# Study of coupling between bending and torsional vibration of cracked rotor system supported by radial active magnetic bearings

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2 types of excitation, steady-state as well as transient. Because of that the breathing model of crack has been adopted. To the author's knowledge the dynamics of a cracked rotor system with an active feedback is investigated only in the articles [5], [7] and [8]. The model rotor system in these studies was consisting of massless flexible shaft with the transverse crack supported by two identical rigid bearings and one magnetic bearing and with disc located at the midspan. The active magnetic bearing and transverse crack are located in the proximity of disc and active magnetic bearing is incorporated into the computational model of rotor system by means of linear force coupling. The simple breathing model of crack developed in [4] is used. 2. Equation of motion of rotor system with transverse fatigue crack It is assumed that the model rotor system with transverse fatigue crack has following properties: (i) the shaft is represented by a flexible beam-like body that is discretized into finite elements, (ii) shaft element is modelled on the basis of Bernouli-Euler beam theory with six degrees of freedom per node (i.e. two translations, two rotations, axial displacement and torsion), (iii) the stationary (i.e. non-rotating) part is considered to be absolutely rigid and fixed, (iv) the rotor is coupled with the stationary part through radial AMB and massless axial bearing, (v) the discs are considered to be absolutely rigid axisymmetric bodies, (vi) inertia and gyroscopic effects of the rotating parts are taken into account, (vii) material damping of the shaft and other kinds of damping are regarded as linear, (viii) the shaft contains a transverse fatigue crack, which influences only the stiffness matrix of the rotor system, (ix) the crack edge is assumed to be straight line, (x) the rotor is loaded by a gravitational force and periodic time histories forces and (xi) the rotor rotates at constant angular speed. The equation of motion of the complete rotor system in the stationary co-ordinate system has the following form M q () t + ( B + η VK SH + ω G) q ( t) + { K[ q( t) ] + ω K C} q( t) = f ( q i) + f ( t) + f, M, A V q OP = q OP ( t), q ( 0) = q0, q( 0) q 0 [ ] C where M, B, G, K q() t, Κ SH, K are the mass matrix, the damping matrix (external damping and damping of material), the gyroscopic effects matrix, the stiffness matrix of the rotor system with the transverse crack, the stiffness matrix of the shaft and the circulation matrix of the rotor system respectively, q, q, q are vectors of generalized nodal displacements, velocities and accelerations respectively, f M, f A, f V, i are vectors of magnetic forces, generalized forces exerting on the rotor system (external and constraint forces) and control currents passing by core of electromagnets respectively, while ω is the angular speed of a shaft, η V is the coefficient of viscous damping of material and t is the time. Further details on element formulation and element matrices of a rotor system without a crack can be found elsewhere [6]. The investigated rotor system is supported by radial AMB consisting of four electromagnets, which are uniformly distributed around the bearing circumference. The motion of rotor in horizontal and vertical direction is controlled by two independent current PD feedback controllers. AMB are incorporated into the equation of motion (1) by means of nonlinear force coupling. Formulas for components of vector of magnetic force exerting on rotor in the horizontal and vertical direction and formulas of current PD feedback controller are given in [3]. = (1) =, (2) 428

3 3. Derivation of stiffness matrix of cracked shaft element The finite elements employed are based on Bernouli-Euler beam theory. The influence of the crack is taken into account through modification of the stiffness matrix obtained by finite element formulation. Let us consider a shaft element with diameter D (R=D/2), length L and a transverse crack situated at a distance x from the first node as shown in fig. 1. The shaft element is loaded by shear forces f 1, f 2 and f 7, f 8, bending moments f 3, f 4 and f 9, f 10, axial forces f 5 and f 11 and torsional moments f 6 and f 12. The presence of the transverse fatigue crack in the shaft element introduces a local Fig. 1. A shaft element with the transverse crack. flexibility, which is described by a local flexibility matrix C e cr. The dimension of the local flexibility matrix is in general dependent on the number of degrees of freedom of the shaft element. Thus, in this case the local flexibility matrix is six by six matrix. e Calculation of the flexibility matrix of cracked shaft element C is determined as a sum of e local flexibility matrix of cracked shaft element C cr and flexibility matrix of shaft element e without the crack C ncr (see [2]). Using concepts of fracture mechanics the elements of local flexibility matrix due to crack are given according to [2] by the following expression Fig. 2. The cross-section of shaft with the crack. 2 f f 2 ( ci, j) = J ( ) cr i a j 0 s s1 α dα d γ, (3) where f i, f j are corresponding loads (forces and moments) for i, j = 1, 2,..., 6 th load index, α is the immediate depth of the crack, a is the total depth of the crack, s 1 is half a width of crack, s 2 is width of crack, γ is the distance between rectangular strip of width d γ and vertical axis of the shaft (see fig. 2) and J(α) is the strain energy density function. The increase in the displacement due to the presence of the crack is derived for the first node of shaft element. The e complete stiffness matrix K cr is then obtained by means of transformation matrix T, which is obtained by considering the static equilibrium of the finite element [1]. The stiffness matrix of cracked element in rotor-fixed coordinate system has the following form K e cr e 1 T ( ) T = T C. (4) The stiffness matrix of cracked element (4) has to be transferred into the stationary coordinate system and then it is placed into correct position in the global stiffness matrix of the cracked rotor system. 429

4 4. Incorporation of breathing of crack into computational model of rotor system Size of elements of local flexibility matrix (3) depends on size of crack opening or closing (fig. 2). For fully open crack the integration limits in expression (3) are from - s1 to s 1. The breathing of crack is incorporated into the computational model of the rotor system by using concept of crack closure line (CCL), which was proposed in the article [1]. The CCL is an imaginary line perpendicular to the crack edge, which is separating opened and closed part of crack (see fig. 2). It is assumed that crack edge is straight line and its total length is equal to 2s 1. The crack edge is divided into lines of equal length and stress intensity factor is calculated in the middle of each line. Positive stress intensity factor indicates tensile stress field and the crack is open. Negative stress intensity factor indicates compressive stress field and crack is closed at that point. The total stress intensity factor is given by K K = K + K + K, (5) I 3 I 3 = σ 3 πα F2 ( α h), K I 4 = σ 4 πα F1 ( α h), K 5 σ 5 πα F1 ( α h) σ 2h ( f + f x) =, 4 π R σ 4γ I 4 ( f f x) where F 1 and F 2 are configuration correction factors for stress intensity factors and meaning of other symbols is evident from fig. 1 and fig Computation of response of rotor system with breathing crack on excitation by centrifugal forces due to unbalance of rotating parts The appropriate direct integration method has been employed to obtain transient response of the rotor system with breathing crack on excitation by unbalance of the rotating parts. From the transient response the steady-state response of a rotor system is estimated. The nonlinear equation of motion (1) of examined a rotor system are solved in the stationary co-ordinate system. Due to breathing of crack the stiffness matrix of the cracked rotor system K [ q( t) ] is dependent on calculated response of the rotor system. Stiffness matrix of the rotor system with the transverse crack K [ q() t ] is assumed to be constant for short period of time (e.g. one degree of rotation), because variation of integration limits in the expression (3) is insignificant. After every short period of time the stiffness matrix needs to be updated to take into account the breathing behaviour of the cracked rotor. For this purpose nodal forces in the stationary co-ordinate system f S are calculated where q is the new response vector obtained by integration of equation of motion (1) and K [ q() t ] is the stiffness matrix of the rotor system with the transverse crack in the stationary co-ordinates. Nodal forces calculated in the stationary co-ordinate system f S must be transformed into the rotor-fixed-co-ordinates f R by an appropriate transformation matrix T R (see [1]). By means of these nodal forces f R the total stress intensity factor (5) along the crack edge is calculated. Therefore the appropriate integration limits in the expression (3) for computing corresponding stiffness matrix of a cracked element are known. I =, f S π R f σ =, π R { K[ q( t) ] + ω K C} q( t) I =, (6) 2 2 h = 2 R γ, (7) =, (8) 430

5 6. Test rotor system with transverse fatigue crack Presented computational procedures were built up in the computer system MATLAB (R2007a) and tested by means of computer simulations on the test rotor system (fig. 3). The investigated rotor system is consisting of the shaft (SH) driven by electromotor (EL) through the clutch (CL) and two discs (D1, D2) fixed on cantilever as shown in the fig. 3. The rotor is connected with the bed plate (BP) by means of two radial AMB (MB1 and MB2). The transverse fatigue crack Fig. 3. A sketch of the test rotor system. (CR) is located on the shaft (SH) in the proximity of disc (D1) on the side of magnetic bearing (MB2). The shaft in the computational model is represented by a beam-like body that is discretized into twenty four finite elements of equal length (95 mm) and more detailed data are given in [3]. Rayleigh damping coefficients α, β and the coefficient of viscous material damping η V are assumed equal to 4 s -1, 0 s and s -1 respectively. The current PD feedback controllers and MB1 and MB2 are assumed to be identical and their parameters are 5 1 given in [3]. The discrete axial bearing stiffness is in the axial direction k x = 1 10 N m and 1 torsional stiffness is k t = 1N m rad. The eccentricity of both discs is equal to mm and is located on positive axis z (see fig. 2). The numerical simulations were performed with several depths of crack a = 13.2 mm, a = 26.4 mm, a = 39.6 mm which are corresponding to 15%, 30%, 45% of diameter of shaft. In some cases of computation of response the additional torsional excitation is applied. The time history of torsional moment is M t = M A sin( ωt t), where M A = 300 N m is the amplitude of torsional moment and ω is the torsional excitation frequency. 7. Results of numerical experiments with test rotor system t The assembly of the Campbell s diagram of uncracked rotor system is made in the neighbourhood of equilibrium position. The results of Modes of vibration and critical several values of critical speeds and its higher harmonics are stated in tab. 1. speeds [rads 1 ] 1 st axial mode 18.1 The response of a rotor system on excitation by centrifugal forces due to unbalance of the rotating parts 1 st 76.1 bending mode has been solved with direct integration method. In the computer system MATLAB the implemented functions 2 nd bending mode ode15s or ode23 were used. After dying out of the 1 st torsional mode initial transient component from obtained response the steady-state response is estimated. The presented frequency spectra were computed from steady-state re- 3 rd bending mode sponse by applying discrete Fourier transform. Tab. 1. Critical speeds. Since cracked shaft is causing coupling of longitudinal, torsional and lateral bending vibration the re- 431

6 sponse of uncracked rotor system is computed with and without torsional excitation. Angular 1 speed is 15.2 rads, which is approximately 1/5 th of first bending critical speed 1 (e.g rads 1 ). Then the harmonic torsional excitation with ω t = 76.1 rads was applied at the sixth node of discretized rotor system (at the disc location). The computed responses shown in fig. 4 and fig. 5 were equivalent to responses computed without torsional excitation. In the case of uncracked rotor system the coupling mechanism between torsional excitation and the response in lateral direction doesn t exist. Response in the torsional direction shows the presence of torsional excitation frequency (fig. 6). The response of rotor system in vertical direction is shown in left fig. 5 and the corresponding frequency spectra contains except rotational frequency its second harmonic frequency with much smaller amplitude caused by nonlinear magnetic forces (right fig. 5). The response of rotor system with the crack of depth 13.2 mm and angular speed rads excited only by centrifugal forces due to unbalance of the rotating parts is now investigated. The lateral bending vibration in both horizontal (fig. 7) and vertical (fig. 8) direction contain rotational frequency and its higher harmonics. Also in the longitudinal (fig. 9) and torsional (fig. 10) vibration the rotational frequency and its higher harmonics were identified. Higher harmonics indicate the coupling phenomenon between the bending and the longitudinal vibration as well as between the bending and the torsional vibration. The unbalance excitation applied on rotor system with cracked shaft generated lateral, axial and torsional vibration. Fig. 4. Time domain response in horizontal direction of an uncracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). Fig. 5. Time domain response in vertical direction of an uncracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). 432

7 Fig. 6. Torsional response in time domain of an uncracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). Next analysis has been carried out with parameters of rotor system from previous analysis 1 and in addition the torsional excitation with frequency ω t = 76.1 rads was applied, which is equal to the 1 st bending critical speed. Although in this case the torsional excitation is applied on the rotor system, the peak of the 1 st bending critical speed (left fig. 11) has the similar value as for rotor without torsional excitation (right fig. 7). The amplitude variations of frequency components of response in the horizontal direction for three depths of crack are shown in right fig. 11. All frequency components are in the same way sensitive on depth of crack. Fig. 7. Time domain response in horizontal direction of a cracked rotor system (left) and corresponding frequency domain response (right). Fig. 8. Time domain response in vertical direction of a cracked rotor system (left) and corresponding frequency domain response (right). 433

8 Fig. 9. Time domain response in axial direction of a cracked rotor system (left) and corresponding frequency domain response (right). Fig. 10. Time domain torsional response of a cracked rotor system (left) and corresponding frequency domain response (right). Fig. 11. Frequency domain response in horizontal direction of a cracked rotor system (left) and variation of amplitudes of rotational frequency and its higher harmonics (right). 1 The cracked rotor with depth of crack a = 13.2 mm rotating at angular speed 17.8 rads, which is not equal to integer fraction of bending critical speed, excited by torsional moment 1 with frequency equal ω t = 43.6 rads (fig. 15) is now studied. The spectra of lateral vibration (right fig. 12 and right fig. 13) and axial vibration (right fig. 14) shows the rotational frequency and its higher harmonics. Then lateral vibration spectra (right fig. 12 and right fig. 13) contain around the torsional excitation frequency ω sums and differences of rotational fre- t 434

9 quency and torsional excitation frequency like are ω t ± ω and ωt ± 2ω. Arguably some of higher harmonics of rotational frequency in these frequency spectra are caused by interaction between nonlinearities of magnetic forces and nonlinear model of breathing crack. The side frequencies are present due to interaction between torsional excitation frequency and the lateral unbalance excitation. Fig. 12. Time domain response in horizontal direction of a cracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). Fig. 13. Time domain response in vertical direction of a cracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). Fig. 14. Time domain response in axial direction of a cracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). 435

10 Fig. 15. Time domain torsional response of a cracked rotor system with torsional excitation (left) and corresponding frequency domain response (right). 8. Conclusion Finite element method has been used for study of steady-state response of rotor system supported by radial AMB with the transverse breathing crack in a shaft. The steady-state response of a rotor system on excitation by centrifugal forces due to unbalance of the rotating parts and in some special cases additional torsional harmonic excitation is studied in both time and frequency domain. The unbalance excitation applied on the cracked rotor system in lateral direction excited response in torsional and axial direction, which is indicating coupling mechanism between bending and torsional vibration and between bending and longitudinal vibration. When an angular speed of rotor was not an integer fraction of its bending critical speed, the frequency spectra contained side frequencies around the frequency of torsional excitation. The results presented in this work can therefore help with detection of cracks in the shaft. Acknowledgement This research work has been supported by the postdoctoral grant of the Grant Agency of the Czech Republic (No. 101/07/P368) and by the research project No. MSM of the Czech Ministry of Education. Their help is gratefully acknowledged. References [1] A. K. Darpe, K. Gupta, A. Chawla, Coupled Bending, Longitudinal and Torsional Vibrations of a Cracked Rotor, Journal of Sound and Vibration 269 (2004), pp [2] A. D. Dimarogonas, S. A. Paipetis, Analytical Methods in Rotor Dynamics, London: Applied Science Publishers, [3] P. Ferfecki, Počítačové modelování rotorové soustavy uložené v radiálních aktivních magnetických ložiskách, Ph.D. thesis, VŠB-TUO Fakulta strojní, Ostrava, [4] R. A. Gasch, Survey of the Dynamics Behaviour of a Simple Rotating Shaft with a Transverse Crack, Journal of Sound and Vibration 160 (1993), pp [5] G. Mani, D. D. Quinn, M. Kasarda, Active Health Monitoring in a Rotating Cracked Shaft Using Active Magnetic Bearings as Force Actuators, Journal of Sound and Vibration 294 (2006), pp [6] J. Slavík, V. Stejskal, V. Zeman, Základy dynamiky strojů, Press ČVUT, Praha, [7] C. Zhu, D. A. Robb, D. J. Ewins, The Dynamics of a Cracked Rotor with an Active Magnetic Bearing, Journal of Sound and Vibration 265 (2003), pp [8] C. Zhu, D. A. Robb, D. J. Ewins, The Dynamics of a Cracked Rotor with an Active Feedback Control System, Proceedings of the 8th International Symposium of Magnetic Bearings, Zurich 2000, pp

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