# Dynamic Analysis of Sandwich Auxetic Honeycomb Plates Subjected to Moving Oscillator Load on Elastic Foundation

Authors: Thanh Trung Tran; Quoc Hoa Pham; Trung Nguyen-Thoi; The-Van Tran
Journal: Advances in Materials Science and Engineering
 (2020)

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## Abstract

Based on Mindlin plate theory and finite element method (FEM), dynamic response analysis of sandwich composite plates with auxetic honeycomb core resting on the elastic foundation (EF) under moving oscillator load is investigated in this work. Moving oscillator load includes spring-elastick and damper c. The EF with two coefficients was modelled by Winkler and Pasternak. The system of equations of motion of the sandwich composite plate can be solved by Newmark’s direct integration method. The reliability of the present method is verified through comparison with the results other methods available in the literature. In addition, the effects of structural parameters, material properties, and moving oscillator loads to the dynamic response of the auxetic honeycomb plate are studied.


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## Body

## 1. Introduction

The composite auxetic material with negative Poisson’s ratio is an advanced material with many outstanding benefits due to the excellent performance of ultra-light, higher stiffness, strength-to-weight ratio, and well energy absorption capacity from loads, especially the types of loads that make the structures vibrate as large as blast loads, moving loads, and impact loads, etc. Therefore, it has been used in many areas including civil and defence. With the advantageous features in many practical applications, this material has been extensively interested in the word by scientists with many different methods, including the field of mechanical calculation.Considering material with negative Poisson’s ratio, Almgren [1] presented an isotropic three-dimensional structure with equal Poisson’s ratio. Whitty et al. [2] designed sandwich composite plates with increased mechanical and thermal properties by the effect of Poisson ratios in in-plane. In [3], Ruzzene and coworkers investigated the wave propagation in sandwich structures with the cyclical auxetic core layer. Jensen and his colleagues [4] survey the effects of Poisson’s ratio on the membrane deformation of the structures under indentation. Qian-Tian and Zhi-Chun [5] studied the wave propagation in sandwich composite plates with the auxetic core. Hadjigeorgiou and Stavroulakis [6] proposed the use of auxetic materials in advanced structures. In [7–9], Teik-Cheng considered the shear deformation, auxetic foundation in plates, and thermal stresses in both plates and shell structures. Prawoto [10] studied and reviewed the structure with negative Poisson’s ratio. Strek and coworkers [11] analyzed the response of dynamic sandwich plates with auxetic cores. Scarpa and Tomlinson [12] considered the theoretical characteristics of the vibration of sandwich composite plates with in-plane negative Poisson’s ratio values. Yang and his colleagues [13] used the FEM to study the geometric effects on the micropolar elastic honeycomb structure. Imbalzano et al. [14] studied the three-dimensional modelling of auxetic sandwich composite plates for localized impact resistance and considered sandwich composite plates with auxetic core layer under blast loadings in Ref [15]. Huang and coworkers [16] investigated the in-plane mechanics of a novel zero Poisson’s ratio honeycomb core. Jin et al. [17] examined the effects of size on the mechanical behavior of hexagonal honeycombs in in-plane. Arbaoui and coworkers [18] analyzed the static bending of multilayer sandwich structures by numerical simulation and experimental methods. Yongcun and his colleagues [19] presented the buckling of a rectangular plate with auxetic material under uniaxial compression. Xu and Deng [20] considered the wave propagation characteristics in thick conventional and auxetic cellular plates. Recently, Duc and his colleagues have a relatively complete research on the mechanical behavior of the sandwich composite structures with negative Poisson’s ratio in the auxetic honeycomb core layer such as investigating dynamic nonlinear response and vibration of sandwich composite plates with auxetic honeycomb core layer by the analytical method [21]; static bending analysis of plate using FEM [22]; nonlinear forced and free vibration response of sandwich composite cylindrical panels [23]; the geometrically nonlinear dynamic response of eccentrically stiffened circular cylindrical shells [24]; double curved shallow shells using TSDT [25]; and evaluation of mechanical properties of composites reinforced by randomly distributed spherical particles with negative Poisson’s ratios [26].In the case of structures on EF, Zenkour and Radwan [27] have recently proposed an exact analytical approach to free vibration analysis of laminated composite and sandwich plates resting on EF using a plate theory with four unknown variables. Duc et al. [28] presented nonlinear thermal dynamic response of shear deformable functionally graded (FG) plates resting on EF. In [29], he and coworkers analyzed the dynamic response of composite double curved shallow shells with auxetic honeycomb core layer resting on EF under blast and damping loads. In his book [30], the results of the nonlinear dynamic analysis of FGM structures on EF are presented. Mahmoudi et al. [31] developed a refined quasi-three-dimensional shear deformation theory to analyze for FG sandwich plates resting on the two-parameter EF under thermomechanical loading.Moving load is one of the most common types of load that affects the structures in practice. Besides, the influence of this load on the dynamic response of the system is very significant. Thus, the structures under the action of the moving load are studied by many scientists around the world. Other typical works such as Song et al. [32] presented the oscillation of FG plates with a moving volume, using Kirchhoff's plate theory and analytical method based on Rayleigh–Ritz’s solution and penalty function. Malekzadeh and coworkers [33] analyzed the dynamic of laminated composite plates subjected to moving load based on the analytical method. Ouyang [34] briefly reviewed a variety of moving-load problems and several analytical solution methods. Nguyen and coworkers [35] programed triple-layer composite plates with layers connected by shear connectors subjected to moving load by FEM. Song et al. [36] investigated the dynamic of rectangular thin plates of arbitrary boundary conditions under moving loads using the analytical method. Fryba [37] has summarized a variety of engineering problems that analyzed the dynamics of structures under moving loads. The above literature reviews indicate mainly the analysis of free vibration and forced vibration when moving load or moving mass is on a straight line parallel to an edge of the plate. The research results on plate structure are influenced by the moving load with oscillation system which takes into account the interaction between structure and load as well as the investigation of any motion orbit of the load is not much.According to the best of authors’ knowledge, the researchers only focus on investigating static bending, free vibration, and the dynamics of the sandwich structures with Poisson’s ratio in auxetic honeycomb layer. Nevertheless, the dynamic response of the sandwich composite plate with honeycomb core layer subject to moving oscillator load has not been published yet. In this article, the dynamic response of plates with the auxetic honeycomb core layer subjected to moving oscillator load resting on EF is investigated numerically. The results of the present work may be applied for computation in many civil engineering problems such as vehicles running on roads, the rails resting on the EF subject to moving oscillator loads, and aeroplanes moving on runways. The sandwich composite plates used in the present study have three layers in which the top and bottom layers are isotropic materials; the core layer is auxetic honeycomb using the same material. In the present formulation, the first-order shear deformation theory (FSDT) combining with FEM is used to build the equation of motion for the sandwich auxetic honeycomb plate and to validate the results of the proposed equations with published works to verify accuracy and reliability. Finally, the influence of geometric parameters, material, load, etc., on the dynamic response of the referred plate under the moving oscillator load is investigated.


## 2. Computational Model and Assumptions

Consider plates have the lengtha, width b, and thickness h. The plate is made from materials with negative Poisson’s ratio in auxetic honeycomb subjected to the moving oscillator load consisting of two masses m1 and m2, and they are connected with spring-elastic k and viscous bonds c as shown in Figure 1.Figure 1
Model of the sandwich composite plate with negative Poisson’s ratio in auxetic honeycomb core layer subjected to moving oscillator load on EF.The moving load moves with any orbit depending on (x (t), y (t)) which is the position of mass m1.The dynamic response of the plate consists of displacement, velocity, acceleration, and stress at any point in the plate. The sandwich composite plate resting on EF is shown in Figure1. The reaction-deflection relation of Winkler–Pasternak foundation is given by [21, 27](1)qe=K1w−K2∇2w,where ∇2=∂2/∂x2+∂2/∂y2 and w is the displacement of the sandwich composite plate with honeycomb core layer; K1 and K2 are, respectively, Winkler foundation stiffness and shear layer stiffness of the Pasternak foundation.


## 3. Material Properties of the Auxetic Honeycomb Core Layer

This paper researches the sandwich composite plate with the auxetic honeycomb core layer with negative Poisson’s ratio. Unit cells of auxetic honeycomb core layer have geometric parameters as [5, 21] shown in Figure 2, in which l is the length of the inclined cell rib, h is the length of the vertical cell rib, θ is the inclined angle, and α and β define the relative cell wall length and the wall’s slenderness ratio, respectively, which are important parameters in honeycomb property. The material property of the honeycomb core layer structures is given by [5, 21](2)E1C=Etl3cosθh/l+sinθsin2θ,E2C=Etl3h/l+sinθcos3θ,ν12C=cos2θh/l+sinθsinθ,G12C=Etl3h/l+sinθh/l22h/l+1cosθ,G13C=Gtlcosθh/l+sinθ,G23C=Gtl1+2sin2θ2cosθh/l+sinθ,ρC=ρt/lh/l+22cosθh/l+sinθ,in which “c” represents core material, and E, G, and ρ are Young’s moduli, shear moduli, and mass density of the material.Figure 2
Model of the cell of the auxetic honeycomb core layer.From Figure3, it can be observed that when the angle turns to be negative values gradually, Poisson’s ratio also comes with negative values. In the range from −80° to 80°, when θ=0, value ν12 is equal to zero. However, in the case of θ=−30° and h/l = 0.5, value ν12 is undefined, and therefore, this line is discontinuous.Figure 3
Effect of cell angle on Poisson’s ratio with differenth/l ratios.


## 4. Finite Element Model for Dynamic Analysis of Sandwich Composite Plates with Negative Poisson’s Ratio in the Auxetic Honeycomb Core Layer

### 4.1. The Equation of Motion for Plate Element

The displacement field of Reddy’s FSDT is of the following form [38]:(3)ux,y,z=u0x,y,0+z⋅ψx,vx,y,z=v0x,y,0+z⋅ψy,wx,y,z=w0x,y,0,where u0,v0,w0 are the displacements at midplane (z = 0) and ψx,ψy are the transverse normal rotations of the y- and x-axes.The deformed components due to bending at the point of coordinatesx,y,z are given by(4)εb=ε0+zγ,where(5)ε0=∂u0∂x,∂v0∂y,∂u0∂y+∂v0∂xT,(6)γ=∂ψx∂x,∂ψy∂y,∂ψx∂y+∂ψy∂xT.The deformed components due to shear at points with coordinatesx,y,z are given by(7)εs=γyzγxz=ψy+∂w0∂yψx+∂w0∂x.Stress due to bending deformation at the point of coordinatesx,y,z is as follows:(8)σb=σxσyτxyT=Dbεb,in which(9)Db=Q11Q120Q12Q22000Q44.The components in equation (9) are determined by the following equation:(10)Q11=E11−ν12ν21,Q22=E21−ν12ν21,Q12=ν12E21−ν12ν21,Q44=E121+ν12.Stress due to shear strain at a pointx,y,z is given by(11)σs=τyzτxzT=Dsεs,(12)Ds=Q5500Q66.The components in equation (12) are determined as follows:(13)Q55=Q66=E121+ν12.In the present study, an eight-node isoparametric element with five degrees of freedom (DOFs) for each node is employed. The DOFs of theith node qei and plate element qe are defined as follows:(14)qei=u0iv0iw0iψxiψyiT,i=1,…,8,(15)qe=qe1qe2qe3qe4qe5qe6qe7qe8T,(16)u0=∑i=18Niξ,ηu0i,v0=∑i=18Niξ,ηv0i,w=∑i=18Niζ,ηw0i,ψx=∑i=18Niξ,ηψxi,ψy=∑i=18Niξ,ηψyi,where Ni (i = 1,…, 8) is the shape function and specified in Appendix A.The equation of motion for a plate element can be derived from the virtual work principle as follows:(17)∫Veδq˙eTρq˙edVe+∫VeδεTDbεdVe+∫VeδγTDsγdVe+∫AeδwTK1w−K2∂2w∂x2+∂2w∂y2dAe−δqeT∫AeNwptdAe=0,in which Ae and Ve are the area and volume of the element, respectively.By substituting equations (5)–(8) and (15) into equation (17), the equation of motion according to the finite element formulation for the plate element is obtained:(18)Mepq¨e+Kep+Kefw+Kefpqe=Fe,where the stiffness matrix Kep is given by(19)Kep=∫AeB1TAB1+B1TBB2+B2TBB1+B2TDB2dxdy+ξ∫AeB3TApB3dxdy,where ξ=5/6 is the shear correction factor. The gradient strain matrices in equation (19) are determined by the following formulas:(20)B1=B11B12B13B14B15B16B17B18,B2=B21B22B23B24B25B26B27B28,B3=B31B32B33B34B35B36B37B38.Note that B1i, B2i, and B3i are shown in Appendix B.Material stiffness matrices are determined by the following formulas:(21)A;B;D=∫−h/2h/21;z;z2Q11Q120Q12Q22000Q44dz,Ap=∫−h/2h/2Q5500Q66dz.Stiffness matrices of elastic foundation:(i)
For Galerkin foundation:(22)Kefw=K1∫AeNwTNwdxdy.(i)
For Pasternak foundation:(23)Kefp=K2∫Ae∂Nw∂xT∂Nw∂x+∂Nw∂yT∂Nw∂ydxdy,where Nw is shown in Appendix C.Mass matrix of the plate element:(24)Mep=∫AeNTρ∫−h/2h/2LTLdzNdxdy,in which(25)N=N1I5×5N2I5×5…N7I5×5N8I5×5,with I5×5 is the unit matrix rank 5 × 5, and(26)L=100z00100z00100.The load vector in the case of distributed uniform loadp (x, y) on the surface of the element is given by(27)Fe=∫AeNTpdxdy,in which(28)p=00px,y,t00T.The stiffness matrices, mass matrices, and nodal load vectors of the element are calculated by using the Gaussian integral method.


### 4.2. Formulation of the Moving Oscillator Load

Consider the plate element subjected to the moving oscillator load as shown in Figure4. At time t, the moving oscillator load is located on an element at the coordinate xt,yt.Figure 4
The plate element model under moving oscillator load.The interaction force caused by the moving load on the plate element is written as follows:(29)Rx,y,t=Pt−m1+m2g−m1w¨−m2f¨,where f represents the displacement along the z-axis of mass m2 and(30)w¨=x¨∂Nw∂x+y¨∂Nw∂y+x˙2∂2Nw∂x2+y˙2∂2Nw∂y2+2x˙y˙∂2Nw∂x∂yqe+2x˙∂Nw∂x+2y˙∂Nw∂yq˙e+Nwq¨e.Rewrite the equation in brief form:(31)w¨=Kewqe+Cewq˙e+Nwq¨e,in which the matrices are defined as follows:(32)Kew=x¨∂Nw∂x+y¨∂Nw∂y+x˙2∂2Nw∂x2+y˙2∂2Nw∂y2+2x˙y˙∂2Nw∂x∂y,(33)Cew=2x˙∂Nw∂x+2y˙∂Nw∂y.Using delta Dirac function, force is distributed by the following formula:(34)px,y,t=Rx,y,tδx−ηmδy−ξm.


### 4.3. The Equation of Motion of the Plate under Moving Oscillator Load

The differential equations of motion of the plate element on EF can be written as follows:(35)Mepq¨e+Cepq˙e+Kep+Kefw+Kefpqe=Fet.In order to find the moving oscillator loadFet, substituting equation (31) into equation (29) and then into equation (34), the following equation is obtained:(36)Fet=Pt−m1+m2gNwT−m1NwTKewqe−m1NwTCewq˙e−m1NwTNeqe−m2NwTf¨,in which Nw receives the value at the coordinate xm,ym which is the position of the moving load.By substituting equation (36) into equation (35), we obtained the following equation:(37)Mep+Memq¨e+Cep+Cemq˙e+Kep+Kefw+Kefp+Kemqe=Pt−m1+m2gNwT−m2NwTf¨,in which Cep=αMep+βKep and α,β are Rayleigh drag coefficients defined in [37].The differential equations of massm1 on moving oscillator load are as follows:(38)m2f¨+cf˙+kf−cw˙−kw=Pt,with(39)w˙=x˙∂Nw∂x+y˙∂Nw∂yq˙e+Nwqe.By substituting equation (39) into equation (38), we obtain(40)m2f¨+cf˙+kf−cx˙∂Nw∂x+y˙∂Nw∂y+kNwqe−cNwq˙e=Pt.By combining equation (37) and equation (40), we have the system of differential equations of motion of the plate element and the mass m2 of the moving load is as follows:(41)Mep+Memm1NwT0m2q¨ef¨+Cep+Cem0−cNwcq˙ef˙+Kep+Kefw+Kefp+Kem0cx˙∂Nw∂x+y˙∂Nw∂y+kNwkqef=Pt−m1+m2gNwTPt,where(42)Mem=m1NwTNw,Cem=m1NwTCew,Kem=m2NwTKew.The system of differential equations of motion of the whole system is obtained after gathering matrices and element load vectors into the matrix and the vector global according to the general algorithm of FEM. Note that when there is a moving load with oscillation system on the plate, the DOFs of the system increase by one, and the extension of element matrices when there are moving loads in equation (41) is added. Rows and columns are extended in the global matrix according to the corresponding total DOFs. They are linear differential equations, which have the coefficient depending on time. To solve these equations, we use the Newmark-beta method [37]:Step 1:
discrete the mesh and determine the element matrixStep 2:
determine the global matrix when there is no moving load by assembling element matricesStep 3:
solve a separate oscillation problem and determine natural frequenciesStep 4:
in each integral time step, determine the element matrix depending on the moving load, assembled into the global matrix (similar to Step 2)Step 5:
solve the system of equations in each time step and determineStep 6:
in each time step, determine stresses and internal forcesStep 7:
export the dynamic response results of the systemFlowchart of the Newmark-beta method is shown in [39, 40].


## 4.1. The Equation of Motion for Plate Element

The displacement field of Reddy’s FSDT is of the following form [38]:(3)ux,y,z=u0x,y,0+z⋅ψx,vx,y,z=v0x,y,0+z⋅ψy,wx,y,z=w0x,y,0,where u0,v0,w0 are the displacements at midplane (z = 0) and ψx,ψy are the transverse normal rotations of the y- and x-axes.The deformed components due to bending at the point of coordinatesx,y,z are given by(4)εb=ε0+zγ,where(5)ε0=∂u0∂x,∂v0∂y,∂u0∂y+∂v0∂xT,(6)γ=∂ψx∂x,∂ψy∂y,∂ψx∂y+∂ψy∂xT.The deformed components due to shear at points with coordinatesx,y,z are given by(7)εs=γyzγxz=ψy+∂w0∂yψx+∂w0∂x.Stress due to bending deformation at the point of coordinatesx,y,z is as follows:(8)σb=σxσyτxyT=Dbεb,in which(9)Db=Q11Q120Q12Q22000Q44.The components in equation (9) are determined by the following equation:(10)Q11=E11−ν12ν21,Q22=E21−ν12ν21,Q12=ν12E21−ν12ν21,Q44=E121+ν12.Stress due to shear strain at a pointx,y,z is given by(11)σs=τyzτxzT=Dsεs,(12)Ds=Q5500Q66.The components in equation (12) are determined as follows:(13)Q55=Q66=E121+ν12.In the present study, an eight-node isoparametric element with five degrees of freedom (DOFs) for each node is employed. The DOFs of theith node qei and plate element qe are defined as follows:(14)qei=u0iv0iw0iψxiψyiT,i=1,…,8,(15)qe=qe1qe2qe3qe4qe5qe6qe7qe8T,(16)u0=∑i=18Niξ,ηu0i,v0=∑i=18Niξ,ηv0i,w=∑i=18Niζ,ηw0i,ψx=∑i=18Niξ,ηψxi,ψy=∑i=18Niξ,ηψyi,where Ni (i = 1,…, 8) is the shape function and specified in Appendix A.The equation of motion for a plate element can be derived from the virtual work principle as follows:(17)∫Veδq˙eTρq˙edVe+∫VeδεTDbεdVe+∫VeδγTDsγdVe+∫AeδwTK1w−K2∂2w∂x2+∂2w∂y2dAe−δqeT∫AeNwptdAe=0,in which Ae and Ve are the area and volume of the element, respectively.By substituting equations (5)–(8) and (15) into equation (17), the equation of motion according to the finite element formulation for the plate element is obtained:(18)Mepq¨e+Kep+Kefw+Kefpqe=Fe,where the stiffness matrix Kep is given by(19)Kep=∫AeB1TAB1+B1TBB2+B2TBB1+B2TDB2dxdy+ξ∫AeB3TApB3dxdy,where ξ=5/6 is the shear correction factor. The gradient strain matrices in equation (19) are determined by the following formulas:(20)B1=B11B12B13B14B15B16B17B18,B2=B21B22B23B24B25B26B27B28,B3=B31B32B33B34B35B36B37B38.Note that B1i, B2i, and B3i are shown in Appendix B.Material stiffness matrices are determined by the following formulas:(21)A;B;D=∫−h/2h/21;z;z2Q11Q120Q12Q22000Q44dz,Ap=∫−h/2h/2Q5500Q66dz.Stiffness matrices of elastic foundation:(i)
For Galerkin foundation:(22)Kefw=K1∫AeNwTNwdxdy.(i)
For Pasternak foundation:(23)Kefp=K2∫Ae∂Nw∂xT∂Nw∂x+∂Nw∂yT∂Nw∂ydxdy,where Nw is shown in Appendix C.Mass matrix of the plate element:(24)Mep=∫AeNTρ∫−h/2h/2LTLdzNdxdy,in which(25)N=N1I5×5N2I5×5…N7I5×5N8I5×5,with I5×5 is the unit matrix rank 5 × 5, and(26)L=100z00100z00100.The load vector in the case of distributed uniform loadp (x, y) on the surface of the element is given by(27)Fe=∫AeNTpdxdy,in which(28)p=00px,y,t00T.The stiffness matrices, mass matrices, and nodal load vectors of the element are calculated by using the Gaussian integral method.


## 4.2. Formulation of the Moving Oscillator Load

Consider the plate element subjected to the moving oscillator load as shown in Figure4. At time t, the moving oscillator load is located on an element at the coordinate xt,yt.Figure 4
The plate element model under moving oscillator load.The interaction force caused by the moving load on the plate element is written as follows:(29)Rx,y,t=Pt−m1+m2g−m1w¨−m2f¨,where f represents the displacement along the z-axis of mass m2 and(30)w¨=x¨∂Nw∂x+y¨∂Nw∂y+x˙2∂2Nw∂x2+y˙2∂2Nw∂y2+2x˙y˙∂2Nw∂x∂yqe+2x˙∂Nw∂x+2y˙∂Nw∂yq˙e+Nwq¨e.Rewrite the equation in brief form:(31)w¨=Kewqe+Cewq˙e+Nwq¨e,in which the matrices are defined as follows:(32)Kew=x¨∂Nw∂x+y¨∂Nw∂y+x˙2∂2Nw∂x2+y˙2∂2Nw∂y2+2x˙y˙∂2Nw∂x∂y,(33)Cew=2x˙∂Nw∂x+2y˙∂Nw∂y.Using delta Dirac function, force is distributed by the following formula:(34)px,y,t=Rx,y,tδx−ηmδy−ξm.


## 4.3. The Equation of Motion of the Plate under Moving Oscillator Load

The differential equations of motion of the plate element on EF can be written as follows:(35)Mepq¨e+Cepq˙e+Kep+Kefw+Kefpqe=Fet.In order to find the moving oscillator loadFet, substituting equation (31) into equation (29) and then into equation (34), the following equation is obtained:(36)Fet=Pt−m1+m2gNwT−m1NwTKewqe−m1NwTCewq˙e−m1NwTNeqe−m2NwTf¨,in which Nw receives the value at the coordinate xm,ym which is the position of the moving load.By substituting equation (36) into equation (35), we obtained the following equation:(37)Mep+Memq¨e+Cep+Cemq˙e+Kep+Kefw+Kefp+Kemqe=Pt−m1+m2gNwT−m2NwTf¨,in which Cep=αMep+βKep and α,β are Rayleigh drag coefficients defined in [37].The differential equations of massm1 on moving oscillator load are as follows:(38)m2f¨+cf˙+kf−cw˙−kw=Pt,with(39)w˙=x˙∂Nw∂x+y˙∂Nw∂yq˙e+Nwqe.By substituting equation (39) into equation (38), we obtain(40)m2f¨+cf˙+kf−cx˙∂Nw∂x+y˙∂Nw∂y+kNwqe−cNwq˙e=Pt.By combining equation (37) and equation (40), we have the system of differential equations of motion of the plate element and the mass m2 of the moving load is as follows:(41)Mep+Memm1NwT0m2q¨ef¨+Cep+Cem0−cNwcq˙ef˙+Kep+Kefw+Kefp+Kem0cx˙∂Nw∂x+y˙∂Nw∂y+kNwkqef=Pt−m1+m2gNwTPt,where(42)Mem=m1NwTNw,Cem=m1NwTCew,Kem=m2NwTKew.The system of differential equations of motion of the whole system is obtained after gathering matrices and element load vectors into the matrix and the vector global according to the general algorithm of FEM. Note that when there is a moving load with oscillation system on the plate, the DOFs of the system increase by one, and the extension of element matrices when there are moving loads in equation (41) is added. Rows and columns are extended in the global matrix according to the corresponding total DOFs. They are linear differential equations, which have the coefficient depending on time. To solve these equations, we use the Newmark-beta method [37]:Step 1:
discrete the mesh and determine the element matrixStep 2:
determine the global matrix when there is no moving load by assembling element matricesStep 3:
solve a separate oscillation problem and determine natural frequenciesStep 4:
in each integral time step, determine the element matrix depending on the moving load, assembled into the global matrix (similar to Step 2)Step 5:
solve the system of equations in each time step and determineStep 6:
in each time step, determine stresses and internal forcesStep 7:
export the dynamic response results of the systemFlowchart of the Newmark-beta method is shown in [39, 40].


## 5. Numerical Results

### 5.1. Accuracy Study

Example 1.
Consider a rectangular plate with the simply supported-free-simply supported-free (SFSF—two short edges are simply supported and two long edges are free) boundary condition (BC). The plate has lengtha = 1 m, width b = a/2, thickness h = a/100, Young’s modulus E = 206.8 GPa, Poisson’s ratio ν=0.29, mass density ρ = 7820 kg/m3 under moving load with mass m = 2.3 kg along y = b/2 and v = 10 m/s as in [36]. The authors use the established program with the different cases of meshes: 4×4, 6×6, 8×8, 10×10, 12×12, 14×14 to examine the displacement of the central point of the plate versus time. The numerical results with different meshes are compared with analytical solutions as shown in Table 1. From this, it is seen that using the mesh 12×12, the present results are in good agreement with the previous literature [36] in both the shape and magnitude as presented in Figure 5. This proved the reliability and the convergence of our finite element formulation.Table 1
Maximum deflectionsw × 10−5 m versus mesh density.
Case of meshes
4
×
4
6
×
6
8
×
8
10
×
10
12
×
12
14
×
14
Present
5.0413
5.3868
5.5803
5.6998
5.7020
5.7020
[36]
5.7015Figure 5
Deflections of the central point of the plate versus time.Example 2.
To be more reliable, we consider a fully simple-support (SSSS) FGM plate with geometry parameters: the lengtha = 1 m, the width b = a, and the thickness h = a/20. The FGM plate is made of ceramic (ZrO2) and metal (Al) with material characteristics as shown in Table 2 and volume factor ratio k = 1. The nondimensional free vibration is computed as the following formula ω¯=ωhρm/Em. We use uniform mesh 12×12. The numerical results of the present method are compared with the analytical solutions [41] in Table 3.
From Table3, it can be observed that the results of this work are in good agreement with the results of [41], and the error of the first four frequencies is less than 2%.Table 2
Properties materials of FGM.
Materials
Young modulus (GPa)
Mass density (kg/m3)
Poisson’s ratio
ZrO2
E
c
=
200
ρ
c
=
5700
μ
c
=
0,3
Al
E
m
=
70
ρ
m
=
2707
μ
m
=
0,3Table 3
Nondimensional frequencyω¯.
ω
¯
Mode 1
Mode 2
Mode 3
Mode 4
Present
0.0161
0.1552
0.2603
2.2136
[41]
0.0158
0.1534
0.2592
2.2140
Error (%)
1.8
1, 2
0.4
0.02
Note thatError%=100×Present−41/41.


### 5.2. Free Vibration of Sandwich Composite Plates with Negative Poisson’s Ratio in Auxetic Honeycomb Core Layer

In this section, an SSSS sandwich composite plate with three layers, the top layer, the bottom layer, and the auxetic honeycomb core layer, are considered. The top layer (thicknessh1) and bottom layer (thickness h3) are isotropic aluminium material with geometrical parameters and materials properties: h1 = h3, E = 69 GPa, G = 26 GPa, ν=0.33, and ρ=2700 kg/m3. The core layer is the same material of other layers; geometrical parameters of the core layer are h2, θ, and ratio h/l and t/l. The plate rests on EF with Winkler foundation stiffness K1 (GPa/m) and shear stiffness of Pasternak K2 (GPa·m). Geometrical parameters of the sandwich composite plate are thickness h = h1 + h2 + h3, length a, and width b = a.The first natural frequencies with different values of the EF parameter and geometrical parameters of the core layer are presented in Table4. It can be seen that the influence of geometrical parameters of the core layer on free vibration of the sandwich composite plates with the auxetic honeycomb core layer does not follow any rule. Note that when the EF parameters increase, natural frequencies of plate increase. Furthermore, numerical results show the EF parameters make the plate become stiffener and Pasternak’s EF (K2) is more effective than Winkler’s foundation (K1). For example, with θ=−35°, h/l = 0.5 when K1 = 0.1 GPa/m and K2 = 0, natural frequency f1 = 224.5746 Hz, and if K1 = 0 and K2 = 0.05 GPa·m, natural frequency f1 = 284.8829 Hz. The first six mode shapes of the sandwich composite plate with honeycomb core layer are shown in Figure 6. In the case of the SSSS plate, the second and the third mode shapes are similar to each other. This phenomenon is suitable for the symmetrical plate structures under the same BC in the practice.Table 4
The first natural frequencies (Hz) of the plate with different values of the EF parameters and geometry parameters of the core layer (h = 0.1 m, h2/h1 = 1.5, h1 = h3, t/l = 0.1385, and a = b = 20h).
(K1, K2)
θ
°
h/l = 0.5
h/l = 1
h/l = 2
h/l = 4
(0, 0)
θ
=
−
10
°
150.0868
151.6094
152.1649
152.4068
θ
=
−
35
°
172.3232
149.7829
151.7055
152.1957
θ
=
−
55
°
158.6420
142.8576
150.7676
151.7532
θ
=
−
80
°
166.0869
58.71573
144.8534
148.7206
(0.1, 0)
θ
=
−
10
°
195.4327
197.4288
198.1573
198.4744
θ
=
−
35
°
224.5746
195.0387
197.5567
198.1988
θ
=
−
55
°
206.6414
185.979
196.3308
197.6211
θ
=
−
80
°
216.4098
76.28653
188.6021
193.6613
(0, 0.05)
θ
=
−
10
°
247.8029
250.3433
251.2704
251.674
θ
=
−
35
°
284.8829
247.3046
250.5073
251.3241
θ
=
−
55
°
262.0626
235.7882
248.9498
250.5905
θ
=
−
80
°
274.4996
96.61242
239.131
245.5620
(0.1, 0.05)
θ
=
−
10
°
277.6223
280.4719
281.5118
281.9645
θ
=
−
35
°
319.2125
277.0644
280.6563
281.5723
θ
=
−
55
°
293.6156
264.1513
278.9102
280.7500
θ
=
−
80
°
307.5683
108.1945
267.9027
275.1136Figure 6
The first six mode shapes of the SSSS plate withh = 0.1 m, h2/h1 = 1.5, h1 = h3, a = 20h, θ=−55°, h/l = 2, t/l = 0.1385, and K1 = K2 = 0.


### 5.3. Numerical Results of the Dynamic Analysis of Sandwich Composite Plates with Negative Poisson’s Ratio in Auxetic Honeycomb Core Layer

In this section, the authors consider an SFSF sandwich composite plate in which the core layer is auxetic honeycomb as Section5.2. All layers have the same material E = 69 GPa, G = 26 GPa, ν=0.33, and ρ=2700 kg/m3, the geometrical parameters of the core layer are h2, θ, and ratio h/l and t/l. Geometrical parameters of the sandwich composite plate are thickness h = h1 + h2 + h3 in which h1 and h3 are the thickness of the top and bottom layers, respectively, length a, and width b = a/2. The plate rests on EF with two parameters K1 and K2. The moving oscillator load (see Figure 1) with parameters m1, m2, k, and c travels on the plate with any orbit with velocity v and acceleration a.


#### 5.3.1. Influence of the Ratioh2/h1

Firstly, let us consider a rectangular plate withh = 0.1 m, a = 150h, θ=−55°, h2/l = 2, t/l = 0.1385, and K1 = K2 = 0. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m moves along y = b/2 with velocity v=15m/s. The ratio h2/h1 = 0, 1, 1.5, 2, and 2.5 (h2/h1 = 0 is case of isotropic plate). The deflection, velocity, and stress of the plate central point are shown in Figure 7, and maximum deflections, velocities, and stress are illustrated in Table 5. Besides, the weight reduction of the sandwich composite plate with honeycomb core layer compared to the isotropic plate is also shown in this table.Figure 7
Dynamic deflections of the centre point of the plate versus time for differenth2/h1 ratios. (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 5
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
h
2/h1 = 0
h
2/h1 = 1
h
2/h1 = 1.5
h
2/h1 = 2
h
2/h1 = 2.5
w (m)
0.0100
0.0099
0.0102
0.0106
0.0111
v (m/s2)
0.0300
0.0319
0.0341
0.0365
0.0390
σ
x
MPa
1.4262
1.3850
1.4129
1.4524
1.5024
Reduced mass (%)
0
27
34
40
45From Figure7 and Table 5, it can be observed that when the ratio hc/ht is increased from 0 to 2.5 (i.e., increasing the thickness of the middle layer since the plate thickness h is fixed), the deflection and the velocity of the central point of the plate are almost not changed. It is clear that the sandwich composite plates with honeycomb core layer help significantly absorb the vibration of structures and thus make the central deflection and velocity of plate reduce. In the case of the ratio h2/h1, it can be found that the mass of plate decreases significantly. Specifically, when h2/h1 = 2.5, the mass of the plate has decreased by 45% compared to the uniform plate. This is an impressive result. Therefore, to reduce the weight of the structure, we suggest using sandwich composite plates with honeycomb core layer instead of the anisotropic plates.


#### 5.3.2. Influence of the Velocityv

Secondly, we consider an SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = K2 = 0, and h2/h1 = 1.5. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m travels along y = b/2 with velocity v = 5, 10, 15, 20, and 25 m/s and acceleration a = 0. The deflection, velocity, and stress of the plate central point are shown in Figure 8, and maximum deflections, velocities, and stress are illustrated in Table 6.Figure 8
Dynamic deflections of the central point of the plate versus time for different velocitiesv (m/s). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 6
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
v = 5 m/s
v = 10 m/s
v = 15 m/s
v = 20 m/s
v = 25 m/s
w (m)
0.0073
0.0082
0.0102
0.0109
0.0110
v (m/s2)
0.0129
0.0245
0.0341
0.0407
0.0438
σ
x
MPa
0.9525
0.9027
1.4129
1.5631
1.6626With material properties and geometrical parameters as in Section5.3.1, the change in velocities is from 5 to 25 m/s. The response of the displacement, velocity, and stress of the plate has very little change. Once again, it can be seen that the advantage of honeycomb sandwich composite plates is that it has a very good absorption capacity.


#### 5.3.3. Influence of the Accelerationa

Next, let us consider a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = K2 = 0, and h2/h1 = 1.5. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m travels along y = b/2 with velocity v0 = 15 m/s and various accelerations a = 0, 5, 10, and 12.5 m/s2. The deflection, velocity, and stress of the plate central point are shown in Figure 9, and maximum deflections, velocities, and stress are illustrated in Table 7.Figure 9
Dynamic deflections of the central point of the plate versus time for different accelerationa (m/s2). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 7
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
a = 0 m/s2
a = 5 m/s2
a = 10 m/s2
a = 12.5 m/s2
w (m)
0.0102
0.0104
0.0106
0.0106
v (m/s2)
0.0334
0.0350
0.0365
0.0369
σ
x
M
P
a
1.4152
1.4766
1.5314
1.5499From Figure9 and Table 7, it can be observed that when the acceleration of the moving load is increased from 0 to 12.5 m/s2, deflection and velocity of the plate central point slightly increase. This proves that the acceleration of the moving load does not affect much the dynamic response of the sandwich composite plate with a honeycomb core layer.


#### 5.3.4. Influence ofK1

Now, we consider a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, and h2/h1 = 1.5, Winkler foundation stiffness K1 = 0, 0.1, 0.2, 0.3, 0.4 MPa/m, and K2 = 0. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m moves along y = b/2 with velocity v = 15 m/s and acceleration a = 0. The deflection, velocity, and stress of the plate central point are shown in Figure 10, and maximum deflections, velocities, and stress are illustrated in Table 8.Figure 10
Deflections of the central point of the plate versus time for different Winkler foundation stiffnessK1 (MPa/m). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 8
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
K
1 = 0.0 (MPa/m)
K
1 = 0.1 (MPa/m)
K
1 = 0.2 (MPa/m)
K
1 = 0.3 (MPa/m)
K
1 = 0.4 (MPa/m)
w (m)
0.0102
0.0008
0.0005
0.0004
0.0003
v (m/s)
0.0341
0.0050
0.0033
0.0033
0.0021
σ
x
M
P
a
1.4129
0.2197
0.1605
0.1375
0.1067Figure10 and Table 8 show the effects of Winkler foundation stiffness on the dynamic response of the sandwich plates with negative Poisson’s ratio. From these results, it can be seen that the EF makes the vibration amplitude of sandwich composite plates with negative Poisson’s ratio reduce strongly. Furthermore, note that the EF according to Winkler's model significantly influences the dynamic response of the sandwich composite plate.


#### 5.3.5. Influence ofK2

Finally, we investigate a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = 0, and shear layer stiffness of Pasternak foundation K2 = 0, 0.05, 0.10, 0.15, 0.20 MPa·m. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 N/m, and c = 2500 Ns/m moves along y = b/2 with velocity v = 15 m/s and acceleration a = 0. The ratio h2/h1 = 1.5. The deflection, velocity, and stress of the plate central point are shown in Figure 11, and maximum deflections, velocities, and stress are shown in Table 9.Figure 11
Effect of the Pasternak foundationK2 (Mpa·m) on the dynamic response of the sandwich plates with negative Poisson’s ratio. (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 9
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
K
2 = 0.0 (MPa·m)
K
2 = 0.05 (MPa·m)
K
2 = 0.1 (MPa·m)
K
2 = 0.15 (MPa·m)
K
2 = 0.2 (MPa·m)
w (m)
0.0102
0.0082
0.0067
0.0057
0.0049
v (m/s2)
0.0341
0.0294
0.0262
0.0234
0.0209
σ
x
MPa
1.4129
1.1070
0.8821
0.7139
0.5857Figure11 and Table 9 show the effects of the Pasternak foundation on the dynamic response of the sandwich composite plates with the honeycomb core layer. When K2 increases, the response of deflections, velocities, and stress of the plate central point reduces and also it has a significant effect on the dynamic response of structures.


## 5.1. Accuracy Study

Example 1.
Consider a rectangular plate with the simply supported-free-simply supported-free (SFSF—two short edges are simply supported and two long edges are free) boundary condition (BC). The plate has lengtha = 1 m, width b = a/2, thickness h = a/100, Young’s modulus E = 206.8 GPa, Poisson’s ratio ν=0.29, mass density ρ = 7820 kg/m3 under moving load with mass m = 2.3 kg along y = b/2 and v = 10 m/s as in [36]. The authors use the established program with the different cases of meshes: 4×4, 6×6, 8×8, 10×10, 12×12, 14×14 to examine the displacement of the central point of the plate versus time. The numerical results with different meshes are compared with analytical solutions as shown in Table 1. From this, it is seen that using the mesh 12×12, the present results are in good agreement with the previous literature [36] in both the shape and magnitude as presented in Figure 5. This proved the reliability and the convergence of our finite element formulation.Table 1
Maximum deflectionsw × 10−5 m versus mesh density.
Case of meshes
4
×
4
6
×
6
8
×
8
10
×
10
12
×
12
14
×
14
Present
5.0413
5.3868
5.5803
5.6998
5.7020
5.7020
[36]
5.7015Figure 5
Deflections of the central point of the plate versus time.Example 2.
To be more reliable, we consider a fully simple-support (SSSS) FGM plate with geometry parameters: the lengtha = 1 m, the width b = a, and the thickness h = a/20. The FGM plate is made of ceramic (ZrO2) and metal (Al) with material characteristics as shown in Table 2 and volume factor ratio k = 1. The nondimensional free vibration is computed as the following formula ω¯=ωhρm/Em. We use uniform mesh 12×12. The numerical results of the present method are compared with the analytical solutions [41] in Table 3.
From Table3, it can be observed that the results of this work are in good agreement with the results of [41], and the error of the first four frequencies is less than 2%.Table 2
Properties materials of FGM.
Materials
Young modulus (GPa)
Mass density (kg/m3)
Poisson’s ratio
ZrO2
E
c
=
200
ρ
c
=
5700
μ
c
=
0,3
Al
E
m
=
70
ρ
m
=
2707
μ
m
=
0,3Table 3
Nondimensional frequencyω¯.
ω
¯
Mode 1
Mode 2
Mode 3
Mode 4
Present
0.0161
0.1552
0.2603
2.2136
[41]
0.0158
0.1534
0.2592
2.2140
Error (%)
1.8
1, 2
0.4
0.02
Note thatError%=100×Present−41/41.


## 5.2. Free Vibration of Sandwich Composite Plates with Negative Poisson’s Ratio in Auxetic Honeycomb Core Layer

In this section, an SSSS sandwich composite plate with three layers, the top layer, the bottom layer, and the auxetic honeycomb core layer, are considered. The top layer (thicknessh1) and bottom layer (thickness h3) are isotropic aluminium material with geometrical parameters and materials properties: h1 = h3, E = 69 GPa, G = 26 GPa, ν=0.33, and ρ=2700 kg/m3. The core layer is the same material of other layers; geometrical parameters of the core layer are h2, θ, and ratio h/l and t/l. The plate rests on EF with Winkler foundation stiffness K1 (GPa/m) and shear stiffness of Pasternak K2 (GPa·m). Geometrical parameters of the sandwich composite plate are thickness h = h1 + h2 + h3, length a, and width b = a.The first natural frequencies with different values of the EF parameter and geometrical parameters of the core layer are presented in Table4. It can be seen that the influence of geometrical parameters of the core layer on free vibration of the sandwich composite plates with the auxetic honeycomb core layer does not follow any rule. Note that when the EF parameters increase, natural frequencies of plate increase. Furthermore, numerical results show the EF parameters make the plate become stiffener and Pasternak’s EF (K2) is more effective than Winkler’s foundation (K1). For example, with θ=−35°, h/l = 0.5 when K1 = 0.1 GPa/m and K2 = 0, natural frequency f1 = 224.5746 Hz, and if K1 = 0 and K2 = 0.05 GPa·m, natural frequency f1 = 284.8829 Hz. The first six mode shapes of the sandwich composite plate with honeycomb core layer are shown in Figure 6. In the case of the SSSS plate, the second and the third mode shapes are similar to each other. This phenomenon is suitable for the symmetrical plate structures under the same BC in the practice.Table 4
The first natural frequencies (Hz) of the plate with different values of the EF parameters and geometry parameters of the core layer (h = 0.1 m, h2/h1 = 1.5, h1 = h3, t/l = 0.1385, and a = b = 20h).
(K1, K2)
θ
°
h/l = 0.5
h/l = 1
h/l = 2
h/l = 4
(0, 0)
θ
=
−
10
°
150.0868
151.6094
152.1649
152.4068
θ
=
−
35
°
172.3232
149.7829
151.7055
152.1957
θ
=
−
55
°
158.6420
142.8576
150.7676
151.7532
θ
=
−
80
°
166.0869
58.71573
144.8534
148.7206
(0.1, 0)
θ
=
−
10
°
195.4327
197.4288
198.1573
198.4744
θ
=
−
35
°
224.5746
195.0387
197.5567
198.1988
θ
=
−
55
°
206.6414
185.979
196.3308
197.6211
θ
=
−
80
°
216.4098
76.28653
188.6021
193.6613
(0, 0.05)
θ
=
−
10
°
247.8029
250.3433
251.2704
251.674
θ
=
−
35
°
284.8829
247.3046
250.5073
251.3241
θ
=
−
55
°
262.0626
235.7882
248.9498
250.5905
θ
=
−
80
°
274.4996
96.61242
239.131
245.5620
(0.1, 0.05)
θ
=
−
10
°
277.6223
280.4719
281.5118
281.9645
θ
=
−
35
°
319.2125
277.0644
280.6563
281.5723
θ
=
−
55
°
293.6156
264.1513
278.9102
280.7500
θ
=
−
80
°
307.5683
108.1945
267.9027
275.1136Figure 6
The first six mode shapes of the SSSS plate withh = 0.1 m, h2/h1 = 1.5, h1 = h3, a = 20h, θ=−55°, h/l = 2, t/l = 0.1385, and K1 = K2 = 0.


## 5.3. Numerical Results of the Dynamic Analysis of Sandwich Composite Plates with Negative Poisson’s Ratio in Auxetic Honeycomb Core Layer

In this section, the authors consider an SFSF sandwich composite plate in which the core layer is auxetic honeycomb as Section5.2. All layers have the same material E = 69 GPa, G = 26 GPa, ν=0.33, and ρ=2700 kg/m3, the geometrical parameters of the core layer are h2, θ, and ratio h/l and t/l. Geometrical parameters of the sandwich composite plate are thickness h = h1 + h2 + h3 in which h1 and h3 are the thickness of the top and bottom layers, respectively, length a, and width b = a/2. The plate rests on EF with two parameters K1 and K2. The moving oscillator load (see Figure 1) with parameters m1, m2, k, and c travels on the plate with any orbit with velocity v and acceleration a.


### 5.3.1. Influence of the Ratioh2/h1

Firstly, let us consider a rectangular plate withh = 0.1 m, a = 150h, θ=−55°, h2/l = 2, t/l = 0.1385, and K1 = K2 = 0. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m moves along y = b/2 with velocity v=15m/s. The ratio h2/h1 = 0, 1, 1.5, 2, and 2.5 (h2/h1 = 0 is case of isotropic plate). The deflection, velocity, and stress of the plate central point are shown in Figure 7, and maximum deflections, velocities, and stress are illustrated in Table 5. Besides, the weight reduction of the sandwich composite plate with honeycomb core layer compared to the isotropic plate is also shown in this table.Figure 7
Dynamic deflections of the centre point of the plate versus time for differenth2/h1 ratios. (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 5
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
h
2/h1 = 0
h
2/h1 = 1
h
2/h1 = 1.5
h
2/h1 = 2
h
2/h1 = 2.5
w (m)
0.0100
0.0099
0.0102
0.0106
0.0111
v (m/s2)
0.0300
0.0319
0.0341
0.0365
0.0390
σ
x
MPa
1.4262
1.3850
1.4129
1.4524
1.5024
Reduced mass (%)
0
27
34
40
45From Figure7 and Table 5, it can be observed that when the ratio hc/ht is increased from 0 to 2.5 (i.e., increasing the thickness of the middle layer since the plate thickness h is fixed), the deflection and the velocity of the central point of the plate are almost not changed. It is clear that the sandwich composite plates with honeycomb core layer help significantly absorb the vibration of structures and thus make the central deflection and velocity of plate reduce. In the case of the ratio h2/h1, it can be found that the mass of plate decreases significantly. Specifically, when h2/h1 = 2.5, the mass of the plate has decreased by 45% compared to the uniform plate. This is an impressive result. Therefore, to reduce the weight of the structure, we suggest using sandwich composite plates with honeycomb core layer instead of the anisotropic plates.


### 5.3.2. Influence of the Velocityv

Secondly, we consider an SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = K2 = 0, and h2/h1 = 1.5. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m travels along y = b/2 with velocity v = 5, 10, 15, 20, and 25 m/s and acceleration a = 0. The deflection, velocity, and stress of the plate central point are shown in Figure 8, and maximum deflections, velocities, and stress are illustrated in Table 6.Figure 8
Dynamic deflections of the central point of the plate versus time for different velocitiesv (m/s). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 6
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
v = 5 m/s
v = 10 m/s
v = 15 m/s
v = 20 m/s
v = 25 m/s
w (m)
0.0073
0.0082
0.0102
0.0109
0.0110
v (m/s2)
0.0129
0.0245
0.0341
0.0407
0.0438
σ
x
MPa
0.9525
0.9027
1.4129
1.5631
1.6626With material properties and geometrical parameters as in Section5.3.1, the change in velocities is from 5 to 25 m/s. The response of the displacement, velocity, and stress of the plate has very little change. Once again, it can be seen that the advantage of honeycomb sandwich composite plates is that it has a very good absorption capacity.


### 5.3.3. Influence of the Accelerationa

Next, let us consider a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = K2 = 0, and h2/h1 = 1.5. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m travels along y = b/2 with velocity v0 = 15 m/s and various accelerations a = 0, 5, 10, and 12.5 m/s2. The deflection, velocity, and stress of the plate central point are shown in Figure 9, and maximum deflections, velocities, and stress are illustrated in Table 7.Figure 9
Dynamic deflections of the central point of the plate versus time for different accelerationa (m/s2). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 7
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
a = 0 m/s2
a = 5 m/s2
a = 10 m/s2
a = 12.5 m/s2
w (m)
0.0102
0.0104
0.0106
0.0106
v (m/s2)
0.0334
0.0350
0.0365
0.0369
σ
x
M
P
a
1.4152
1.4766
1.5314
1.5499From Figure9 and Table 7, it can be observed that when the acceleration of the moving load is increased from 0 to 12.5 m/s2, deflection and velocity of the plate central point slightly increase. This proves that the acceleration of the moving load does not affect much the dynamic response of the sandwich composite plate with a honeycomb core layer.


### 5.3.4. Influence ofK1

Now, we consider a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, and h2/h1 = 1.5, Winkler foundation stiffness K1 = 0, 0.1, 0.2, 0.3, 0.4 MPa/m, and K2 = 0. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m moves along y = b/2 with velocity v = 15 m/s and acceleration a = 0. The deflection, velocity, and stress of the plate central point are shown in Figure 10, and maximum deflections, velocities, and stress are illustrated in Table 8.Figure 10
Deflections of the central point of the plate versus time for different Winkler foundation stiffnessK1 (MPa/m). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 8
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
K
1 = 0.0 (MPa/m)
K
1 = 0.1 (MPa/m)
K
1 = 0.2 (MPa/m)
K
1 = 0.3 (MPa/m)
K
1 = 0.4 (MPa/m)
w (m)
0.0102
0.0008
0.0005
0.0004
0.0003
v (m/s)
0.0341
0.0050
0.0033
0.0033
0.0021
σ
x
M
P
a
1.4129
0.2197
0.1605
0.1375
0.1067Figure10 and Table 8 show the effects of Winkler foundation stiffness on the dynamic response of the sandwich plates with negative Poisson’s ratio. From these results, it can be seen that the EF makes the vibration amplitude of sandwich composite plates with negative Poisson’s ratio reduce strongly. Furthermore, note that the EF according to Winkler's model significantly influences the dynamic response of the sandwich composite plate.


### 5.3.5. Influence ofK2

Finally, we investigate a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = 0, and shear layer stiffness of Pasternak foundation K2 = 0, 0.05, 0.10, 0.15, 0.20 MPa·m. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 N/m, and c = 2500 Ns/m moves along y = b/2 with velocity v = 15 m/s and acceleration a = 0. The ratio h2/h1 = 1.5. The deflection, velocity, and stress of the plate central point are shown in Figure 11, and maximum deflections, velocities, and stress are shown in Table 9.Figure 11
Effect of the Pasternak foundationK2 (Mpa·m) on the dynamic response of the sandwich plates with negative Poisson’s ratio. (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 9
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
K
2 = 0.0 (MPa·m)
K
2 = 0.05 (MPa·m)
K
2 = 0.1 (MPa·m)
K
2 = 0.15 (MPa·m)
K
2 = 0.2 (MPa·m)
w (m)
0.0102
0.0082
0.0067
0.0057
0.0049
v (m/s2)
0.0341
0.0294
0.0262
0.0234
0.0209
σ
x
MPa
1.4129
1.1070
0.8821
0.7139
0.5857Figure11 and Table 9 show the effects of the Pasternak foundation on the dynamic response of the sandwich composite plates with the honeycomb core layer. When K2 increases, the response of deflections, velocities, and stress of the plate central point reduces and also it has a significant effect on the dynamic response of structures.


## 5.3.1. Influence of the Ratioh2/h1

Firstly, let us consider a rectangular plate withh = 0.1 m, a = 150h, θ=−55°, h2/l = 2, t/l = 0.1385, and K1 = K2 = 0. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m moves along y = b/2 with velocity v=15m/s. The ratio h2/h1 = 0, 1, 1.5, 2, and 2.5 (h2/h1 = 0 is case of isotropic plate). The deflection, velocity, and stress of the plate central point are shown in Figure 7, and maximum deflections, velocities, and stress are illustrated in Table 5. Besides, the weight reduction of the sandwich composite plate with honeycomb core layer compared to the isotropic plate is also shown in this table.Figure 7
Dynamic deflections of the centre point of the plate versus time for differenth2/h1 ratios. (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 5
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
h
2/h1 = 0
h
2/h1 = 1
h
2/h1 = 1.5
h
2/h1 = 2
h
2/h1 = 2.5
w (m)
0.0100
0.0099
0.0102
0.0106
0.0111
v (m/s2)
0.0300
0.0319
0.0341
0.0365
0.0390
σ
x
MPa
1.4262
1.3850
1.4129
1.4524
1.5024
Reduced mass (%)
0
27
34
40
45From Figure7 and Table 5, it can be observed that when the ratio hc/ht is increased from 0 to 2.5 (i.e., increasing the thickness of the middle layer since the plate thickness h is fixed), the deflection and the velocity of the central point of the plate are almost not changed. It is clear that the sandwich composite plates with honeycomb core layer help significantly absorb the vibration of structures and thus make the central deflection and velocity of plate reduce. In the case of the ratio h2/h1, it can be found that the mass of plate decreases significantly. Specifically, when h2/h1 = 2.5, the mass of the plate has decreased by 45% compared to the uniform plate. This is an impressive result. Therefore, to reduce the weight of the structure, we suggest using sandwich composite plates with honeycomb core layer instead of the anisotropic plates.


## 5.3.2. Influence of the Velocityv

Secondly, we consider an SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = K2 = 0, and h2/h1 = 1.5. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m travels along y = b/2 with velocity v = 5, 10, 15, 20, and 25 m/s and acceleration a = 0. The deflection, velocity, and stress of the plate central point are shown in Figure 8, and maximum deflections, velocities, and stress are illustrated in Table 6.Figure 8
Dynamic deflections of the central point of the plate versus time for different velocitiesv (m/s). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 6
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
v = 5 m/s
v = 10 m/s
v = 15 m/s
v = 20 m/s
v = 25 m/s
w (m)
0.0073
0.0082
0.0102
0.0109
0.0110
v (m/s2)
0.0129
0.0245
0.0341
0.0407
0.0438
σ
x
MPa
0.9525
0.9027
1.4129
1.5631
1.6626With material properties and geometrical parameters as in Section5.3.1, the change in velocities is from 5 to 25 m/s. The response of the displacement, velocity, and stress of the plate has very little change. Once again, it can be seen that the advantage of honeycomb sandwich composite plates is that it has a very good absorption capacity.


## 5.3.3. Influence of the Accelerationa

Next, let us consider a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = K2 = 0, and h2/h1 = 1.5. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m travels along y = b/2 with velocity v0 = 15 m/s and various accelerations a = 0, 5, 10, and 12.5 m/s2. The deflection, velocity, and stress of the plate central point are shown in Figure 9, and maximum deflections, velocities, and stress are illustrated in Table 7.Figure 9
Dynamic deflections of the central point of the plate versus time for different accelerationa (m/s2). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 7
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
a = 0 m/s2
a = 5 m/s2
a = 10 m/s2
a = 12.5 m/s2
w (m)
0.0102
0.0104
0.0106
0.0106
v (m/s2)
0.0334
0.0350
0.0365
0.0369
σ
x
M
P
a
1.4152
1.4766
1.5314
1.5499From Figure9 and Table 7, it can be observed that when the acceleration of the moving load is increased from 0 to 12.5 m/s2, deflection and velocity of the plate central point slightly increase. This proves that the acceleration of the moving load does not affect much the dynamic response of the sandwich composite plate with a honeycomb core layer.


## 5.3.4. Influence ofK1

Now, we consider a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, and h2/h1 = 1.5, Winkler foundation stiffness K1 = 0, 0.1, 0.2, 0.3, 0.4 MPa/m, and K2 = 0. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 MN/m, and c = 2500 Ns/m moves along y = b/2 with velocity v = 15 m/s and acceleration a = 0. The deflection, velocity, and stress of the plate central point are shown in Figure 10, and maximum deflections, velocities, and stress are illustrated in Table 8.Figure 10
Deflections of the central point of the plate versus time for different Winkler foundation stiffnessK1 (MPa/m). (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 8
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
K
1 = 0.0 (MPa/m)
K
1 = 0.1 (MPa/m)
K
1 = 0.2 (MPa/m)
K
1 = 0.3 (MPa/m)
K
1 = 0.4 (MPa/m)
w (m)
0.0102
0.0008
0.0005
0.0004
0.0003
v (m/s)
0.0341
0.0050
0.0033
0.0033
0.0021
σ
x
M
P
a
1.4129
0.2197
0.1605
0.1375
0.1067Figure10 and Table 8 show the effects of Winkler foundation stiffness on the dynamic response of the sandwich plates with negative Poisson’s ratio. From these results, it can be seen that the EF makes the vibration amplitude of sandwich composite plates with negative Poisson’s ratio reduce strongly. Furthermore, note that the EF according to Winkler's model significantly influences the dynamic response of the sandwich composite plate.


## 5.3.5. Influence ofK2

Finally, we investigate a SFSF rectangular plate withh = 0.1 m, a = 150h, h/l = 2, t/l = 0.1385, K1 = 0, and shear layer stiffness of Pasternak foundation K2 = 0, 0.05, 0.10, 0.15, 0.20 MPa·m. The moving oscillator load with m1 = m2 = 100 kg, k = 1.6 N/m, and c = 2500 Ns/m moves along y = b/2 with velocity v = 15 m/s and acceleration a = 0. The ratio h2/h1 = 1.5. The deflection, velocity, and stress of the plate central point are shown in Figure 11, and maximum deflections, velocities, and stress are shown in Table 9.Figure 11
Effect of the Pasternak foundationK2 (Mpa·m) on the dynamic response of the sandwich plates with negative Poisson’s ratio. (a) Deflection w versus time, (b) velocity v versus time, and (c) stress σx versus time.
(a)
(b)
(c)Table 9
Maximum deflections, velocities, and stress of the plate central point.
Maximum values
K
2 = 0.0 (MPa·m)
K
2 = 0.05 (MPa·m)
K
2 = 0.1 (MPa·m)
K
2 = 0.15 (MPa·m)
K
2 = 0.2 (MPa·m)
w (m)
0.0102
0.0082
0.0067
0.0057
0.0049
v (m/s2)
0.0341
0.0294
0.0262
0.0234
0.0209
σ
x
MPa
1.4129
1.1070
0.8821
0.7139
0.5857Figure11 and Table 9 show the effects of the Pasternak foundation on the dynamic response of the sandwich composite plates with the honeycomb core layer. When K2 increases, the response of deflections, velocities, and stress of the plate central point reduces and also it has a significant effect on the dynamic response of structures.


## 6. Conclusion

In the paper, the dynamic response of the sandwich composite plates with negative Poisson’s ratio under moving oscillator load on EF is mainly studied. Reddy’s first-order shear deformation plate theory is used to form the basic equations. By using FEM, the equation system of motion of the sandwich composite plates subjected to moving load is solved. From the numerical results, we can withdraw some of the following points:(i)
Authors built a numerical method to analyze the dynamical response of sandwich composite plates with honeycomb core layer under moving oscillator load on EF.(ii)
This investigation further clarifies the superiority of sandwich composite plates with negative Poisson’s ratio in auxetic honeycomb core layer under moving oscillator load. It can be noted that using sandwich auxetic honeycomb plates significantly reduces the mass and increases the vibration absorbability of the structures.(iii)
This study also analyzes and discusses the effects of material and geometrical properties and EF on the dynamic response of sandwich composite plates with negative Poisson’s ratio in auxetic honeycomb core layer.(iv)
The proposed method can analyze the static bending, dynamic response, and nonlinear problems of structures with the different honeycomb core layers, which can be not easily solved by analytical methods.



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*Source: 6309130-2020-05-12.xml*