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Journal of Applied Mathematics
Volume 2011, Article ID 842805, 17 pages
Research Article

An Investigation on the Nonlinear Free Vibration Analysis of Beams with Simply Supported Boundary Conditions Using Four Engineering Theories

1Young Research Club, Islamic Azad University, Sari Branch, P.O. Box 48161-194, Mazandaran Province, Sari, Iran
2School of Mechanical Engineering, Sharif University of Technology, P.O. Box 14588-89694, Tehran, Iran

Received 9 May 2011; Revised 5 September 2011; Accepted 8 September 2011

Academic Editor: C. Conca

Copyright © 2011 R. A. Jafari-Talookolaei et al. This is an open access article distributed under the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.


The objective of this study is to present a brief survey on the geometrically nonlinear free vibrations of the Bernoulli-Euler, the Rayleigh, shear, and the Timoshenko beams with simple end conditions using the Homotopy Analysis Method (HAM). Expressions for the natural frequencies, the transverse deflection, postbuckling load-deflection relation to, and critical buckling load are presented. The results of nonlinear analysis are validated with the published results, and excellent agreement is observed. The effects of some parameters, such as slender ratio, the rotary inertia, and the shear deformation, are examined as other parameters are fixed.

1. Introduction

Many structures, such as bridges, buildings, and spacecraft arms can be modeled as flexible beams. Most phenomena in these structures are essentially nonlinear and are described by nonlinear equations. Therefore, the study of nonlinear effects on vibration analysis of beam structures is of great importance for engineers working in the field of aerospace, mechanical, and civil engineering.

Although the nonlinear free vibrations of classical Bernoulli-Euler’s beams have been investigated by many researchers [15], only a few publications have been devoted to include the effects of shear deformation and rotary inertia [6, 7].

Azrar et al. [1] have developed a semianalytical approach to the nonlinear dynamic response problem based on Lagrange’s principle and the harmonic balance method. An analytical method for determining the vibration modes of geometrically nonlinear beams under various edge conditions has been presented by Qaisi [2]. Nonlinear modal analysis approach based on invariant manifold method is utilized to obtain the nonlinear normal modes of a clamped-clamped beam for large amplitude displacements by Xie et al. [3]. Singh et al. [4] have developed a spline-based differential quadrature method based on sextic cardinal spline functions. Guo and Zhong [5] have investigated nonlinear vibrations of thin beams based on sextic cardinal spline functions, a spline-based differential quadrature method.

Rao et al. [6] have studied the large-amplitude free vibrations of slender beams using the continuum and finite element methods. The large-amplitude free-vibration behavior of the beams with central point-concentrated masses has investigated by Rao et al. [7].

The main purpose of this survey is to present the analytical expressions for geometrically nonlinear vibration of Bernoulli-Euler’s, Rayleigh’s, shear, and Timoshenko’s beams with simple end conditions. First it is obtained nonlinear differential equations of motion on the basis of the four engineering beam theories. It is assumed that only the fundamental mode is excited and the governing nonlinear partial differential equation is reduced to a single nonlinear ordinary differential equation. These equations have been solved analytically in time domain using HAM. Finally, some numerical examples are shown for the nonslender and slender beams to signify the differences among the beam models.

2. Problem Formulation

2.1. Timoshenko’s Beam Model

Consider a straight Timoshenko’s beam of length 𝐿, a uniform cross-sectional area 𝐴, the mass per unit length of 𝑚, the moment of area of the cross-section 𝐼, Young's modulus 𝐸, and shear modulus 𝐺 that is subjected to an axial force of magnitude 𝑁 as shown in Figure 1. We assume that the beam is made of a homogenous and isotropic material.

Figure 1: A schematic of a simple ends’ beam.

The kinetic energy 𝑇 and the potential energy 𝑈 of the vibrating beam can be written as1𝑇=2𝐿0𝑚𝑤2,̂𝑡+̂𝑢2,̂𝑡+𝑟2𝜓2,̂𝑡1𝑑̂𝑥,𝑈=2𝐿0𝑀𝜓,𝑁𝜀1̂𝑥+𝑉𝛾+̂𝑥𝑑̂𝑥=2𝐿0𝐸𝐼𝜓2,̂𝑥+𝑘𝐴𝐺𝛾2+𝐸𝐴𝜀2𝑑̂𝑥,(2.1) where 𝛾 represents the shear angle (𝑤𝛾=,̂𝑥𝜓), 𝑤̂𝑢,, and 𝜓 are axial, transverse displacements, and the cross-section rotation due to the bending moment, ̂𝑥 is the axial coordinate of the beam, 𝑟 is the radius of gyration (𝐼/𝐴), and 𝑘 is beam’s cross-sectional shape factor. Also 𝑀, 𝑉, and 𝑁̂𝑥 are the bending moment, shear, and axial forces, respectively. Comma denotes differentiate with respect to ̂𝑥 or ̂𝑡.

The strain-displacement relation of the Von-Karman type is used: 𝜀=̂𝑢,+𝑤̂𝑥2,̂𝑥2.(2.2) Applying Hamilton's principle, the governing equations of motion and boundary conditions are obtained as follows:𝑚𝑤,̂𝑡̂𝑡𝑤𝑘𝐴𝐺,̂𝑥𝜓,̂𝑥𝐸𝐴̂𝑢,+𝑤̂𝑥2,̂𝑥2𝑤,̂𝑥,̂𝑥=0,(2.3)𝑚̂𝑢,̂𝑡̂𝑡𝐸𝐴̂𝑢,+𝑤̂𝑥2,̂𝑥2𝑤,̂𝑥,̂𝑥=0,(2.4)𝑚𝑟2𝜓,̂𝑡̂𝑡𝑤𝑘𝐴𝐺,̂𝑥𝜓(𝐸𝐼𝜓,)̂𝑥,̂𝑥=0,(2.5)(𝑀𝛿𝜓)𝐿0𝑁=0,(̂𝑥𝛿̂𝑢)𝐿0𝑁𝑤=0,(𝑉𝛿𝑤+̂𝑥,𝛿̂𝑥𝑤)𝐿0=0.(2.6) Upon neglecting the axial inertia, (2.3)-(2.4) can be combined to get the first type of differential equations which is coupled as follows: 𝑚𝑤,̂𝑡̂𝑡𝑤𝑘𝐴𝐺(,̂𝑥𝜓),+𝑁𝑤̂𝑥̂𝑥,̂𝑥̂𝑥=0,(2.7)𝑚𝑟2𝜓,̂𝑡̂𝑡𝑤𝑘𝐴𝐺,̂𝑥𝜓(𝐸𝐼𝜓,)̂𝑥,̂𝑥=0,(2.8) where the axial force 𝑁̂𝑥 is given by𝑁=̂𝑥𝑁𝐸𝐴2𝐿𝐿0𝑤2,̂𝑥𝑑̂𝑥(2.9) In the next step, (2.7)-(2.8) can be combined to arrive at the following decoupled equation of motion:𝑤𝐸𝐼,𝑤̂𝑥̂𝑥̂𝑥̂𝑥+𝑚,̂𝑡̂𝑡𝑚𝑟2+𝑚𝐸𝐼𝑤𝑘𝐴𝐺,̂𝑡̂𝑡+̂𝑥̂𝑥𝑚𝑟2𝑤𝑘𝐴𝐺,̂𝑡̂𝑡̂𝑡̂𝑡𝑤+𝑁̂𝑥,+̂𝑥̂𝑥𝑚𝑟2𝑤𝑘𝐴𝐺,̂𝑡̂𝑡̂𝑥̂𝑥𝐸𝐼𝑤𝑘𝐴𝐺,+̂𝑥̂𝑥̂𝑥̂𝑥𝑚𝑟2𝑁𝑘𝐴𝐺̂𝑥,̂𝑡̂𝑡𝑤,̂𝑥̂𝑥=0.(2.10) It may be noted that the equation of motion governing the rotation of the cross-section 𝜓 can be obtained using the same procedures to derive an equation similar to (2.10) with 𝑤 being replaced by𝜓.

2.2. Shear Beam Model

This model neglects the effect of rotary inertia in Timoshenko’s beam model. Therefore, the kinetic energy can only be rewritten as1𝑇=2𝐿0𝑚𝑤2,̂𝑡+̂𝑢2,̂𝑡𝑑̂𝑥.(2.11) Applying the same process mentioned in Timoshenko’s beam model, the decoupled equation of motion is given by𝑤𝐸𝐼,𝑤̂𝑥̂𝑥̂𝑥̂𝑥+𝑚,̂𝑡̂𝑡𝑚𝐸𝐼𝑤𝑘𝐴𝐺,̂𝑡̂𝑡+𝑁𝑤̂𝑥̂𝑥̂𝑥,̂𝑥̂𝑥𝐸𝐼𝑤𝑘𝐴𝐺,̂𝑥̂𝑥̂𝑥̂𝑥=0,(2.12) and the boundary conditions are as listed in (2.6).

2.3. Rayleigh’s Beam Model

In the Rayleigh beam model, the effect of shear deformation is neglected in Timoshenko’s beam model. Unlike the Timoshenko and Shear beam models, there is only one dependent variable for the Rayleigh beam. Therefore, the potential and kinetic energies can be rewritten as1𝑇=2𝐿0𝑚𝑤2,̂𝑡+̂𝑢2,̂𝑡+𝑟2𝑤2,̂𝑡1̂𝑥𝑑̂𝑥,𝑈=2𝐿0𝑀𝑤,+𝑁𝜀1̂𝑥̂𝑥̂𝑥𝑑̂𝑥=2𝐿0𝑤𝐸𝐼2,̂𝑥̂𝑥+𝐸𝐴𝜀2𝑑̂𝑥.(2.13) Again following the same procedure outline before, the equation of motion in this case is obtained as𝑤𝐸𝐼,𝑤̂𝑥̂𝑥̂𝑥̂𝑥+𝑚,̂𝑡̂𝑡𝑚𝑟2𝑤,̂𝑡̂𝑡+𝑁𝑤̂𝑥̂𝑥̂𝑥̂𝑥̂𝑥=0.(2.14) The general form of any kind of boundary conditions can be listed as(𝑁̂𝑥𝛿̂𝑢)𝐿0𝑁𝑤=0,(𝑉𝛿𝑤+̂𝑥,𝛿𝑤̂𝑥𝑤+𝑀𝛿,)̂𝑥𝐿0=0.(2.15)

2.4. Bernoulli-Euler’s Beam Model

In this model both effects of shear deformation and rotary inertia are neglected in the Timoshenko beam model. Like the Rayleigh beam model, there is one dependent variable in this model. Based on above assumptions, the potential and kinetic energies in this case are1𝑇=2𝐿0𝑚𝑤2,̂𝑡+̂𝑢2,̂𝑡1𝑑̂𝑥,𝑈=2𝐿0𝑀𝑤,+𝑁𝜀1̂𝑥̂𝑥̂𝑥𝑑̂𝑥=2𝐿0𝑤𝐸𝐼2,̂𝑥̂𝑥+𝐸𝐴𝜀2𝑑̂𝑥.(2.16) Similarly, the equation of motion is𝑤𝐸𝐼,𝑤̂𝑥̂𝑥̂𝑥̂𝑥+𝑚,̂𝑡̂𝑡+𝑁𝑤̂𝑥,̂𝑥̂𝑥=0,(2.17) and the boundary conditions are as listed in (2.15).

3. Method of Solution

3.1. Homotopy Analysis Method (HAM)

For convenience of the reader, we will first present a brief description of the HAM [8]. Consider the following nonlinear homogeneous differential equations: 𝑁[]𝑊(𝑡)=0,(3.1) where 𝑁 are nonlinear operators, 𝑡 denotes the independent variable, and 𝑊(𝑡) are unknown functions.

Liao constructed the so-called zero-order deformation equation as follows [8]:𝜙(1𝑞)𝐿(𝑡;𝑞)𝑊0[𝜙](𝑡)=𝑞(𝑡)𝑁(𝑡;𝑞),(3.2) where 𝑞[0,1] is an embedding parameter,(,(𝑡)) are nonzero auxiliary parameter and function; respectively, 𝐿 is an auxiliary linear operator, 𝑊0(𝑡) are initial guesses of 𝑊(𝑡), and 𝜙(𝑡;𝑞) are unknown functions. As 𝑞 increases from 0 to 1, the 𝜙(𝑡;𝑞) varies from the initial approximation to the exact solution. In other words,𝜙(𝑡;0)=𝑊0(𝑡) and𝜙(𝑡;1)=𝑊(𝑡).

Differentiating once more from (3.2) with respect to the embedding parameter 𝑞 and then setting 𝑞=0, we obtain the first-order deformation equation as𝐿𝑊1[𝜙](𝑡)=(𝑡)𝑁(𝑡;𝑞)𝑞=0,(3.3) which gives the first-order approximation for the 𝑊(𝑡) [8]. The higher-order approximation of the solution can be obtained by calculating the 𝑚th-order (𝑚>1) deformation equation which can be calculated by differentiating (3.3) 𝑚 times with respect to the 𝑞, then setting 𝑞=0, and finally dividing them by 𝑚! [8].

To solve (3.2), we employ Taylor’s expansion series for 𝜙(𝑡;𝑞) as𝜙(𝑡;𝑞)=𝜙(𝑡;0)+𝑚=11𝜕𝑚!𝑚𝜙(𝑡;𝑞)𝜕𝑞𝑚||||𝑞=0𝑞𝑚=𝑊0(𝑡)+𝑚=1𝑊𝑚(𝑡)𝑞𝑚,(3.4) where 𝑊𝑚(𝑡) is called the 𝑚th-order derivative of unknown function 𝜙(𝑡;𝑞).

3.2. Application of the HAM in Beam Vibrations
3.2.1. Timoshenko’s Beam Model

For convenience, the following nondimensional variables are used:𝑥=̂𝑥𝐿𝑤,𝑤=𝑟̂𝑡,𝑡=𝐸𝐼𝑚𝐿4.(3.5) In order to obtain the nonlinear natural frequency, we start to apply the Galerkin method on which the deflection for the beam with simply supported boundary conditions is expressed as𝑤(𝑥,𝑡)=𝑊(𝑡)sin(𝜋𝑥)(3.6) As a result (2.10) can be written as follows:𝑊,𝑡𝑡𝑡𝑡+𝜂4𝜆+𝜂2𝜋211+𝜆+𝜋𝑁44𝑊,𝑡𝑡+𝜂4𝜋2𝜋2𝜆𝜋𝑁2+𝜂2𝜆𝑊+𝜂2𝜋44𝜂2𝜆+𝜋2𝑊3+𝜋42𝑊𝑊2,𝑡+𝑊2𝑊,𝑡𝑡=0,(3.7) in which𝐿𝜂=𝑟𝐸,𝜆=𝑁𝑘𝐺,𝑁=𝐸𝐴.(3.8) The beam centroid is subjected to the following initial conditions:𝑊(0)=𝑎,𝑊,𝑡(0)=0,𝑊,𝑡𝑡(0)=0,𝑊,𝑡𝑡𝑡(0)=0.(3.9) Note that the free vibration of a system without damping is expressed by a kind of periodic function which has to satisfy the initial conditions. For our case this function simply iscos𝑛𝜔NL𝑡,𝑛=1,2,3,.(3.10) It is well known that for a Timoshenko’s beam with simple ends there are two natural frequencies. In other words, there are two sinusoidal modes of different frequencies corresponding to the same spatial mode. This is the most important difference between Timoshenko’s model and the other beam models with simple end conditions. This phenomenon is also reported by many researchers [911]. Thus, the initial guess for 𝑊(𝑡) is as follows:𝑊0(𝑡)=𝐴1𝜔cosNL1𝑡+𝐴2𝜔cosNL2𝑡.(3.11) To satisfy the initial conditions given in (3.9), we apply those conditions into (3.11) which yields the following expressions for the 𝐴1 and 𝐴2 as𝐴1=𝑎𝜔1NL1/𝜔NL22,𝐴2=𝑎𝜔1NL2/𝜔NL12.(3.12) Now we use (3.11) in conjunction with HAM to get the natural frequencies and the beam deflection. To construct the homotopy function, the auxiliary linear operator is selected as𝐿[]=𝜕𝜙(𝑡;𝑞)4𝜙(𝑡;𝑞)𝜕𝑡4+𝜔2NL1+𝜔2NL2𝜕2𝜙(𝑡;𝑞)𝜕𝑡2+𝜔NL1𝜔NL22.(3.13) For convenience and simplify, from (3.7), the nonlinear operator is defined as𝑁[]=𝜕𝜙(𝑡;𝑞)4𝜙(𝑡;𝑞)𝜕𝑡4+𝛼1𝜕2𝜙(𝑡;𝑞)𝜕𝑡2+𝛼2𝜙(𝑡;𝑞)+𝛼3𝜙(𝑡;𝑞)3+𝛼4𝜙(𝑡;𝑞)2𝜕2𝜙(𝑡;𝑞)𝜕𝑡2+𝜙(𝑡;𝑞)𝜕𝜙(𝑡;𝑞)𝜕𝑡2,(3.14) where𝛼1=𝜂4𝜆+𝜂2𝜋211+𝜆+𝜋𝑁44,𝛼2=𝜂4𝜋2𝜋2𝜆𝜋𝑁2+𝜂2𝜆,𝛼3=𝜂2𝜋44𝜂2𝜆+𝜋2,𝛼4=𝜋42.(3.15) In our case, to obtain the first-order approximation, (3.3) for the function of 𝜙(𝑡;𝑞) can be expressed as𝐿𝑊1||(𝑡)=(𝑡)𝑁[𝜙(𝑡;𝑞)]𝑞=0,𝑊1(0)=0,𝑊1,𝑡(0)=0,𝑊1,𝑡𝑡(0)=0,𝑊1,𝑡𝑡𝑡(0)=0.(3.16) Assuming =1and(𝑡)=1 and substituting (3.11) in (3.16), (3.16) becomes as𝑊1,𝑡𝑡𝑡𝑡+𝜔2NL1+𝜔2NL2𝑊1,𝑡𝑡+𝜔NL1𝜔NL22=𝐵1cos3𝜔NL1𝑡+𝐵2cos3𝜔NL2𝑡+𝐵3𝜔cosNL12𝜔NL2𝑡+𝐵4𝜔cosNL1+2𝜔NL2𝑡+𝐵5cos2𝜔NL1𝜔NL2𝑡+𝐵6cos2𝜔NL1+𝜔NL2𝑡+𝐵7𝜔cosNL1𝑡+𝐵8𝜔cosNL2𝑡,(3.17) where 𝐵’s are given in the appendix, equation (A.1).

These solutions should obey the general form of the base function. This means that the coefficients of the secular terms (cos(𝜔NL1𝑡),cos(𝜔NL2𝑡)) must be zero, that is, 𝐵7=0and𝐵8=0. This provides two algebric equations which yield the nonlinear frequencies: 𝜔4NL1𝛼1𝜔2NL1+𝛼2𝜔4NL2𝜔4NL12+34𝛼3𝑎2𝜔4NL2+2𝜔4NL112𝛼4𝑎2𝜔6NL1+𝜔4NL1𝜔2NL2+𝜔2NL1𝜔4NL2𝜔=0,4NL2𝛼1𝜔2NL2+𝛼2𝜔4NL2𝜔4NL12+34𝛼3𝑎2𝜔4NL1+2𝜔4NL212𝛼4𝑎2𝜔6NL2+𝜔4NL1𝜔2NL2+𝜔2NL1𝜔4NL2=0,(3.18) and solving (3.17), the 𝑊1(𝑡) is obtained as follows:𝑊1𝜔(𝑡)=𝐴cosNL1𝑡𝜔+𝐵cosNL2𝑡+𝐷1cos3𝜔NL1𝑡+𝐷2cos3𝜔NL2𝑡+𝐷3𝜔cosNL12𝜔NL2𝑡+𝐷4𝜔cosNL1+2𝜔NL2𝑡+𝐷5cos2𝜔NL1𝜔NL2𝑡+𝐷6cos2𝜔NL1+𝜔NL2𝑡,(3.19) where 𝐴 and 𝐵 are obtained by applying initial conditions. Coefficients 𝐴, 𝐵 and 𝐷’s have been presented in the appendix, equation (A.2).

Thus, the first-order approximation of the 𝑊(𝑡) becomes as𝑊(𝑡)=𝑊0(𝑡)+𝑊1(𝑡).(3.20)

3.2.2. Shear, Rayleigh’s, and Bernoulli-Euler’s Beam Models

The second-order differential equations of motion for either of shear (S), Rayleigh’s (R), or Bernoulli- Euler’s (B) beams using Galerkin’s method are the same and is𝑊,𝑡𝑡+𝛼𝑖1𝑊+𝛼𝑖3𝑊3=0,𝑖=S,R,B.(3.21) Under the following initial conditions𝑤(0)=𝑎,𝑤,𝑡(0)=0.(3.22) Using the same method described for Timoshenko’s beam model, we obtain the nonlinear natural frequency and the deflection of the beam as follows.

Nonlinear natural frequency: 𝜔NL=123𝛼𝑖3𝑎3+4𝛼𝑖1.(3.23)

Deflection of the beam, 𝛼𝑊(𝑡)=𝑖3𝑎332𝜔2NL𝜔+𝑎cosNL𝑡𝛼𝑖3𝑎332𝜔2NLcos3𝜔NL𝑡,(3.24) where 𝛼𝑖𝑗,(𝑖=S,R,B,𝑗=1,3) are given in the appendix equation (A.3).

4. Postbuckling Load-Deflection Relation and Critical Buckling Load

In order to obtain the postbuckling load-deflection relation one can set all time derivative terms in (3.7) and (3.21) equal to zero which yields the following.

Rayleigh’s and Bernoulli-Euler’s beam models, 𝜋𝑁=24𝜂24+𝑊2.(4.1)

Shear and Timoshenko’s beam models, 𝜋𝑁=24𝜂2𝜂2+𝜆𝜋24𝜂2+𝜂2+𝜆𝜋2𝑊2.(4.2) Neglecting the contribution of 𝑊 in the above equations, the critical buckling load can be determined as follows:

Rayleigh and Bernoulli-Euler Beam Models: 𝑁cr=𝜋𝜂2.(4.3)

Shear and Timoshenko’s beam models: 𝑁cr=𝜋2𝜂2+𝜆𝜋2.(4.4) As it is seen the postbuckling and critical buckling load predicted by Timoshenko model is the same as the one from shear model. The same trend exists between the Rayleigh and the Bernoulli-Euler beam models.

5. Results and Discussions

In all computational cases, the shear coefficient is taken as 𝑘=5/6 and the axial force 𝑁=0 unless mentioned otherwise.

In order to demonstrate the effectiveness of the HAM, the natural frequencies of a simply supported beam are compared with other existing results in literatures. To do this the nondimensional nonlinear frequencies are calculated, and the results are listed for different 𝜂 and 𝑎 in the Table 1.

Table 1: Comparison of the nondimensional nonlinear frequency (𝜔NL).

It can be observed that there is an excellent agreement between the results obtained from HAM and those reported by Rao et al. [6]. It should be noted that only the first natural frequency of a Timoshenko’s beam is reported in [6] and in a rather numerical way using FE method. It is clear that Bernoulli-Euler’s beam predicts the nonlinear natural frequency higher than the one obtained by other theories. Also the nonlinear natural frequency obtained by Timoshenko’s beam theory is lower than the one obtained by other theories.

The convergence and accuracy of the presented analytical solution is validated against variation of nondimensional amplitude ratio of the beam center versus 𝑡 using the Runge-Kutta method. Figure 2 illustrates comparison between these results. They are found to be in a very good agreement.

Figure 2: Variation of the nondimensional amplitude of the beam center versus 𝑡.

Now, after being satisfied in obtaining results given by this method, in the next step we try to investigate the effect of different parameters on the natural frequencies of four above-mentioned beam models.

The variation of the nonlinear frequency versus nondimensional amplitude for all different models and for two different slenderness ratios (𝜂) using HAM is shown in Figure 3. For thin beams and based on the three employed beam models, namely, Bernoulli-Euler’s, Rayleigh’s and shear beams, almost no difference is seen for different nondimensional amplitude. However, due to the nature of Timoshenko’s simply supported beam, there are two 𝜔NL variations. In the region 0<𝑎<5 as it is seen from Figure 3(a), the 𝜔NL1 represents a linear fashion variation whereas the 𝜔NL2 variation remains almost unchanged. In the region 𝑎>20 as it is seen from the same figure, variation of 𝜔NL1 follows a rather constant trend and 𝜔NL2 follows a linear fashion variation. For the intermediate region, that is, 5<𝑎<20 frequencies of both 𝜔NL1 and 𝜔NL2 keep changing in a rather nonlinear fashion. As can be seen from Figure 3(a), for thick beams this interval occurs in small amplitude. In other words, for thin and thick beams though in each amplitude both nonlinear frequencies (𝜔NL1,𝜔NL2) are involved in beam's response; however, one of them is mostly affected by variation of the nondimensional amplitude.

Figure 3: Nonlinear frequency versus nondimensional amplitude (- - -) Bernoulli-Euler’s, () Rayleigh’s, (– - –) shear, and (—) Timoshenko’s Beam Models.

Figure 4 shows nondimensional natural frequency ((𝜔NL1) for Timoshenko’s beam model) versus small nondimensional amplitude for different slenderness ratio and different beam models. It can be seen that the natural nonlinear frequency obtained from four different beam models yields the same value provided that the beam's slenderness ratio at least 100. This finding is also reported by [11] in which only the linear analysis was conducted.

Figure 4: Nonlinear frequency versus nondimensional amplitude for different slenderness ratios (- - -) Bernoulli-Euler’s, () Rayleigh’s, (-·-) shear, and (—) Timoshenko’s beam models.

The variations of the nonlinear frequency ratios versus nondimensional amplitude of Timoshenko beam for different slenderness ratios are depicted in Figure 5. As it is seen from these figures the point of shift from flat to nonflat and vise versa of 𝜔NL increases by increasing the beam slenderness ratio.

Figure 5: Nonlinear frequency versus nondimensional amplitude for different beam slenderness ratios (- - -) 𝜂=10, ()𝜂=20, (-·-) 𝜂=35, and (—) 𝜂=50.

Variation of the nondimensional postbuckling load (𝑁) versus nondimensional amplitude is shown in Figure 6. As it is seen the postbuckling load predicted by all four different models are the same for rather a slender beam (𝜂=50). On the other hand for the thick beams (𝜂=10), the differences between Timoshenko and/or Shear beam models with Rayleigh’s and/or Bernoulli-Euler’s beam models are more pronounced. Furthermore, for both cases the postbuckling load increases as amplitude (𝑎) increases.

Figure 6: Variation of the nondimensional buckling load versus nondimensional amplitude for different beam models and different slenderness ratios (- - -) Bernoulli-Euler’s and Rayleigh’s, (—) shear and Timoshenko’s beam models.

Variation of the critical buckling load (𝑁cr) versus slenderness ratio for the problem under investigation is shown in Figure 7. It can be observed that for 𝜂>12, that is, thin beams, all four models give the same answer for the critical buckling load.

Figure 7: Variation of the critical buckling load(𝑁cr) versus slenderness ratio for different beam models (- - -) Bernoulli-Euler’s and Rayleigh’s, (—) shear and Timoshenko’s beam models.

6. Conclusion

The natural frequencies, deflection, postbuckling, and critical buckling loads of four approximate models of Bernoulli-Euler’s, Rayleigh’s, shear, and Timoshenko’s beams having nonlinear behavior and simply supported end conditions have been discussed in the present survey in detail. A first-order approximation of HAM is used primarily to get the analytic expressions for the nonlinear frequencies and consequently the related deflection of neutral axis of such beam. Based on derived results, the followings are concluded.(1)For a wide range of cases having large amplitude vibration, HAM seems to be very efficient in handling such problems.(2)For thin and thick beams with simple end conditions, there are two nonlinear natural frequencies. Depending on the value of amplitude, only one of these two nonlinear natural frequencies is mostly affected by variation of the amplitude. There is a primary interval of amplitude in which the lower frequency varies in a linear fashion, and the higher frequency remains almost constant. Contrary to this trend, at the ending interval the frequency variations change their character; that is, the lower frequency is almost constant and the higher frequency varies in a linear fashion. Moreover, in the intermediate interval both frequencies increase in a nonlinear fashion. It is also seen that for the thick beams this intermediate range shifts to the left; that is, the primary interval shrinks further. In addition, it is shown that the point of shift from flat to nonflat part of 𝜔NL curves and vise versa increases by increasing the beam slenderness ratio.(3)The Bernoulli-Euler beam model predicts the nonlinear natural frequency higher than the other theories. Whereas, the nonlinear natural frequency obtained by Timoshenko’s beam theory is lower than the other theories.(4)It can be seen that the natural nonlinear frequency obtained from four different beam models yields the same value for the beams with slenderness ratio of at least 100.(5)The shear effect is always more dominant than the rotary effect for a given geometry and material in predicting postbuckling and critical buckling loads. Therefore, it is highly recommended to use either of the shear or the Timoshenko models for a thick beam. Nonetheless, in the postbuckling and critical buckling loads analysis and in a more general view the shear beam model conveys less complexity than the Timoshenko beam model. (6)The postbuckling and critical buckling loads predicted by all four different models are the same for rather a slender beam. On the other hand, for the thick beams the differences between Timoshenko’s and/or shear beam models with Rayleigh’s and/or Euler-Bernoulli’s beam models are more pronounced. Furthermore, for both cases the postbuckling load (𝑁) increases as amplitude increases.




The authors are grateful for the encouragement and useful comments of the reviewers.


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