Vibration and postbuckling of a functionally graded beam subjected to nonconservative forces
Qing Lu Li^{1} , Jing Hua Zhang^{2}
^{1, 2}Department of Engineering Mechanics, Lanzhou University of Technology, Lanzhou, Gansu, China
^{1}Corresponding author
Journal of Vibroengineering, Vol. 18, Issue 8, 2016, p. 49014913.
https://doi.org/10.21595/jve.2016.16824
Received 14 January 2016; received in revised form 14 July 2016; accepted 18 August 2016; published 31 December 2016
JVE Conferences
Vibration and postbuckling of beams made from functionally graded materials (FGM) subjected to uniformly and tangentially compressing follower forces are studied in this paper. Based on the accurately and geometrically nonlinear theory for extensible beams, the dynamic governing equations for FGM beams under nonconservative load are formulated. By using a shooting method to solve the nonlinearly differential equations numerically, the responses of postbuckling and free vibration in the vicinity of postbuckling configuration are obtained, in which the hingedfixed boundary conditions of beam are considered. Effects of material gradient parameter on the critical buckling, postbuckling and lower frequencies of the FGM beam are discussed in details.
Keywords: functionally graded materials, nonconservative load, postbuckling, free vibration, frequency.
1. Introduction
Follower forces are nonconservative whose lines of action are affected by the deformation of the elastic system on which they act. Bolotin [1] analyzed the stability of a clampedfree rod subjected to a concentrated compressing follower force at the free end, named Beck’s problem. Leipholz [2] studied infinitesimal stability of elastic beams subjected to uniformly and linearly continuously distributed, tangential follower forces. By using finite element method, Vitaliani et al. [3] and Detinko [4] investigated the stability problem of cantilever beams and semicircle arches with large elastic deformations under end tip follower load, the loaddisplacement diagram and loadsquare of frequency curves were present. Based on the theory of accuratly axial extension beam, Li and Zhou [5] studied the postbuckling behaviors of a homogenous beam with hingedfixed ends under uniformly distributed tangential follower forces.
The concept of functionally gradient materials (FGM) was first introduced in 1984 by a group of materials scientists in Japan [6, 7]. FGM are both macroscopically and microscopically heterogeneous composites which are normally made from a mixture of ceramics and metals with continuous composition gradation from pure ceramic on one surface to full metal on the other. This leads to gradual and smooth change in the material profile as well as the effective material properties, making them distinguish from the conventional fibermatrix composites and preferable in many fields of engineering applications. So, the studies of the mechanical behaviors of FGM structures under the mechanical and thermal loadings have being attracted more and more attentions of scientists and also have become a new research field of solid mechanics. Xiang and Yang [8] studied both free and forced vibrations of an FGM beam with variable thickness under thermally induced initial stresses based on the Timoshenko beam theory. Reddy and Chin^{}[9] investigated the dynamic thermoelastic response of functionally graded cylinders and plates. Li et al. [10] studied the thermal postbuckling of FGM Timoshenko beams with material properties changing continuously in the thickness direction by a power law function. Li et al. [11] further examined free vibration of FGM beams with surfacebonded piezoelectric layers in thermal environment. Employing the finite element method, Bhangale and Ganesan [12] carried out thermoelastic buckling and vibration analysis of a sandwich beam made from FGM. Ying et al. [13]^{}presented solutions for bending and free vibration of FGM beams resting on a WinklerPasternak elastic foundation. Aydogdu and Taskin [14] investigated the free vibration behavior of a simply supported FGM beam based on the theory of classical beam. Yang and Chen [15] studied the free vibration and elastic buckling of FGM beams with open edge cracks by using classical beam theory. The authors [16] of this paper carried out free vibration of FGM EulerBernoulli beams with postbuckling configuration subjected to axial force and presented characteristics curves of the first three lower frequencies versus the load parameters. Zhang and Zhou [17] defined a physical neutral surface that is different from the geometric midsurface of a plate and studied the bending, vibration and nonlinear bending behaviors of FGM doubly curved shallow shin shell. Ma and Lee [18] derived governing equations for both the static behavior and dynamic response of FGM beams on the physical neutral surface. Ramesh and Mohan Rao [19] discussed the natural frequencies of vibration of a rotating pretwisted functionally graded cantilever beam. Ghiasian et al. [20] studied the static and dynamic buckling of an FGM beam subjected to uniform temperature rise loading and uniform compression. Recently, Li et al. [21] studied the free vibration of FGM beams based on both the classical and firstorder shear deformation beam theories. Recently, Zhang et al. [22] present the thermal shock responses of cylindrical shell by differential quadrature solutions.
To the author’s knowledge, fewer researchers have given much attention to the static behaviors of FGM beams subjected to nonconservative forces. Especially, the dynamic response of buckled FGM beams due to follower loadings has not been appeared. Therefore, the present paper focus on the harmonic responses of vibrations of postbuckled FGM beams under distributed tangential follower force. We consider an FGM elastic beam with one end hinged and the other fixed, subjected to a uniformly distributed tangentially compressing follower forces. First, the dynamic governing equations of this problem are established on the basis of the geometrically nonlinear theory for axially extensible beams. By assuming that the amplitude of the vibrated beam is small and its response is harmonic, the above mentioned nonlinear partial differential equations are reduced to two sets of coupled ordinary differential equations. One is for the static postbuckling deformation and the other is for the linear vibration in postbuckling configurations. Finally, by using the shooting method in conjunction with an analytical continuation, the postbuckling configurations as well as the linear responses of the prebuckling and postbuckling FGM beam are obtained. The equilibrium paths as well as the postbuckling configurations of the deformed FGM beam are plotted. The characteristic relationships between frequencies and load parameter are presented numerically. The material properties are assumed to vary continuously through the thickness of the beam. Effects of material properties on the postbuckling behavior and vibration of the FGM beam are discussed in detail.
2. Basic equations
Consider a beam made from functionally graded material, with initial length $l$ and rectangular crosssection of width $b$ and height $h$, subjected to a distributed nonconservative force $\stackrel{}{q}$ along the axis of the beam as shown in Fig. 1. Take $x$coordinate along the axial line of the beam, $z$coordinate along the thickness. According to Leipholz’s point of view [2], divergence type systems, which are fairly frequent in practice, can be treated by disregarding damping.
Fig. 1. A hingedfixed functionally graded material beam subjected to uniformly distributed follower force
2.1. Material properties of FGM
It is assumed that the material properties (such as the Young’s modulus, $E$, mass density, $\rho $) of the beam vary along the height of the beam and obey the following relation [9]:
where the subscripts $c$ and $m$ denote the ceramic and metallic constituents, respectively, ${V}_{c}$ denotes the volume fraction of ceramic and follows a simple power low as:
where $n$ is the gradient index of FGM. According to this distribution, the bottom surface of the functionally graded beam is pure metal and the top surface is pure ceramic, and for different values of $n$ one can obtain different volume fractions of ceramics. Fig. 2 shows the variations of volume fraction of ceramic constituent through the thickness of beam for various values of $n$ calculated from Eq. (2). Generally, Poisson’s ratio $v$ varies in a small range. For simplicity, we assume $v$ be a constant for functionally graded materials. The effective Young’s modulus $E\left(z\right)$ and mass density $\rho \left(z\right)$ of the FGM beam follow the distribution law of Eqs. (1) and (2), namely:
where ${E}_{c}$ and ${E}_{m}$ are elastic modulus of ceramic and metal, ${\rho}_{c}$ and ${\rho}_{m}$ are mass density of ceramic and metal respectively.
Fig. 2. Variations of the volume fraction of ceramic versus the dimensionless thickness of the functionally grated material beam for different values of $n$
On the assumption that the crosssection remains plane during deforming, one can get the strain at an arbitrary point in the crosssection:
The stressstrain relation of FGM beam is:
Axis force $N\left(x\right)$ and bending moment $M\left(x\right)$ are:
in which ${A}_{1}=bh{E}_{c}{F}_{1}/2$, ${B}_{1}=b{h}^{2}{E}_{c}{F}_{2}/4$, ${D}_{1}=b{h}^{3}{E}_{c}{F}_{3}/8$, ${F}_{1}$, ${F}_{2}$ and ${F}_{3}$ are the resultant stiffness coefficients calculated by:
${f}_{E}\left(\eta ,n,{r}_{E}\right)={\left(\frac{1+\eta}{2}\right)}^{n}+{r}_{E}\left[1{\left(\frac{1+\eta}{2}\right)}^{n}\right],{r}_{E}=\frac{{E}_{m}}{{E}_{c}},\eta =\frac{2z}{h}.$
The axial force in the cross section also can be expressed as:
where $H$ and $V$ are the internal resultant forces in the longitudinal and transverse directions, respectively.
Substituting Eq. (6) into Eqs. (7)(8), obtains:
where $I=b{h}^{3}/12$.
2.2. Systems of governing differential equations
On the basis of the theory of extensible beams [23], the governing equations of the problem are written as follows:
$\frac{\partial V}{\partial x}=R{\rho}_{0}\frac{{\partial}^{2}w}{\partial {t}^{2}}+R\stackrel{}{q}\mathrm{s}\mathrm{i}\mathrm{n}\theta ,$
$\frac{\partial u}{\partial x}=R\mathrm{c}\mathrm{o}\mathrm{s}\theta 1,$
$\frac{\partial w}{\partial x}=R\mathrm{s}\mathrm{i}\mathrm{n}\theta ,$
$\frac{\partial \theta}{\partial x}=\frac{1}{{F}_{2}^{2}{F}_{1}{F}_{3}}\left[\frac{4{F}_{2}}{{E}_{c}b{h}^{2}}\left(H\mathrm{c}\mathrm{o}\mathrm{s}\theta +V\mathrm{s}\mathrm{i}\mathrm{n}\theta \right)\frac{8{F}_{1}}{{E}_{c}b{h}^{3}}M\right],$
$\frac{\partial M}{\partial x}=R\left(H\mathrm{s}\mathrm{i}\mathrm{n}\theta +V\mathrm{c}\mathrm{o}\mathrm{s}\theta \right)R{I}_{0}\frac{{\partial}^{2}\theta}{\partial {t}^{2}},$
where $t$ is the time variable, $u(x,t)$ and $v(x,t)$ are the longitudinal and transverse displacements of points on the beam axis, respectively, $M(x,t)$ is the bending moment, $\theta (x,t)$ is the angle between the deformed beam axis and the $x$axis, $R$ is the stretch of the central axis, ${\rho}_{0}$ and ${I}_{0}$ are the mass and inertia moment per unit length of the beam, defined by:
Substituting Eq. (4) into Eq. (13), one obtains:
in which ${F}_{4}$ and ${F}_{5}$ calculated by:
where:
For the convenience in the following analysis, the governing equations need to be transformed into dimensionless forms. Nondimensional parameters are introduced as follows:
$\delta =\frac{l}{h},\tau =\frac{t}{{l}^{2}}\sqrt{\frac{{E}_{c}I}{{\rho}_{0}}}.$
Thus, the nondimension governing equations can be expressed as:
$\frac{\partial {V}^{\mathrm{*}}}{\partial X}=R\frac{{\partial}^{2}W}{\partial {\tau}^{2}}+Rq\mathrm{s}\mathrm{i}\mathrm{n}{\theta}^{\mathrm{*}},$
$\frac{\partial U}{\partial X}=R\mathrm{c}\mathrm{o}\mathrm{s}{\theta}^{\mathrm{*}}1,$
$\frac{\partial W}{\partial X}=R\mathrm{s}\mathrm{i}\mathrm{n}{\theta}^{\mathrm{*}},$
$\frac{\partial {\theta}^{\mathrm{*}}}{\partial X}=\frac{1}{{F}_{2}^{2}{F}_{1}{F}_{3}}\left[\frac{{F}_{2}}{3\delta}\left({H}^{\mathrm{*}}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}^{\mathrm{*}}+{V}^{\mathrm{*}}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}^{\mathrm{*}}\right)\frac{2{F}_{1}}{3}{M}^{\mathrm{*}}\right],$
$\frac{\partial {M}^{\mathrm{*}}}{\partial X}=R\left({H}^{\mathrm{*}}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}^{\mathrm{*}}+{V}^{\mathrm{*}}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}^{\mathrm{*}}\right)+\frac{{F}_{5}}{4{\delta}^{2}{F}_{4}}\frac{{\partial}^{2}{\theta}^{\mathrm{*}}}{\partial {\tau}^{2}},$
where:
Up to now, we have arrived at the dynamics governing equations of the geometrical nonlinear deformations of FGM extensible beams subjected to a distributed nonconservative force along the axis of the beam.
3. Solutions to the problem
To analyze the free vibration of the FGM beam in the vicinity of static prebuckling and postbuckling equilibrium states, we seek solutions of Eq. (16) of the following form:
$W\left(X,\tau \right)={W}_{s}\left(X\right)+{W}_{d}\left(X,\tau \right),$
${\theta}^{*}\left(X,\tau \right)={\theta}_{s}\left(X\right)+{\theta}_{d}\left(X,\tau \right),$
${H}^{\mathrm{*}}\left(X,\tau \right)={H}_{s}\left(X\right)+{H}_{d}\left(X,\tau \right),$
${V}^{\mathrm{*}}\left(X,\tau \right)={V}_{s}\left(X\right)+{V}_{d}\left(X,\tau \right),$
${M}^{\mathrm{*}}\left(X,\tau \right)={M}_{s}\left(X\right)+{M}_{d}\left(X,\tau \right),$
where ${U}_{s}\left(X\right)\text{,}$${W}_{s}\left(X\right)\text{,}$${\theta}_{s}\left(X\right)\text{,}$${H}_{s}\left(X\right)\text{,}$${V}_{s}\left(X\right)\text{,}$${M}_{s}\left(X\right)$ indicate the solution of statically postbuckling of the beam, ${U}_{d}(X,\tau )$, ${W}_{d}(X,\tau )$, ${\theta}_{d}(X,\tau )$, ${H}_{d}(X,\tau )$, ${V}_{d}(X,\tau )$, ${M}_{d}(X,\tau )$ indicate the dynamic responses near the nonlinear static configuration for the FGM beam.
3.1. Postbuckling solutions
The governing equations of the nonlinear static problem of the FGM beam under a compressing follower loading can be obtained from Eq. (18) by neglecting the inertial terms, which are as follows:
$\frac{d{W}_{s}}{dX}={R}_{s}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s},$
$\frac{d{\theta}_{s}}{dX}=\frac{1}{{F}_{2}^{2}{F}_{1}{F}_{3}}\left[\frac{{F}_{2}}{3\delta}\left({H}_{s}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}+{V}_{s}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}\right)\frac{2{F}_{1}}{3}{M}_{s}\right],$
$\frac{d{H}_{s}}{dX}={R}_{s}q\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s},$
$\frac{d{V}_{s}}{dX}={R}_{s}q\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s},$
$\frac{d{M}_{s}}{dX}={R}_{s}\left({H}_{s}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}+{V}_{s}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}\right),$
$R=1+\frac{1}{{F}_{2}^{2}{F}_{1}{F}_{3}}\left[\frac{{F}_{3}}{6{\delta}^{2}}\left({H}_{s}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}+{V}_{s}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}\right)\frac{{F}_{2}}{3\delta}{M}_{s}\right].$
The boundary conditions of the FGM beam corresponding to the postbuckling, in dimensionless form, are shown in Fig. 1 can be expressed as:
${U}_{s}={W}_{s}={\theta}_{s}=0,\mathrm{}\mathrm{}\mathrm{}\mathrm{a}\mathrm{t}\mathrm{}\mathrm{}\mathrm{}X=1.\mathrm{}$
The last one of Eqs. (20a) is a normalized condition prescribed for the beam, and $\beta $ is the rotational angle of the crosssection of the left end.
Due to the strong nonlinearity and the coupling including in Eq. (19), it is very difficult to obtain any analytical solution of this problem. Herein, a shooting method is employed to numerically solve the problem. First, we transform the twopoint boundary value problem into an initial value problem with some unknown initial parameters. Then, the unknown parameters are estimated to start computations by the fourth order RungeKutta method, and these estimates are modified by the NewtonRaphson method until specified boundary conditions at the terminal point are satisfied. The details about this numerical approach can be found in the literatures, such as works by Li and Zhou [23]; William et al. [24]; Ma and Wang [25].
Based on the bifurcation theory, the minimum eigenvalue of the linearized problem $(\beta \to 0)$ of the nonlinear boundary value problem is the critical load of the beam instability, denoted by ${q}_{cr}$. When $q<{q}_{cr}$ the initial equilibrium of the beam is stable, when $q>{q}_{cr}$, the beams go into postbuckled state.
In what follows, a ceramic (zirconia) and metal (Ti6AI4V) system of FGM is considered. The material parameters, Young’s modulus, Poisson’s ratio and mass density are taken from the Ref. [25] which are summarized in Table 1.
Table 1. Material parameters of ceramic, zirconia and metal Ti6AL4V in the FGM system
Materials

Young’s modulus $E$ (GPa)

Poission’s ratio $v$

Mass density ${\rho}_{0}$ (kg/m)

Zirconia

244.27

0.3

4429

Ti6AL4V

122.56

0.3

3000

In the following numerical computation for the postbuckling deformation of the FGM beam, we prescribe the slenderness $\delta =$ 30. For some specific values of $\beta $, postbuckling equilibrium configurations are presented. The postbuckling equilibrium paths of the FGM beam with different values of $(\beta ,q)$ are calculated and plotted in Fig. 3, in which the gradient index of FGM beam is $n=$ 0.5.
Fig. 3. Equilibrium configurations of the hingedfixed beam with some prescribed values of $\beta $ under follower force ($n=\text{0.5}$)
Fig. 4. Characteristic curves of $\beta $ vs. load $q$ of the hingedfixed FGM beam with different values of $n$
The typical postbuckling paths of hingedfixed FGM beams are shown in Figs. 46. Fig. 4 and Fig. 5 show the rotational angle and left end displacement vs. load parameter $q$ for different values of material constant $n$, respectively. Fig. 6 show right end bending moment vs. load parameter $q$ for different values of material constant $n$. The postbuckling behaviors of ceramic beam are in excrement agreement with those of homogenous beam [5]. It is clear that the deflection of the FGM beam is higher than that of pure ceramic beam (i.e. $n=0$). It also can be seen that the loaddeformation behaviors of FGM beam are similar than those of homogenous beams. The coordinates of intersection points of the curves with $q$axis give the dimensionless critical buckling load parameters ${q}_{cr}$. It is clear that the critical buckling load decreases with the increase of the value of $n$. This is because pure ceramic beam has higher stiffness than the functionally graded beam.
Fig. 5. Characteristic curves of $U\left(0\right)$ vs. load $q$ of the hingedfixed FGM beam with different values of $n$
Fig. 6. Characteristic curves of $m\left(1\right)$ vs. load $q$ of the hingedfixed FGM beam with different values of $n$
3.2. Small vibration solutions
Now, we pay attention to the dynamic response of FGM beam under uniformly distributed follower loading.
In this investigation, we focus on steadystate vibrations corresponding to infinitesimal deformations that are superimposed upon the static postbuckled configuration.
Substitute Eqs. (18) into Eqs. (16), using Eq. (19), letting $\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{d}={\theta}_{d}$, $\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{d}=\text{1,}$ and neglecting the nonlinear terms, we obtain the linear equations governing the small free vibration of the FGM beam as follows:
$\frac{\partial {W}_{d}}{\partial \mathrm{\xi}}={\mathrm{\Lambda}}_{s}{\theta}_{d}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}+{\mathrm{\Lambda}}_{d}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s},$
$\frac{\partial {\theta}_{d}}{\partial \mathrm{\xi}}=\frac{1}{{F}_{2}^{2}{F}_{1}{F}_{3}}\left\{\frac{{F}_{2}}{3\delta}\left[\right({V}_{d}{H}_{s}{\theta}_{d})\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}+({H}_{d}+{V}_{s}{\theta}_{d}\left)\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}\right]\frac{2{F}_{1}}{3}{M}_{d}\right\},$
$\frac{\partial {H}_{d}}{\partial \mathrm{\xi}}=\mathrm{\Lambda}\frac{{\partial}^{2}{U}_{d}}{\partial {\tau}^{2}}+{\mathrm{\Lambda}}_{d}q\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}{\mathrm{\Lambda}}_{s}q{\theta}_{d}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s},$
$\frac{\partial {V}_{d}}{\partial \mathrm{\xi}}=\mathrm{\Lambda}\frac{{\partial}^{2}{W}_{d}}{\partial {\tau}^{2}}+{\mathrm{\Lambda}}_{d}q\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}+{\mathrm{\Lambda}}_{s}q{\theta}_{d}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s},$
$\frac{\partial {M}_{d}}{\partial \mathrm{\xi}}=\mathrm{\Lambda}\left({H}^{\mathrm{*}}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}^{\mathrm{*}}+{V}^{\mathrm{*}}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}^{\mathrm{*}}\right)+\frac{{F}_{5}}{4{\delta}^{2}{F}_{4}}\frac{{\partial}^{2}{\theta}^{\mathrm{*}}}{\partial {\tau}^{2}},$
where:
3.3. Harmonic responses of the linear vibrations
We find the solution in synchronous of terms as:
where $\omega $ is the natural frequency of the beam system, and ${u}_{d}$, ${w}_{d}$, ${\mathrm{\Theta}}_{d}$, ${h}_{d}$, ${v}_{d}$, ${m}_{d}$ are shape functions. The substitution of Eq. (21) into Eq. (16) and (19) achieve the ordinary differential equations for the amplitude functions as:
$\frac{d{v}_{d}}{dX}={R}_{s}\left[{\omega}^{2}{w}_{d}+q{\mathrm{\Theta}}_{d}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}\right]+\mathrm{\Gamma}q\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s},$
$\frac{d{u}_{d}}{dX}={R}_{s}{\mathrm{\Theta}}_{d}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}+\mathrm{\Gamma}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s},$
$\frac{d{w}_{d}}{dX}={R}_{s}{\mathrm{\Theta}}_{d}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}+\mathrm{\Gamma}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s},$
$\frac{d{\Theta}_{d}}{dX}=\frac{1}{{F}_{2}^{2}{F}_{1}{F}_{3}}\left\{\frac{{F}_{2}}{3\delta}\left[({h}_{d}+{V}_{s}{\mathrm{\Theta}}_{d})\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}+({v}_{d}{H}_{s}{\mathrm{\Theta}}_{d})\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}\right]\frac{2{F}_{1}}{3}{m}_{d}\right\},$
$\frac{d{m}_{d}}{dX}={R}_{s}\left[({h}_{d}+{V}_{s}{\mathrm{\Theta}}_{d})\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}+({v}_{d}{H}_{s}{\mathrm{\Theta}}_{d})\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}\right]$
$\mathrm{}\mathrm{}\mathrm{}\mathrm{}\mathrm{}\mathrm{}\frac{{F}_{5}}{4{F}_{4}{\delta}^{2}}{\omega}^{2}{\mathrm{\Theta}}_{d}+\mathrm{\Gamma}\left({V}_{s}\mathrm{c}\mathrm{o}\mathrm{s}{\theta}_{s}{H}_{s}\mathrm{s}\mathrm{i}\mathrm{n}{\theta}_{s}\right),$
where:
For the beam with hingedfixed ends, the boundary conditions corresponding to vibration problem are as follows:
${u}_{d}={w}_{d}={\theta}_{d}=0,\mathrm{}\mathrm{}\mathrm{}\mathrm{a}\mathrm{t}\mathrm{}\mathrm{}X=1.$
If the beam is not buckled, that is ${U}_{s}={W}_{s}={\theta}_{s}={V}_{s}={M}_{s}=0$, then Eqs. (22)(25) govern the linear vibrations of prebuckled beam.
3.3.1. Numerical simulation and discussions
Here, we also use the shooting method to seek for the numerical solutions to this problem, and specify the geometric parameter $\lambda =$ 150. The two boundary value problems are solved simultaneously in the numerical computation, in order to ensure the same discrete points for both the postbuckling solution and the vibration ones are the same in the numerical integration of the fourthorder RungeKutta method.
In the special case that the beam is homogenous $(n=$ 0$)$pure ceramic beam, the first divergence load parameter obtained in this paper is 57.07, which is very close to the accurate solution [3, 5]. Fig. 7 shows the relation curves between the nonconservative load parameters and nondimensional frequency for different vibration modes.
Figs. 8 shows the characteristic curves of the first three order dimensionless natural frequencies versus the dimensionless load parameter of the FGM beam with different values of the material constant $n$ in prebuckling (dashed lines) and postbuckling (solid lines) states. As it usually appears, the firstorder frequency becomes zero when $q={q}_{cr}$. The results in Figs. 7 show that a beam with a specified value of $n$ in an unbuckled state, the first three frequencies decease with the increase of the load parameters. This decrease in frequencies with loading is attributed to the fact that the loading induced compressive stress weakens the beam stiffness. In the postbuckling domain, the first and second order frequencies increasing with the increasing $q$, but the thirdorder frequencies decrease with the increasing $q$. However, as the postbuckling deformation become more significant, the frequencies seem to tend to a constant. As expected, the fundamental frequencies approach zero at the buckling points. The material constant $n$ has obvious effects on the frequencies. With the increase in the value of $n$, the frequencies decrease in the prebuckling states but increase in the postbuckling states except for the thirdorder frequencies of the FGM beam. We also find that the loadfrequency curves are continuous but not smooth at the point of the critical load ${q}_{cr}$. Li et al. [26] reveal that this is because a bifurcation points of the corresponding equilibrium paths over which the beam goes into its secondary equilibrium state, i.e. postbuckling state.
Fig. 7. Variations of the first three frequencies $\omega $ near the nonlinear buckled configuration with load parameters $q$ for a homogeneous ceramics beam
a) The fundamental frequency
b) The secondmode frequency
c) The thirdmode frequency
4. Conclusions
The postbuckling and small free vibration in the vicinity of the postbuckling configurations of hingedfixed FGM beam subjected to distributed nonconservative forces were investigated. Based on the geometrically nonlinear theory considering the extensibility of the FGM beam, governing equations for the problem were derived, and then a shooting method technique is employed to numerically solve the nonlinear equations. The deformed configuration, postbuckling equilibrium paths and characteristic curves of the lower order frequencies during prebuckling and postbuckling vs. the load parameters were presented. Effects of material constant $n$on the critical buckling loading, the postbuckling behaviors and the frequencies of the FGM beam were discussed in detail. The following conclusions are arrived from present study.
Fig. 8. Variation of the first three frequencies$\mathrm{\omega}$ near the nonlinear buckled configuration with load parameters $q$ for a hingedfixed FGM beam for different values of material constant $n$
a) The fundamental frequency
b) The secondmode frequency
c) The thirdmode frequency
1. Postbuckling behaviors of FGM beam exhibit bifurcations under follower force. The critical buckling load decreases and the postbuckling deflection increases with increasing the volume of $n$, it is due to the increase of the volume fraction of the metal reduces the bending stiffness of the whole beam.
2. All the lower order frequencies of the FGM beam decrease monotonously with the increase of the load parameter in the prebuckling states. The fundamental frequency of the FGM beam becomes zero at the critical load. However, for the FGM beam in a postbuckling domain, the first and second order frequencies increase along with the increment of load parameters, except for the third order frequencies. The dimensionless frequency for FGM beam decreases with the increase of the gradient index $n$.
3. The numerical results show that the buckling can be reduced or delayed by adjusting the material gradient index and the distributed nonconservative force. At the same time, the adjustment of the natural frequency of the nonconservative system is realized.
Acknowledgements
QUL’s and JHZ’s work was supported by the National Natural Science Foundation of China (11262010, 11272278), the Fundamental Research Funds for the Universities of Gansu and the National Science Foundation of Gansu Province (2012RJZA028). The authors gratefully acknowledge these supports.
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Cited By
International Journal of Mechanics Research
鹏程 戚

2021
