another interesting example of measured result concerning surge; this time it illustrates course of axial force during surge. Axial force was measured by 3 individual strain gauge load cells and Fig. 11 shows signals of all 3 load cells at two different regimes of surge. In the 1 st case (left) maximum axial force deviation was 2,0 kn, while in the second case it was only 1,2 kn; the average value of steady-state axial force after the regime surge ended was only about 400 N and 280 N. 4. PROJECT INSPIRED BY TURBOCHARGER ROTOR BEHAVIOUR The idea of influencing the rotor behaviour by active control of journal bearings was inspired by behaviour of TCH rotors, which indicate symptoms of instability of outer oil film, but the instability does not develop due to non-synchronous excitation. Experiments carried out in the frame of grant project GA ČR on test stand build in VSB TU Ostrava proved, that it is possible to increase substantially rotor stability by active control of journal bearings [4]. This research will continue in years with slightly modified test stand (se Fig. 12), enabling also experiments with elastic shaft. Fig. 12 Test stand with elastic shaft for active control of journal bearings 5. CONCLUSIONS Turbocharger rotors show some specific features, which are described in the paper. Calculated results indicate rotor instability, which occurs only in limited number of cases with rotating floating bushing. Instability can be suppressed by modification of bearing geometry, from which most effective is usage of non-rotating bushings. Experimental results acquired on TCH inspired research project on affecting rotor behaviour by active control of journal bearings by piezoactuators. 6. ACKNOWLEDGEMENT This work was supported by the Czech Science Foundation under project No. 101/12/2520 Active damping of rotor vibration by parametric excitation of sliding bearings. 7. REFERENCES [1] Šimek, J.: Measurement of NR20SJ turbocharger relative vibration of the rotor and floating bushing. Technical report TECHLAB No [2] Šimek, J.-Svoboda, R.: Calculation of bearing support and dynamic analysis of NR20SJ turbocharger rotor. Technical report TECHLAB No [3] Šimek, J.: Measurement of TCR12 turbocharger vibration with external excitation. Technical report TECHLAB No [4] Tůma, J.et al: A laboratory test stand for active control of journal bearings. Dynamics of Machines 2010, p