# Dynamic Analysis of a Timoshenko Beam Subjected to an Accelerating Mass Using Spectral Element Method

Authors: Guangsong Chen; Linfang Qian; Qiang Yin
Journal: Shock and Vibration
 (2014)

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## Abstract

This paper presents formulations for a Timoshenko beam subjected to an accelerating mass using spectral element method in time domain (TSEM). Vertical displacement and bending rotation of the beam were interpolated by Lagrange polynomials supported on the Gauss-Lobatto-Legendre (GLL) points. By using GLL integration rule, the mass matrix was diagonal and the dynamic responses can be obtained efficiently and accurately. The results were compared with those obtained in the literature to verify the correctness. The variation of the vibration frequencies of the Timoshenko and moving mass system was researched. The effects of inertial force, centrifugal force, Coriolis force, and tangential force on a Timoshenko beam subjected to an accelerating mass were investigated.


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## Body

## 1. Introduction

Dynamic response of structures subjected to a moving force or moving mass is an important issue in engineering problems. For example, the trains have experienced great advances characterized by increasingly higher speeds and weights of vehicles. As a result, the dynamic response, as well as stresses, can be significantly higher than that before or static loads. The problem arose from the observations is a structure subjected to moving masses. Many researchers studied these problems and many studies are presented in the literatures. For examples, references [1–3] assuming the moving load to be a moving force have given some analytical solutions. Abu-Hilal [4] studied the dynamic response of a double Euler-Bernoulli beam due to a moving constant load. Fryba [5] extensively analyzed the solution of moving loads on structures. Rao [6] gave a detailed analysis of the vibration of a beam excited by a moving oscillator using a perturbation method. Zibdeh and Juma [7] presented the dynamic response of a rotating beam subjected to a random moving load using analytical and numerical methods. Akin and Mofid [8] investigated the dynamic behavior of Bernoulli-Euler beams carrying a moving mass with different boundary conditions using analytical-numerical method, and achievements by other researchers are presented in literatures [9–14].The studies mentioned are based on Bernoulli-Euler beam, while the moment of inertia and shear deformation should be taken into account when ratio of the height to span is large. References [15–20] studied the dynamic response of Timoshenko beams subject to moving force. Ross [21] studied the problem of a viscoelastic Timoshenko beam subjected to a step-loading using the Laplace transform method. Katz et al. [22] solved the dynamic response of a rotating shaft subjected to a moving load with constant velocity using the modal analysis method and an integral transformation method. Yavari et al. [23] determined the dynamic behavior of Timoshenko beams due to a moving mass using the discrete element technique (DET). Lee [24] obtained the dynamic responses of a Timoshenko beam subjected to a moving mass using the AMM.Two different kinds of spectral element method (SEM) have been developed to analyse various problems of engineering, namely, spectral element method in frequency domain (FSEM) and TSEM. The FSEM proposed by Doyle [25] is accurate and suited for simple 1D or 2D problems [26–28], while 3D problems or complex geometry are analyzed difficultly by this method and it does not work for nonlinear problems. TSEM is first used by Patera [29] to fluid dynamics; it has characteristics of spectral analysis and finite element method, thus, it is highly accurate and suited for complex geometry. This method has been successfully applied to many problems, that is, wave, fluid, seismology and acoustics [30–35].In this paper, a Timoshenko beam spectral element has been developed on the basis of Legendre polynomials-based spectral finite element. The beam is discretized into a very small number of elements with ten degrees of freedom each. The equations of motion in matrix form for a Timoshenko beam due to a moving mass are derived by using the Hamilton principle. The shape functions of the vertical displacement and the bending rotation for a Timoshenko beam element are formulated by employing Lagrange interpolation supported on the GLL points. The element mass matrix, stiffness matrix, and damping matrix are obtained by GLL integration rule. The effects of inertial forces are considered by the added mass, stiffness, and damping matrix. By assembling element matrices and element nodal vectors, respectively, the global equations of motion for a Timoshenko beam subjected to a moving mass are obtained. The vibration frequencies of the Timoshenko and a moving mass system were analyzed; in addition, the effects of the various forces considering the inertia of the mass will be investigated in detail. The presented formulations can be applied to solve the eigenvalue problem and dynamic responses for a Timoshenko beam with various boundary conditions by direct integration using generalized-ω1 method.


## 2. Theory and Formulation

A uniform Timoshenko beam acted on by a moving massm with a variable speed v(t) and acceleration a(t) along the beam is shown in Figure 1. A coordinate system is assumed to be fixed in the inertial frame with the x–axis parallel to the undeformed longitudinal axis of the beam and the y–axis pointing vertically downward in reverse direction as the gravitational acceleration g.Figure 1
Schematic figure of a Timoshenko beam and a moving mass system.According to the Timoshenko’s beam theory, the deformed beam can be described by the rotation of the across-area and the shear deformation [36]. The relation of the all rotation, section rotation, and shear strain is as follows:
(1)∂y∂x=φ+γ,
where φ and γ are, respectively, the bending rotation and shear strain.As the shear strainγ≠0, the curvature of the Timoshenko beam is different from that of the Euler-Bernoulli beam and it can be expressed as follows:
(2)R=-dφdx,
where R is the curvature of the Timoshenko beam.The Timoshenko beam is discretized into a very small number of spectral elements with equal length,l each. Every beam element consists of five nodes; each node has two degrees of freedom, that is, vertical displacement y and section rotation φ. The nodes are not uniformly distributed on the element. These nodes, denoted by ξj∈[-1,1],j=1,…,k, are the GLL points which are the (k) roots of
(3)(1-ξ2)L′k-1(ξ)=0,
where Lk-1′(ξ) is the derivative of the Legendre polynomial of degree (k-1).Figure2 shows the n-th beam element on which the moving mass applies, at time t. The position x∈[0l] of the arbitrary point in the element can be transfered to the local coordinate along the axis of the beam element ξ∈[-1,1] by
(4)ξ=2lx-1,
where l is the length of the element.Figure 2
Thenth Timoshenko beam spectral element with an accelerating mass.The vertical displacementy and section rotation φ of an arbitrary point on the beam element can be expressed as follows:
(5)y=[N1N2N3N4N5]{y1ny2ny3ny4ny5n}=[N]{yn},φ=[N1N2N3N4N5]{φ1nφ2nφ3nφ4nφ5n}=[N]{φn},
where {yn} and {φn} are, respectively, the element nodal displacement and section rotation vector. The shape functions Ni which are the orthogonal shape functions that denote Lagrange interpolation supported on the k GLL points can be expressed as follows:
(6)Ni(ξ)=∏j=1,j≠ikξ-ξjξi-ξj,
where k is the number of nodes.The equation of motion for the Timoshenko beam and moving mass system can be obtained by using the Hamilton principle, expressed as follows:(7)δ∫t1t2Ldt=0,
where L is Lagrange function; it can be represented by
(8)L=T+U+W,
where T, U, and W are, respectively, the kinetic energy, potential energy, and work done by external force. δ is the variation operator; t1 and t2 are any instance. The kinetic energy of the Timoshenko beam element is given by
(9)T=12ρA∫0l(∂y(x,t)∂t)2dξ+12ρI∫0l(∂φ(x,t)∂t)2dξ=12ρA∫-11y˙(ξ,t)2l2dξ+12ρI∫-11φ˙(ξ,t)2l2dξ,
where ρ is the density of the beam, A is the cross-sectional area, I is the second order moment of area, and the superposed dot stands for differentiation with respect to time. The potential energy of the Timoshenko beam element is expressed as follows:
(10)U=12∫0lEIR2dx+12∫0lGAκ(∂y∂x-φ)2dx=12∫-11l2EI(2l)2φ′(ξ)2dξ+12∫-11l2GAκ2l(y′(ξ)-φ)2dξ,
where E is the Young’s modulus, G is the shear modulus, and κ is the shear coefficient depending on the shape of the cross-section; the prime symbol (′) denotes differentiation with respect to ξ.When the moving mass is located atx=xc, the virtual work done by the moving force is represented by
(11)δW=m∫0l(g-∂2y∂t2+2v∂2y∂x∂t+v2∂2y∂x2+a∂y∂x)c×δy(ξ)δ(x-xc)dx=m∫-11l2(+v2(2l)2y′′(ξ)+a2ly′(ξ)g-y¨(ξ)+2v2ly˙′(ξ)ccccccccccccc+v2(2l)2y′′(ξ)+a2ly′(ξ))c×δy(ξ)δ(ξ+1-(2/l)xc)l/2dξ,
where g is the gravitational acceleration. δ(ξ+1-(2/l)xc) represents the Dirac delta function at ξ=(2/l)xc-1. By introducing (5), the expressions of the differentiations with respect to local coordinate ξ and time t for y and φ in (9)–(11) can be written as (12a)y¨=[N]{y¨n},(12b)φ¨=[N]{φ¨n},(12c)y′=[N′]{yn},(12d)φ′=[N′]{φn},(12e)y′′=[N′′]{yn},(12f)y˙′=[N′]{y˙n}.By substituting (12a)–(12f) into (8) and (8) into (7), respectively, and using the Hamilton principle, the equation of motion for a Timoshenko beam element that acted on by a moving mass in matrix form can be expressed as
(13)Meq¨e+Ceq˙e+Keqe=Fe,
where Me, Ce, and Ke are the mass, damping, and stiffness matrices of the Timoshenko beam element, respectively, Fe is the equivalent nodal force vector of the element, and q¨eq˙e, and q denote the nodal acceleration, velocity, and displacement vector of the element, respectively.Where(14)q={yeφe}Me, Ce, and Ke can be expressed as(15a)Me=MTe+Mfe,(15b)Ce=Cfe,(15c)Ke=Kbe+Kse+Kf1e+Kf2e,where Mfe, Cf, Kf1e, and Kf2e are, respectively, added mass, damping, and stiffness matrix coming from the moving load; the matrices MTe, Mfe, Cfe, Kbe, Kse, Kf1e, Kf2e, and Feare matrices defined as(16a)MTe=∫-11NT[ρA00ρI]Ndξ,(16b)Mfe=m∫-11NTNδ(ξ+1-2lxc)dξ,(16c)Cfe=2mv∫-11NTN'δ(ξ+1-2lxc)dξ,(16d)Kbe=EI∫-11BbTBbdξ,(16e)Kse=GAk∫-11BsTBsdξ,(16f)Kf1e=mv2∫-11NTN''δ(ξ+1-2lxc)dξ,(16g)Kf2e=ma∫01NTN'δ(ξ+1-2lxc)dξ,(16h)Fe=∫-11mgNTδ(ξ+1-2lxc)dξin which(17a)N=[N10N20N30N40N500N10N20N30N40N5],(17b)N'=2l[N1'0N2'0N3'0N4'0N5'00N1'0N2'0N3'0N4'0N5'],(17c)N''=(2l)2×[N1''0N2''0N3''0N4''0N5''00N1''0N2''0N3''0N4''0N5''],(17d)Bb=[0-2lN1′0-2lN2′0-2lN3′0-2lN4′0-2lN5′](17e)Bs=[-2lN1′-N1′-2lN2′-N2′-2lN3′-N3′ccccc-2lN4′-N4′-2lN5′-N5′],where the superscript T denotes the transpose.The matricesMTe, Kbe, and Kse should be calculated numerically by GLL integral which is different from Gauss integral; the integral nodes are the GLL nodes which contain the boundary nodes. It can be written as follows:(18a)MTe=∫-11NT[ρA00ρI]Ndξ=∑i=1nωiNT(ξi)[ρA00ρI]N(ξi),(18b)Kbe=EI∫-11BbTBbdξ=EI∑i=1nωiBbT(ξi)Bb(ξi),(18c)Kse=GAk∫-11BsTBsdξ=GAk∑i=1nωiBsT(ξi)Bs(ξi),where EI and GA are the bending stiffness and shear stiffness, respectively; ρA and ρI are the inertia properties coefficient; weights ωi are determined numerically by
(19)ωi=2k(k+1)·1[Lk(xi)]2,i=0,1,…,k.By assembling element matrices and element nodal vectors, respectively, the global equations of motion for a Timoshenko beam subjected to a moving mass can be obtained as follows:(20)Mq¨+Cq˙+Kq=F,
where the matrices M, C, and K are the global mass, damping, and stiffness matrices, respectively, of the Timoshenko beam, the vectors q¨, q˙, and q are the global nodal acceleration, velocity, and displacement vectors, respectively, and the vector F is the equivalent nodal force. If the added matrices of the mass and the equivalent force vector are zero, undamped natural frequencies and vibration node-shape of the beam are obtained from homogenous solution of (20). For this case, (20) reduces to
(21)Mq¨+Kq=0.
The solution of (21) can be expressed as(22a)q=Aejωt,(22b)q¨=-ω2Aejωt.Introducing (22a) and (22b) into (21) gives
(23)(K-ω2M)Φejωt=0.
As Φejωt≠0, the solution of (23) exits only if the determinant of (K-ω2M) is zero.Consider the following:(24)det⁡(K-ω2M)=0.
By solving (24) and (23), the frequencies and the corresponding vectors of the dynamic system can be obtained.In order to obtain the dynamic response of the Timoshenko beam and the moving mass system, the generalized-α time integration method is used. It has more stability and lower numerical dissipation and numerical dispersion. In the generalized-α method, (20) is transformed to
(25)K*=a0M(1-αm)+a1C(1-αf)+K(1-αf),(26)F*=F[(1-αf)ti+1+αfti]-αfKqi+C[a1(1-αf)ui+α4(1-αf)u˙i+a5(1-αf)u¨i-αfu˙i]+M[a0(1-αm)qi+a2(1-αm)q˙ihhhhhhh+a3(1-αm)q¨i-αmq¨i],
where a0, a1, a2, a3, a4, a5, a6, and a7 are, respectively, the same as that in Newmark-β method; that is,
(27)a0=1βΔt2,a1=γβΔt,a2=1βΔt,a3=12β-1a4=γβ-1,a5=Δt2(γβ-2),a6=Δt(1-γ),a7=γΔt,
where αm, αf, β, and γ are the parameters which can control the precision and stability of the algorithm. In this study, the following values are used: αf=αm=1/2, β=1/4, and γ=1/2; the details of the generalized-α method can be referred to in [37].The updated equations for the displacement velocity and acceleration vectors are expressed as(28)K*qi+1=F*,q¨i+1=a0(qi+1-qi)-a2q˙i-a3q¨i,q˙i+1=q˙i+a6q¨i+a7q¨i+1.


## 3. Numerical Calculations and Analysis

To illustrate this method and verify its effectiveness and correctness, examples 1 and 2 are solved; the solutions obtained by the proposed method are compared with those obtained by other methods (including exact analytical solutions). Example 3 is solved to present the variation of the vibration frequencies of the system during the mass moves on the beam. Example 4 is solved to present the effects of inertial force, centrifugal force, Coriolis force, and tangential force on a Timoshenko beam subjected to an accelerating mass.


### 3.1. Example  1

Three types of prismatic Timoshenko beams considering the natural vibration problems are solved by the present method. The parameters of the beams are the same with different boundaries, simple-simple, clamped-free, and clamped-simple. The properties of the beam are modulus of elasticityE=2.1×1011 N/m2, modulus of rigidity G=3E/8, density ρ=7850 kg/m3, beam length L=0.4 m, the cross section in the form of a rectangle with width b=0.02 m and height h=0.08 m, and cross-sectional shape factor κ=2/3.To test the convergence and validity of the method, the first nine-order natural frequencies obtained by presented method are compared with the analytical solution presented by Huang [38]. Fifth order of the spectral element is deployed to solve the problem, and the numbers of the beam spectral elements are from 1 to 16. The natural frequencies determined by the proposed method and the exact value calculated on the basis of [38] and Ruta [39] are shown in Table 1 to Table 3.Table 1
Free vibration frequencies for simple-simple Timoshenko beam.
ω
1
ω
2
ω
3
ω
4
ω
5
ω
6
ω
7
ω
8
ω
9
n
=
4
6838.8338
23190.888
43444.615
64925.614
86738.326
108481.00
111981.29
120647.23
129920.78
n
=
8
6838.8336
23190.827
43443.498
64939.222
86711.057
108431.78
111981.29
120647.23
130004.36
n
=
12
6838.8336
23190.827
43443.493
64939.186
86710.905
108431.36
111981.29
120647.23
130003.65
n
=
16
6838.8336
23190.827
43443.493
64939.185
86710.899
108431.34
111981.29
120647.23
130003.61
Exact [38]
6838.8336
23190.827
43443.493
64939.185
86710.889
108431.34
111981.29
120647.23
130003.61Tables1, 2, and 3 show that the first nine frequencies of the Timoshenko beam are convergent by using 16 five-order Timoshenko beam spectral elements. As the excellent characteristic of the SEM that the mass matrix is diagonal because of the choice of Lagrange interpolation function supported on the GLL points in conjunction with the GLL integration rule, the values of the natural frequencies are equal to those of the exact.Table 2
Free vibration frequencies for clamped-free Timoshenko beam.
ω
1
ω
2
ω
3
ω
4
ω
5
ω
6
ω
7
ω
8
ω
9
n
=
4
2529.4927
13279.912
31045.083
50829.910
71550.753
91995.694
110969.19
119289.42
131460.96
n
=
8
2529.4927
13279.905
31044.792
50825.847
71565.121
91995.100
110976.56
119244.54
131607.53
n
=
12
2529.4927
13279.905
31044.791
50825.835
71565.050
91994.834
110976.00
119244.58
131606.55
n
=
16
2529.4927
13279.905
31044.791
50825.843
71565.047
91994.824
110975.98
119244.57
131606.52
Exact [38]
2529.4927
13279.905
31044.791
50825.843
71565.047
91994.824
110975.98
119244.57
131606.52Table 3
Free vibration frequencies for clamped-simple Timoshenko beam.
ω
1
ω
2
ω
3
ω
4
ω
5
ω
6
ω
7
ω
8
ω
9
n
=
4
9741.9478
26150.351
45547.079
66197.391
87399.691
108645.85
114294.62
128761.33
131483.10
n
=
8
9741.9469
26150.251
45545.516
66212.036
87376.814
108601.58
114295.46
128739.40
131611.56
n
=
12
9741.9469
26150.251
45545.510
66211.995
87376.651
108601.16
114295.44
128739.41
131610.66
n
=
16
9741.9469
26150.251
45545.510
66211.993
87376.644
108601.14
114295.44
128739.40
131610.63
Reference [39]
9741.9469
26150.251
45545.510
66211.993
87376.644
108601.14
114295.44
128739.40
131610.63As the proposed method has the advantage of the finite elements and the spectral method, it can analyse free vibration problem with various boundary conditions and good accuracy.


### 3.2. Example 2

For the purpose of verification, a Timoshenko beam and a moving mass system neglecting the damping effect with boundary condition simple-simple were considered. The beam is discretized to 16 elements by five-order Timoshenko beam spectral element. The dynamic response of the system is obtained by using generalized-α method with αf=αm=1/2, β=1/4, and γ=1/2. The same parameters and material properties of the Timoshenko beam as defined in Lee [24] are used in the numerical simulation, that is, L=1 m, E=2.07×1011 N/m2, G=7.76×1010 N/m2, κ=0.9, and ρ=7700 kg/m3. The cross-sectional area A of the beam is computed from the radius of gyration r0 defined by a nondimensional parameter (Rayleigh’s coefficient) β=r0π/L. The prescribed travelling speed of the moving mass, v, is similarly defined by a nondimensional parameter given by τ=v/vcr, where vcr=(π/L)EI/ρA. The acceleration of the moving mass a is similarly defined by a nondimensional parameter λ=amL/EI.The vertical displacement of the beam is normalized byyst, defining normalized displacement U=y/yst, where yst=MgL3/48EI is the beam static displacement at mid-point when a load Mg is applied at the same point for an Euler-Bernoulli beam.Figure3 shows the normalized displacements under the moving mass by λ=1, λ=0, and λ=-0.05, respectively. The numerical results for the displacements agreed very much with the reported results by AMM using ten-term assumed functions [24].Normalized displacement under the equivalent moving loads with the initial speed given byτ=0.11. β=0.03. (a) λ=1, (b) λ=0, and (c) λ=-0.05.
(a)
(b)
(c)


### 3.3. Example 3

With only the equivalent force vector that is zero, the natural frequencies of the beam and moving masses system are obtained. For this case, we rewrite (20) as
(29)Ap˙+Bp=0,
where
(30)A=[M00-K],B=[0MMC],p={q˙q}
And the eigenvalue problem relatives to (25) can be expressed as
(31)Au=λBu,vTA=λvTB,
where λ is the vibration frequency and u and v are the right and left vector of the system, respectively.The properties and parameters of the beam are the same as Example 1. Figures4, 5, 6, 7, 8, and 9 show the first two normalized frequencies of the same Timoshenko beam subjected to a moving mass with different boundary conditions. It has a constant mass with m=10kg and different initial speed of the mass with v=5, 10, 20 m/s in Figure 4 to Figure 6, while in Figure 7 to Figure 9 the initial speed is the same with v=10 m/s and the masses are m=2, 5 and 10 kg, respectively. the acceleration is a=0 m/s2. The vibration frequencies ωs normalized by the natural frequencies ω of the beam with same boundary are denoted by f defined by f=ωs/ω.The first two normalized natural frequencies of simple-simple Timoshenko beam subjected to a moving mass with different velocities.
(a)
(b)The first two normalized natural frequencies of clamped-free Timoshenko beam subjected to a moving mass with different velocities.
(a)
(b)The first two normalized natural frequencies of clamped-simple Timoshenko beam subjected to a moving mass with different velocities.
(a)
(b)The first two normalized natural frequencies of simple-simple Timoshenko beam subjected to a moving mass with different masses.
(a)
(b)The first two normalized natural frequencies of clamped-simple Timoshenko beam subjected to a moving mass with different masses.
(a)
(b)The first two normalized natural frequencies of clamped-simple Timoshenko beam subjected to a moving mass with different masses.
(a)
(b)Figure4 to Figure 9 indicate the natural frequencies of the beam and a moving mass system that are modified periodicity during the mass moving on the beam, and the number of the periods is equal to the order of the frequencies; it also shows that frequencies are not more than those of the Timoshenko beam. It can also be seen in Figure 3 to Figure 9 that the frequencies are not sensitive to the velocity of the mass, while they can be sensitive to the mass, as the velocity affects the added damping and stiffness matrices, and the added stiffness matrix is smaller than that of the beam with bending stiffness EI and shear stiffness GA, while the added mass affected with the inertia of the mass is corresponding to the inertia property coefficient ρA of the mass matrix in the motion equations.


### 3.4. Example 4

To analyse the effect of the different forces considering the inertia of the moving mass, the Timoshenko beam and a moving mass system are introduced with the same parameters as Example 2. The inertial force, Coriolis force, centrifugal force, and tangential force at the moving mass point are, respectively, shown in Figure10.Normalized force under the equivalent moving loads with the initial speed given byτ=0.5. β=0.03. (a) λ=1, (b) λ=0, and (c) λ=-1. Dashed line: inertia force; dotted dashed line: Coriolis force; dotted line: centrifugal force; solid line: tangential force.
(a)
(b)
(c)Motion equations show that all the forces are the important members of the loads considering the effect of the moving mass and cannot be neglected. Figure10 shows that the tangential force is small than the others. The effect of the loads on the dynamic response of the Timoshenko beam is shown in Figure 11.Normalized Displacement under the equivalent moving loads with the initial speed given byτ=0.5. β=0.03. (a) λ=1, (b) λ=0, and (c) λ=-1. Solid line: without neglected; dashed line: with inertia force neglected; dotted dashed line: with Coriolis force neglected; dotted line: with centrifugal force neglected; bold dotted line: with tangential force neglected.
(a)
(b)
(c)It shows that the displacements with the tangential force neglected agreed with the displacement without neglected any forces, because the property of the Timoshenko beam is that the ratio of the height to span is large; therefore, the deformation of the beam is smaller, andy′(ξ) is very small. The Coriolis force is the important role of the loads. From the formulations of the motion equations, the effect of Coriolis force is appearing in the damping matrix; hence, with Coriolis force neglected, the displacements are larger than those without any neglected forces. Thus, the effect of the Coriolis force should be considered. The effect of Centrifugal force and inertia force are appearing in the stiff and mass matrix; therefore, with Centrifugal force and inertia being neglected, both the amplitude and the phase are different from those without any neglected forces.


## 3.1. Example  1

Three types of prismatic Timoshenko beams considering the natural vibration problems are solved by the present method. The parameters of the beams are the same with different boundaries, simple-simple, clamped-free, and clamped-simple. The properties of the beam are modulus of elasticityE=2.1×1011 N/m2, modulus of rigidity G=3E/8, density ρ=7850 kg/m3, beam length L=0.4 m, the cross section in the form of a rectangle with width b=0.02 m and height h=0.08 m, and cross-sectional shape factor κ=2/3.To test the convergence and validity of the method, the first nine-order natural frequencies obtained by presented method are compared with the analytical solution presented by Huang [38]. Fifth order of the spectral element is deployed to solve the problem, and the numbers of the beam spectral elements are from 1 to 16. The natural frequencies determined by the proposed method and the exact value calculated on the basis of [38] and Ruta [39] are shown in Table 1 to Table 3.Table 1
Free vibration frequencies for simple-simple Timoshenko beam.
ω
1
ω
2
ω
3
ω
4
ω
5
ω
6
ω
7
ω
8
ω
9
n
=
4
6838.8338
23190.888
43444.615
64925.614
86738.326
108481.00
111981.29
120647.23
129920.78
n
=
8
6838.8336
23190.827
43443.498
64939.222
86711.057
108431.78
111981.29
120647.23
130004.36
n
=
12
6838.8336
23190.827
43443.493
64939.186
86710.905
108431.36
111981.29
120647.23
130003.65
n
=
16
6838.8336
23190.827
43443.493
64939.185
86710.899
108431.34
111981.29
120647.23
130003.61
Exact [38]
6838.8336
23190.827
43443.493
64939.185
86710.889
108431.34
111981.29
120647.23
130003.61Tables1, 2, and 3 show that the first nine frequencies of the Timoshenko beam are convergent by using 16 five-order Timoshenko beam spectral elements. As the excellent characteristic of the SEM that the mass matrix is diagonal because of the choice of Lagrange interpolation function supported on the GLL points in conjunction with the GLL integration rule, the values of the natural frequencies are equal to those of the exact.Table 2
Free vibration frequencies for clamped-free Timoshenko beam.
ω
1
ω
2
ω
3
ω
4
ω
5
ω
6
ω
7
ω
8
ω
9
n
=
4
2529.4927
13279.912
31045.083
50829.910
71550.753
91995.694
110969.19
119289.42
131460.96
n
=
8
2529.4927
13279.905
31044.792
50825.847
71565.121
91995.100
110976.56
119244.54
131607.53
n
=
12
2529.4927
13279.905
31044.791
50825.835
71565.050
91994.834
110976.00
119244.58
131606.55
n
=
16
2529.4927
13279.905
31044.791
50825.843
71565.047
91994.824
110975.98
119244.57
131606.52
Exact [38]
2529.4927
13279.905
31044.791
50825.843
71565.047
91994.824
110975.98
119244.57
131606.52Table 3
Free vibration frequencies for clamped-simple Timoshenko beam.
ω
1
ω
2
ω
3
ω
4
ω
5
ω
6
ω
7
ω
8
ω
9
n
=
4
9741.9478
26150.351
45547.079
66197.391
87399.691
108645.85
114294.62
128761.33
131483.10
n
=
8
9741.9469
26150.251
45545.516
66212.036
87376.814
108601.58
114295.46
128739.40
131611.56
n
=
12
9741.9469
26150.251
45545.510
66211.995
87376.651
108601.16
114295.44
128739.41
131610.66
n
=
16
9741.9469
26150.251
45545.510
66211.993
87376.644
108601.14
114295.44
128739.40
131610.63
Reference [39]
9741.9469
26150.251
45545.510
66211.993
87376.644
108601.14
114295.44
128739.40
131610.63As the proposed method has the advantage of the finite elements and the spectral method, it can analyse free vibration problem with various boundary conditions and good accuracy.


## 3.2. Example 2

For the purpose of verification, a Timoshenko beam and a moving mass system neglecting the damping effect with boundary condition simple-simple were considered. The beam is discretized to 16 elements by five-order Timoshenko beam spectral element. The dynamic response of the system is obtained by using generalized-α method with αf=αm=1/2, β=1/4, and γ=1/2. The same parameters and material properties of the Timoshenko beam as defined in Lee [24] are used in the numerical simulation, that is, L=1 m, E=2.07×1011 N/m2, G=7.76×1010 N/m2, κ=0.9, and ρ=7700 kg/m3. The cross-sectional area A of the beam is computed from the radius of gyration r0 defined by a nondimensional parameter (Rayleigh’s coefficient) β=r0π/L. The prescribed travelling speed of the moving mass, v, is similarly defined by a nondimensional parameter given by τ=v/vcr, where vcr=(π/L)EI/ρA. The acceleration of the moving mass a is similarly defined by a nondimensional parameter λ=amL/EI.The vertical displacement of the beam is normalized byyst, defining normalized displacement U=y/yst, where yst=MgL3/48EI is the beam static displacement at mid-point when a load Mg is applied at the same point for an Euler-Bernoulli beam.Figure3 shows the normalized displacements under the moving mass by λ=1, λ=0, and λ=-0.05, respectively. The numerical results for the displacements agreed very much with the reported results by AMM using ten-term assumed functions [24].Normalized displacement under the equivalent moving loads with the initial speed given byτ=0.11. β=0.03. (a) λ=1, (b) λ=0, and (c) λ=-0.05.
(a)
(b)
(c)


## 3.3. Example 3

With only the equivalent force vector that is zero, the natural frequencies of the beam and moving masses system are obtained. For this case, we rewrite (20) as
(29)Ap˙+Bp=0,
where
(30)A=[M00-K],B=[0MMC],p={q˙q}
And the eigenvalue problem relatives to (25) can be expressed as
(31)Au=λBu,vTA=λvTB,
where λ is the vibration frequency and u and v are the right and left vector of the system, respectively.The properties and parameters of the beam are the same as Example 1. Figures4, 5, 6, 7, 8, and 9 show the first two normalized frequencies of the same Timoshenko beam subjected to a moving mass with different boundary conditions. It has a constant mass with m=10kg and different initial speed of the mass with v=5, 10, 20 m/s in Figure 4 to Figure 6, while in Figure 7 to Figure 9 the initial speed is the same with v=10 m/s and the masses are m=2, 5 and 10 kg, respectively. the acceleration is a=0 m/s2. The vibration frequencies ωs normalized by the natural frequencies ω of the beam with same boundary are denoted by f defined by f=ωs/ω.The first two normalized natural frequencies of simple-simple Timoshenko beam subjected to a moving mass with different velocities.
(a)
(b)The first two normalized natural frequencies of clamped-free Timoshenko beam subjected to a moving mass with different velocities.
(a)
(b)The first two normalized natural frequencies of clamped-simple Timoshenko beam subjected to a moving mass with different velocities.
(a)
(b)The first two normalized natural frequencies of simple-simple Timoshenko beam subjected to a moving mass with different masses.
(a)
(b)The first two normalized natural frequencies of clamped-simple Timoshenko beam subjected to a moving mass with different masses.
(a)
(b)The first two normalized natural frequencies of clamped-simple Timoshenko beam subjected to a moving mass with different masses.
(a)
(b)Figure4 to Figure 9 indicate the natural frequencies of the beam and a moving mass system that are modified periodicity during the mass moving on the beam, and the number of the periods is equal to the order of the frequencies; it also shows that frequencies are not more than those of the Timoshenko beam. It can also be seen in Figure 3 to Figure 9 that the frequencies are not sensitive to the velocity of the mass, while they can be sensitive to the mass, as the velocity affects the added damping and stiffness matrices, and the added stiffness matrix is smaller than that of the beam with bending stiffness EI and shear stiffness GA, while the added mass affected with the inertia of the mass is corresponding to the inertia property coefficient ρA of the mass matrix in the motion equations.


## 3.4. Example 4

To analyse the effect of the different forces considering the inertia of the moving mass, the Timoshenko beam and a moving mass system are introduced with the same parameters as Example 2. The inertial force, Coriolis force, centrifugal force, and tangential force at the moving mass point are, respectively, shown in Figure10.Normalized force under the equivalent moving loads with the initial speed given byτ=0.5. β=0.03. (a) λ=1, (b) λ=0, and (c) λ=-1. Dashed line: inertia force; dotted dashed line: Coriolis force; dotted line: centrifugal force; solid line: tangential force.
(a)
(b)
(c)Motion equations show that all the forces are the important members of the loads considering the effect of the moving mass and cannot be neglected. Figure10 shows that the tangential force is small than the others. The effect of the loads on the dynamic response of the Timoshenko beam is shown in Figure 11.Normalized Displacement under the equivalent moving loads with the initial speed given byτ=0.5. β=0.03. (a) λ=1, (b) λ=0, and (c) λ=-1. Solid line: without neglected; dashed line: with inertia force neglected; dotted dashed line: with Coriolis force neglected; dotted line: with centrifugal force neglected; bold dotted line: with tangential force neglected.
(a)
(b)
(c)It shows that the displacements with the tangential force neglected agreed with the displacement without neglected any forces, because the property of the Timoshenko beam is that the ratio of the height to span is large; therefore, the deformation of the beam is smaller, andy′(ξ) is very small. The Coriolis force is the important role of the loads. From the formulations of the motion equations, the effect of Coriolis force is appearing in the damping matrix; hence, with Coriolis force neglected, the displacements are larger than those without any neglected forces. Thus, the effect of the Coriolis force should be considered. The effect of Centrifugal force and inertia force are appearing in the stiff and mass matrix; therefore, with Centrifugal force and inertia being neglected, both the amplitude and the phase are different from those without any neglected forces.


## 4. Conclusion

The equation of motion in matrix form has been formulated for the dynamic response of a Timoshenko beam subjected to a moving mass by using TSEM. The inertia effect of the moving mass can easily be taken into account by assembling added mass, damping, and stiffness matrices to the global mass, damping, and stiffness matrices. The TSEM has the advantages of the spectral method and finite element method; the degrees of freedom can be less with the characteristic of high accuracy; the eigenvalue and dynamic problem can be obtained efficiently with the accurate diagonal mass matrix.The variation of the vibration frequencies of the Timoshenko and a moving mass system was obtained. Numerical results for the Timoshenko beam and a moving mass system indicate that the inertia effect of the moving load cannot be neglected, the Coriolis, inertia, and centrifugal forces take a more important role than tangential force in the moving mass system, Coriolis, inertia, centrifugal, and tangential forces take the role of the damping, mass, and stiffness matrices, respectively.



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*Source: 768209-2014-02-27.xml*