Modal Properties of Drive Train in Horizontal-Axis Wind Turbine

Transcription

1 MECAHITECH 11, vol. 3, year: 011 Modal Properties of Drive Train in Horizontal-Axis Wind Turbine Michael Todorov 1, Geoi Vukov 1 Technical Univeity of Sofia, Univeity of Forestry-Sofia 8, St. Kliment Ohridsky blvd., 1000 Sofia, Bulgaria michael.todorov@tu-sofia.bg Abstract The staring point in the reseah of the vibrations is the finding of natural frequencies and mode shapes, which indicate the frequency range of interest. A dynamic multibody model for determination of the toional vibrations of a wind turbine drive train is presented in this paper. The model of a wind turbine consists of a rotor with rigid blades, elastic shafts, a drive train and a generator. The drive train has a gearbox with three gear stages. The gear stages include two high-speed stages (helical gear pai) and a low-speed planetary gear stage (three identical planets with spur teeth, sun and a fixed ring wheel). The model consists of 10 bodies and has 11 degrees of freedom. The model takes into account the stiffness of the engaged tooth pai and shafts. In this model the aerodynamic and generator torques are applied as external loads. Computer simulation is performed by MATLAB. The natural frequencies and vibration modes are obtained for an industrial wind turbine. Keywords Wind turbine, drive train, toional vibrations, modal analysis. Introduction The wind eney application has been growing rapidly for the last few yea. In the last ten yea the global installed capacity of wind eney has increased 0 times. This trend is expected to continue in Europe. The increase in the rotor and hence size of the turbine leads to a complicated design of the drive train in the wind turbine beside higher requirements of turbine reliability. Design calculations for a wind turbine base on simulation of mechanical loads on the turbine components caused by external foes. The external foes are the wind, the electricity grid and sea waves for offshore applications. The multi-body simulation techniques are used to analyze the loads on internal components of drive trains. The simplest model with one degree of freedom (DOF) for each drive train component is used to investigate only toional vibrations in the drive train. In this model all bodies have one DOF, i.e. the rotation around their axis of symmetry. Therefore, the coupling of two bodies involves DOF s. Gear contact foes between two wheels are modelled with a linear spring acting in the plane of action along the contact line (normal to the tooth surface), [17, 1]. More complex model with 6 DOF s for each drive train component is used for investigation of the influence of bearing stiffness on the internal dynamics of the drive train. All drive train components are treated as rigid bodies. The linkages in the multi-body model, representing the bearing and tooth flexibilities, have 1 DOF s, [38]. Finally, it is used a flexible model in which the drive train components are modelled as finite element models instead of rigid bodies, [, 4]. This model adds a possibility of calculating stress and deformation in the drive train components in some time. Any addition to the model leads to additional information about dynamics of the drive train but makes the modelling and the simulation more complicated. The modern wind turbines have a planetary gearbox. Studies on the vibrations in a planetary gear system have been done in [, 10, 18, 19, 0, 3, 34]. The tooth meshes are modelled as a linear spring with stiffness which is a time function. For this reason the vibration equations of a planetary gear system are differential equations with periodic coefficients, [, 5, 18, 19 0, 3]. References [6, 7, 11, 1, 14, 19, 3] investigate the vibrations of compound planetary gea. The applications of these modelling techniques on different drive trains of wind turbines are presented in [13, 15, 16, 4-7, 8, 9, 31-33]. References [35, 36] present the numerical investigations for the given wind turbine in this paper, where the meshes stiffness are modelled as constant springs. In this case the differential equations, which describe the toional vibrations of the wind turbine, have constant coefficients. Reference [37] presents a dynamical model of an wind turbine, but the meshes stiffness are modelled as a time function. 160

2 MECAHITECH 11, vol. 3, year: 011 This paper is based on the work [37]. The Lagrange s equations are used to obtain the equations of the toional vibrations of the wind turbine, [3, 8,, 30, 40]. Computational modal analysis of an example wind turbine is presented. Dynamic Model of Wind Turbine The wind turbine consists of a rotor, a drive train and a generator (Fig.1). The drive train has a gearbox with three stages. The gear stages include two high-speed parallel gear stages (helical gear pai) and a lowspeed planetary gear stage (three identical planets with spur teeth, sun and fixed ring wheel) (Fig.). Figure 1: Schematic sketch of wind turbine Figure : Sketch of gearbox: h-hull, c-carrier, p1,,3-planets, s-sun, g1,,3-gea The dynamic multi-body model is shown in Fig.3. It consists of a rotor with 3 rigid blades, a low-speed elastic shaft, a gearbox with 3 gear stages, a high-speed elastic shaft and a generator rotor. Thus, the model consists of 10 bodies and 11 DOF s. 161

3 MECAHITECH 11, vol. 3, year: 011 Figure 3. Dynamical model of wind turbine The gear contact foes between wheels are modeled by linear spring acting in the plane of action along the contact line (normal to the tooth surface), [1, 41]. The stiffness gear is defined as a normal distributed tooth foe in a normal plane causing the deformation of one or more engaging tooth pai, over a distance of 1 μm, normal to a envolvent profile in a normal plane, [9]. This deformation results from the bending of the teeth in contact between the two gear wheels, the one of which is fixed and the other is loaded. The stiffness varies in the time and can be expressed in a time Fourier series form, [1, 18-0, 34, 37]. Damping and friction foes are not included. These assumptions are valid for heavily to moderately loaded gea that are correctly for a lae wind turbine, [17, 1]. It is also accepted that the masses of the planets are identical. The vector of the generalized co-ordinates is q (1) T h c r p1 p p3 s g1 g g3 where (i=h,c,r,p1,p,p3,s,g1,g,g3,gn) are the rotational angles of the ring (gearbox hull), carrier, rotor (hub), i planet 1, planet, planet 3, sun, gear 1, gear, gear 3 and the generator rotor (Fig. 3). The differential equations, describing the toional vibrations of the wind turbine, are [ M ]{ q } [ C( C ]{ q} { T} () where M is the inertia matrix and C is the stiffness matrix. The matrix C results from the carrier rotation. The vector of the external foes, caused by the wind and the electricity grid, is T T T aero T gen (3) where T aero and T gen are the aerodynamic and electromagnetic torques. The non-zero numbe of inertia matrix M are m1,1 Jh Jc mclcr Jr mrlcr 3J p 3m plcr 3mp J s mslcr I g mg lg I g 3 mg 3 lg 3 ; m1, m,1 Jc Jr 3J p 3mp ; m1,3 m3,1 m,3 m3, m3, 3 J r ; gn I g 1 mg 1 lg 1 m1,4 m4,1 m1,5 m1,6 m6,1 m5,1 m,4 m4, m,5 m5, m,6 m6, m4,4 m5,5 m6,6 I p; m1,7 m7,1 m7, 7 I s ; m1,8 m8,1 m8,8 I g1 ; m1,9 m9,1 m9,9 I g ; m1,10 m10,1 m10,10 I g3 ; m, Jc Jr 3J p 3mp ; m11,11 I gn. 16

4 MECAHITECH 11, vol. 3, year: 011 The non-zero numbe of stiffness matrix C are c 1,1 Cs 1 l s 1 ls C ( 3 rr 3 3 6rR C ps( cos Cg1( l l C ( l l ; g1 g g3 g g rR C ps ( cos ; r p rr cos t cost C ps( cos t c1, c,1 C ( c1,4 c4,1 C ( cost ; c1,5 c5,1 C ( C ps( p r p rr cos60 t cos60 t r r r cos60 t r r cos60 t; c p c1,6 c6,1 C ( p s p r r r cos 60 t r r cos 60 t R p c p r p cos60 t cos60 t; C ps ( c1,7 c7,1 C ps( 3 cos ; c1,8 c8,1 Cg1 ( 1l g1 lg cos cos ; c9,1 Cg1( 1l g1 lg cos cos Cg3( l g lg3 cos cos ; c C r l l cos cos ; c C ( 3r 3r C ( 3 c1,9, rr c p ps 3 ; c3,3 1 ; c,4 c4, C ( r p cos t C ps ( r p cos t ; c5, C ( r p cos60 t C ps ( r p cos60 t ; c6, C ( r p cos60 t C ps ( r p cos60 t ; c1,10 10,1 g3( g3 g g3 c,3 C c,5 c,6 c,7 c7, 3C ps( cos; c4,4 c5,5 c6,6 C ( C ps ( ; c4,7 c7,4 C ps( cost; c5,7 c7,5 C ps( cos60 t ; c6,7 c7,6 C ps( cos60 t; c7,7 C 3C ps( ; c7,8 C ; c 8,8 C Cg 1 ( 1 cos cos sin ; c 8,9 Cg 1 ( 1 cos cos sin ; C g 1 ( 1 Cg 3 ( cos cos sin ; c 9,9 c9,10 Cg3( 3cos cos sin ; c 10,10 C 3 Cg 3 ( 3 cos cos sin ; c10,11 c11,11 C3. The non-zero number of C is c 3 m, p r. c Numerical Results The rotor, drive train and generator characteristics of an example turbine can be seen in Ref. [37]. To determinate the natural frequencies and mode shapes the time invariant system is considered. All mesh stiffnesses are considered to be constant and equal to their average stiffness over one mesh cycle. All externally apllied foes are assumed to be zero. The natural frequencies and modes shapes are obtained by Equation () 1 [ ] [ C] [ E] q 0 M (4) All calculations are accomplished using the codes of MATLAB. The natural frequencies in Hz are The mode shapes are shown in Fig.4 163

5 Proceedings of International Conference On Innovations, Recent Trends And Challenges In Mechatronics, MECAHITECH 11, vol. 3, year: 011 Figure 4: Mode shapes of wind turbine For beter ilustration the mode shapes are also presented in Fig Fig.5. Mode shapes for natural frequency 78 Hz. Solid lines are the equilibrium positions and dashed lines are the Fig.6. Mode shapes for natural frequency 387 Hz. Solid lines are the equilibrium positions and dashed lines are the 164

6 MECAHITECH 11, vol. 3, year: 011 Fig.7. Mode shapes for natural frequency 88 Hz. Solid lines are the equilibrium positions and dashed lines are the Fig.8. Mode shapes for natural frequency 37 Hz. Solid lines are the equilibrium positions and dashed lines are the Fig.9. Mode shapes for natural frequency 34 Hz. Solid lines are the equilibrium positions and dashed lines are the 165

7 MECAHITECH 11, vol. 3, year: 011 Fig.10. Mode shapes for natural frequency 318 Hz. Solid lines are the equilibrium positions and dashed lines are the Fig.11. Mode shapes for natural frequency 19 Hz. Solid lines are the equilibrium positions and dashed lines are the Fig.1. Mode shapes for natural frequency 178 Hz. Solid lines are the equilibrium positions and dashed lines are the 166

8 MECAHITECH 11, vol. 3, year: 011 Fig.13. Mode shapes for natural frequency 6 Hz. Solid lines are the equilibrium positions and dashed lines are the Fig.14. Mode shapes for natural frequency.5 Hz. Solid lines are the equilibrium positions and dashed lines are the Conclusions This work identifies the properties of the natural frequency spectra and vibration modes of a wind turbine with a complex drive train. The developed model and results, obtained by its help, are useful for the gearbox designing and scientific reseahes. The prediction of natural frequencies allows taking actions for avoiding of resonance regimes in the gearbox and wind turbine designing. This model gives an opportunity for investigation of toional vibrations excited by wind and electromagnetic loads. The results can be used for wind turbine vibrodiagnostics. Bibliography [1] A. Ams, M. Lorenz, G. Dunchev, I. Kralov, P. Sinapov, Investigation of dynamic loads in a two-stage cylindrical gear, Proc. BulTrans-009, Sozopol, September 4-6, 009, pp (in Bulgarian). [] V. Ambarisha, R. Parker, Suppression of Planet Mode Response in Planet Gear Dynamics Through Mesh Phasing, J. Vibration and Acoustics, 006, vol.18, pp [3] F. Amirouche, Fundamentals of Multibody Dynamics Theory and Applications, Birkhäuser, Boston, 006. [4] A. Andeson, L. Vedmar, A Dynamic Model to Determine Vibrations in Involute Helical Gea, J. Sound and Vibrations, 60, pp.195-1, 003. [5] C.-J. Bahk, R. Parker, Analytical Solution for the Nonlinear Dynamics of Planetary Gea, J. Computational and Nonlinear Dynamics, April 011, vol.6/ pp , [6] V. Batinić, Modal Analysis of Planetary Gear Trains, J. Mechanical Engineering Design, vol.4, No 1, 001, pp [7] F. Choy, Y. Ruan, J. Zakrajsek, F. Oswald, Modal Simulation of Gearbox Vibration with Experimental Correlation, AIAA , 199. [8] M. Coutinho, Dynamic Simulations of Multibody Systems, New-York, Springer-Verlag, 001. [9] Deutsches Institut für Normung, Calculation of Load Capacity of Cylindrical gea, DIN 3990,

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