Abstract
This paper is concerned with free vibration analysis of plate/shell coupled structures with two opposite edges simply supported by the method of reverberationray matrix. The equations of motion of the flat plate and the open circular cylindrical shell, respectively based on the classical thin plate theory and the Flügge thin shell theory, are introduced. Analytical solutions of the combination of a traveling wave form along the circumferential direction and a standing wave form along the axial direction are obtained. The method of reverberationray matrix is applied to derive the equation of the natural frequencies for the plate/shell coupled structures. The semianalytical natural frequencies are obtained with the employment of the golden section search algorithm. The semianalytical calculation results of three typical plate/shell coupled structures are presented and the results are compared with those obtained by the finite element method. The comparison shows that the calculation results obtained in this paper are of high accuracy and that the formulation presented in this manuscript are validated for free vibration analysis of plate/shell coupled structures.
1. Introduction
Thin plates and thin shells are extensively used in civil, mechanical and aeronautical engineering as well as in naval architecture and ocean engineering. Most of the practical engineering structures, such as the fuselages of aircraft, the ship hulls and the ocean platforms, etc., involve plate/shell coupled structures. The vibration behaviors of such coupled structures attract much attention from engineers in their practical designs. Quite a few experimental and analytical studies have been conducted on vibration analysis of plates and shells, but not so many researches are concerned with the plate/shell coupled structures.
Vibration of coupled structures with plate components has been studied by many researchers in the past decades. Vibration behaviors of folded plates are analyzed in literature [19]. Vibration characteristics of and power flow transmission through the Lshaped plate are studied in [1020]. Vibration behaviors of boxtype structures are investigated by a few researchers. Dickinson and Warburton [21] analyzed the free flexural vibrations of open and closed rectangular boxes and presented a theoretical solution using a sine series. Popplewell [22] studied the free vibration of a boxtype structure and the natural frequencies and normal modes are presented. Handa [23] analyzed the inplane vibration of boxtype structures by a finite element method. By considering the spatial properties of distributed forces in terms of their Fourier components and hypothesizing that the uniform component is dominant, Fulford and Petersson [24, 25] accounted for the spatially distributed wavefield at the connections of the builtup structures and the vibratory power for the boxlike structure supported by an infinite platelike recipient were considered. Lee and Wooh [26] presented the free vibration analysis of folded structures and box beams made of composite materials using a fournoded Lagrangian and Hermite finite element that incorporates high order transverse shear deformation and rotary inertia and the significance of the high order plate theory in analyzing folded structures is enunciated. Lin and Pan studied the vibration characteristics of a boxtype structure using the finite element method [27], and subsequently investigated the sound radiation characteristics of a boxtype structure with the employment of the finite element and boundary element methods [28]. More recently, Chen et al. [29] developed an analytical approach to investigate the vibration behaviors of a boxtype builtup structure and energy transmission through the structure.
Most of the available researches on plate/shell coupled structures are concerned with closed circular cylindrical shells with end plates. Yamada et al. [30] presented the free vibration analysis of a circular cylindrical doubleshell system closed by end plates. Schlesinger [31] investigated the transmission of elastic waves from a cylinder to an attached flat plate with the wave approach. Tso and Hansen [32] also studied the transmission of vibration waves through cylinder/plate junctions. Stanley and Ganesan [33] determined the natural frequencies of cylindrical shells with a circular plate attached at arbitrary locations for various boundary conditions using the semianalytical finite element method. Tso and Hansen [34] presented a theoretical and experimental study of the transmission of vibration through a two element structure which consists of a cylindrical shell coupled to an end plate. Wu et al. [35] analyzed the vibroacoustic coupling between a finite circular cylindrical shell closed at each end by a piece of circular plate and its enclosed cavity by using the coveringdomain method, which transforms the calculation of the scattering sound field of a complicatedshaped close cavity to that of a series of simply regularshaped close shells. Wang et al. [36] formulated a substructure approach to investigate the power flow characteristics of a platecylindrical shell system subject to both conservative and dissipative coupling conditions. Liang and Chen [37] investigated the natural frequencies and mode shapes for a conical shell with an annular end plate or a round end plate by means of the transfer matrix method. Subsequently, Liang et al. [38] extended the transfer matrix method to analyze a composite laminated conicalplate shell. Recently, much attentions are paid to vibration analysis of joined cylindrical, conical or spherical shells [3950].
This paper presents an analytical formulation for the free vibration analysis of plate/shell coupled structures with two opposite edges simply supported. Firstly, the force and moment resultants in a thin plate and in an open circular cylindrical shell (OCCS) are presented and the equations of motion of the thin plate and the OCCS, respectively based on the classical thin plate theory and the Flügge thin shell theory, are introduced. Then, analytical solutions of the combination of a traveling wave form along the circumferential direction and a standing wave form along the axial direction are obtained for both of the thin plate and the OCCS with two opposite edges simply supported. Subsequently, the method of reverberationray matrix (MRRM) is employed to derive the equation of natural frequencies for plate/shell coupled structures and the golden section search algorithm is applied to find the semianalytical natural frequencies of the plate/shell coupled structures. Finally, the calculation results of three typical plate/shell coupled structures are presented and the results are compared with those obtained by the finite element method.
2. Formulation
According to the classical thin plate theory and the Flügge thin shell theory, the force and moment resultants and the governing differential equations for the basic components of the plate/shell coupled structures are presented at the beginning of this section. Analytical solutions of the combination form of a traveling wave along one direction and a standing wave along the other direction are obtained for both of the thin plate and the OCCS with two opposite edges simply supported. After that, the displacements and the force and moment resultants are expressed in matrix form to derive the scattering matrix, the phase matrix and the permutation matrix, which are subsequently used to formulate the reverberationray matrix and to obtain the equation of natural frequencies of the plate/shell coupled structures. Finally, the golden section search algorithm is applied to find the natural frequencies of the plate/shell coupled structures.
2.1. Force and moment resultants in a plate
The force and moment resultants in a thin plate are shown in Fig. 1, in which the positive directions are indicated. Based on the generalized Hooke’s law, the straindisplacement relations and the stressstrain relations of the element of the thin plate, the force and moment resultants in the plate can be expressed in terms of the inplane longitudinal, inplane shear and outplane displacements as follows:
where $u$, $v$ and $w$ denote the inplane longitudinal, inplane shear and outplane displacements along $x$, $y$ and $z$ directions, respectively. $C=Eh/(1{\mu}^{2})$ and $D=E{h}^{3}/12\left(1{\mu}^{2}\right)$ are the membrane stiffness and the bending stiffness of the plate, where $E$ is Young’s modulus, $\mu $ is Poisson’s ratio, and $h$ is the thickness of the plate. ${N}_{x}$ and ${N}_{y}$ denote the inplane normal forces, ${N}_{xy}$ and ${N}_{yx}$, the inplane shear forces, ${M}_{x}$ and ${M}_{y}$, the bending moment, ${M}_{xy}$ and ${M}_{yx}$, the torsional moment, and ${Q}_{xz}$ and ${Q}_{yz}$, the outplane shear forces, ${V}_{xz}$ and ${V}_{yz}$, represent the Kirchhoff effective shear force resultants of the first kind acting on the crosssections perpendicular to the $x$ and $y$ directions, respectively.
Fig. 1Force and moment resultants in a thin plate
2.2. Governing differential equations and the solutions of the plate
According to the dynamic equilibrium of forces in the $x$, $y$ and $z$ directions and the relations of the force and moment resultants with the displacements defined by Eqs. (1)(10), the governing differential equations for the free vibration of a thin plate are obtained as follows:
where $\rho $ is the mass density of the plate.
Taking the Fourier transforms of Eqs. (11)(13), the governing differential equations can be expressed in the frequency domain as:
where a tide over a symbol represents the corresponding physical quantity in the frequency domain, ${k}_{L}=\omega {\left[\left(1{\mu}^{2}\right)\rho /E\right]}^{1/2}$ denotes the inplane longitudinal wave number of the thin plate, and $\omega $ is the circular frequency.
With respect to a thin plate simply supported at $x=$ 0 and $x={L}_{x}$, the inplane longitudinal, inplane shear, and outplane displacements can be expressed as the series summation of the products of the modal waves along $x$ direction and the traveling waves along $y$ direction [51]:
where ${k}_{x}=m\pi /{L}_{x}$ denotes the wave number in the $x$ direction, $m$ is the mode number and ${L}_{x}$ represents the length of the plate. ${k}_{y1}={\left({k}_{x}^{2}{k}_{F}^{2}\right)}^{1/2}$ and ${k}_{y2}={\left({k}_{x}^{2}+{k}_{F}^{2}\right)}^{1/2}$ are respectively the wave numbers along the $y$ direction for the propagating and evanescent waves corresponding to the outplane displacement. ${k}_{y3}={\left({k}_{x}^{2}{k}_{L}^{2}\right)}^{1/2}$ and ${k}_{y4}={\left({k}_{x}^{2}{k}_{S}^{2}\right)}^{1/2}$ are respectively the wave numbers along the $y$ direction for the propagating waves corresponding to the inplane displacements. ${k}_{F}={\left({\omega}^{2}\rho h/D\right)}^{1/4}$ is the in vacuo flexural wave number of the plate. ${k}_{S}=\omega {\left[2\left(1+\mu \right)\rho /E\right]}^{1/2}$ denotes the inplane shear wave number. Wave amplitudes corresponding to the arriving wave and the departing wave are respectively indicated by ${a}_{i}$ and ${d}_{i}$ ($i=$ 14), in which ($i=$ 1, 2) for flexural waves of the outplane displacement and ($i=$ 3, 4) for longitudinal waves and shear waves of the inplane displacements.
The rotation of the normal to the midplane of the plate about the $x$ direction is defined as:
According to Eq. (19), the abovementioned rotation can be expressed in frequency domain as:
Substituting Eqs. (17)19) into the Fourier transforms of Eqs. (2), (3), (5) and (10) yields the frequency domain expressions of the force and moment resultants of the plate:
For an arbitrary axial mode number $m$, Eqs. (17)(19) and (21) can be expressed in matrix form as:
where ${\mathbf{W}}_{d}$ denotes the displacement vector of the plate, ${\mathbf{H}}_{m}\left(x\right)$ indicates the axial mode matrix, and ${\mathbf{W}}_{d}^{*}$ represents the vector of the traveling wave solutions corresponding to the displacement vector. They are presented in detail as follows:
in which ${\mathbf{A}}_{d}$ and ${\mathbf{D}}_{d}$ are coefficient matrices, $\mathbf{a}$ and $\mathbf{d}$ are amplitude vectors corresponding to the arriving wave and the departing wave, respectively. ${\mathbf{P}}_{h}\left(y\right)$ represents the phase matrix. They are presented in detail as follows:
Similarly, for an arbitrary axial mode number $m$, Eqs. (22)(25) can be expressed in matrix form as:
where the physical significance and expression of ${\mathbf{H}}_{m}\left(x\right)$ are the same as those presented in Eq. (28). ${\mathbf{W}}_{f}$ denotes the force vector of the plate, and ${\mathbf{W}}_{f}^{*}$ represents the vector of the wave solutions corresponding to the force vector. They are presented in detail as follows:
in which the physical significances and expressions of ${\mathbf{P}}_{h}\left(y\right)$, $\mathbf{a}$ and $\mathbf{d}$ are the same as those defined in Eq. (29). ${\mathbf{A}}_{f}$ and ${\mathbf{D}}_{f}$ are coefficient matrices corresponding to the arriving wave and the departing wave of the force and moment resultants of the plate. They are presented in detail as follows:
where ${\mu}_{2}$ is a nondimensional parameter presented in Appendix A.1.
2.3. Force and moment resultants in an open circular cylindrical shell
The force and moment resultants in an OCCS are shown in Fig. 2, in which the positive directions are indicated. Based on the Flügge thin shell theory, the force and moment resultants in the OCCS can be expressed in terms of the axial, circumferential and radial displacements as follows:
where $u$, $v$ and $w$ denote the displacement components in the axial ($x$), circumferential ($\theta $), and radial ($z$) directions, respectively. $C=Eh/1{\mu}^{2}$ and $D=E{h}^{3}/12\left(1{\mu}^{2}\right)$ are the membrane stiffness and the bending stiffness of the shell, where $E$ is Young’s modulus, $\mu $ is Poisson’s ratio, $h$ is the thickness and $R$ is the radius of the shell. $\lambda ={h}^{2}/12{R}^{2}$ is a dimensionless parameter, which is related with the ratio of the shell thickness to the shell radius. ${\lambda}_{1}$, ${\lambda}_{2}$, ${\mu}_{1}$, ${\mu}_{2}$ and ${\mu}_{3}$ are dimensionless parameters presented in Appendix A.1. ${N}_{x}$ and ${N}_{\theta}$ denote the inplane normal forces, ${N}_{x\theta}$ and ${N}_{\theta x}$, the inplane shear forces, ${M}_{x}$ and ${M}_{\theta}$, the bending moment, ${M}_{x\theta}$ and ${M}_{\theta x}$, the torsional moment, and ${Q}_{xz}$ and ${Q}_{\theta z}$, the outplane shear forces acting on the crosssections perpendicular to the axial and circumferential directions, respectively. Besides, ${V}_{xz}$ and ${V}_{\theta z}$ indicate the Kirchhoff effective shear force resultants of the first kind, namely, the inplane shear force resultants, and ${F}_{x\theta}$ and ${F}_{\theta x}$, the Kirchhoff effective shear force resultants of the second kind, namely, the outplane shear force resultants acting on the crosssections perpendicular to the axial and circumferential directions, respectively.
Fig. 2Force and moment resultants in an open circular cylindrical shell
2.4. Governing differential equations and solutions of the open circular cylindrical shell
The governing differential equations for the free vibration of an OCCS based on the Flügge thin shell theory can be written as:
$+\mu R\frac{\partial w}{\partial x}{R}^{2}\frac{\left(1{\mu}^{2}\right)\rho}{E}\frac{{\partial}^{2}u}{\partial {t}^{2}}=0,$
$+2\lambda {R}^{2}\frac{{\partial}^{4}w}{\partial {x}^{2}\partial {\theta}^{2}}+\lambda \frac{{\partial}^{4}w}{\partial {\theta}^{4}}+2\lambda \frac{{\partial}^{2}w}{\partial {\theta}^{2}}+{\lambda}_{1}w+{R}^{2}\frac{\left(1{\mu}^{2}\right)\rho}{E}\frac{{\partial}^{2}w}{\partial {t}^{2}}=0.$
Taking the Fourier transforms of Eqs. (53)(55), the governing differential equations can be expressed in the frequency domain as:
$+2\lambda {R}^{2}\frac{{\partial}^{4}\stackrel{~}{w}}{\partial {x}^{2}\partial {\theta}^{2}}+\lambda \frac{{\partial}^{4}\stackrel{~}{w}}{\partial {\theta}^{4}}+2\lambda \frac{{\partial}^{2}\stackrel{~}{w}}{\partial {\theta}^{2}}+{\lambda}_{1}\stackrel{~}{w}{R}^{2}{k}_{L}^{2}\stackrel{~}{w}=0.$
With respect to an OCCS simply supported at $x=$ 0 and $x={L}_{x}$, the axial, circumferential and radial displacements can be expressed as the series summation of the products of the mode waves along the axial direction and the traveling waves along the circumferential direction:
where ${k}_{x}=m\pi /{L}_{x}$ denotes the wave number in the x direction, $m$ is the mode number and ${L}_{x}$ represents the length of the shell. ${a}_{i}$ and ${d}_{i}$ ($i=$ 14) represent wave amplitudes corresponding to the arriving wave and the departing wave, respectively. ${\alpha}_{i}$ and ${\beta}_{i}$($i=$ 14) are amplitude coefficients of the axial and circumferential waves. The expressions of ${\alpha}_{i}$ and ${\beta}_{i}$ are defined as follows:
and ${k}_{\theta i}$ ($i=$ 14) are circumferential wave numbers defined by the following equation:
$+\frac{\left({\xi}_{6}^{2}{k}_{\theta}^{2}+{\xi}_{7}^{2}{\xi}_{8}^{2}+{\xi}_{2}^{2}{\xi}_{9}^{2}\right)\left({\mu}_{1}\lambda {k}_{\theta}^{4}{\xi}_{10}^{2}{k}_{\theta}^{2}+{\xi}_{3}^{2}{\xi}_{8}^{2}\right)}{\lambda}=0,$
where ${\mu}_{i}$ ($i=$ 13), ${\lambda}_{i}$ ($i=$ 12) and ${\xi}_{i}$ ($i=$ 110) are nondimensional parameters presented in detail in Appendix A1 and Appendix A2.
The rotation of the normal to the midsurface of the shell about the axial direction is defined as:
According to Eqs. (60) and (61), the abovementioned rotation can be expressed in frequency domain as:
Meanwhile, substitution of Eqs. (59)(61) into the Fourier transforms of Eqs. (40), (44), (50) and (52) yields the frequency domain expressions of the force and moment resultants in the OCCS:
For an arbitrary axial mode number $m$, Eqs. (59)(61) and (66) can be expressed in matrix form as:
where ${\mathbf{W}}_{d}$ denotes the displacement vector of the OCCS, ${\mathbf{H}}_{m}\left(x\right)$ indicates the axial mode matrix and ${\mathbf{W}}_{d}$^{*} represents the vector of the circumferential traveling wave solutions corresponding to the displacement vector. They are presented in detail as follows:
in which ${\mathbf{P}}_{h}\left(\theta \right)$ denotes the phase matrix. $\mathbf{a}$ and $\mathbf{d}$ are amplitude vectors corresponding to the arriving wave and the departing wave of the displacements of the OCCS, respectively. They are presented in detail as follows:
and ${\mathbf{A}}_{d}$ and ${\mathbf{D}}_{d}$ are coefficient matrices corresponding to the arriving wave and the departing wave of the displacements of the OCCS, respectively. The elements of the coefficient matrices ${\mathbf{A}}_{d}$ and ${\mathbf{D}}_{d}$ are listed as follows:
where $j=$ 1, 2, 3, 4.
Similarly, for an arbitrary axial mode number $m$, Eqs. (67)(70) can be expressed in matrix form as:
where the physical significance and expression of ${\mathbf{H}}_{m}\left(x\right)$ are the same as those presented in Eq. (71). ${\mathbf{W}}_{f}$ denotes the force vector of the shell, and ${\mathbf{W}}_{f}^{*}$ represents the vector of the circumferential wave solutions corresponding to the force vector. They are presented in detail as follows:
in which ${\mathbf{A}}_{f}$ and ${\mathbf{D}}_{f}$ are coefficient matrices corresponding to the arriving wave and the departing wave of the force and moment resultants of the OCCS, respectively. The elements of the coefficient matrices ${\mathbf{A}}_{f}$ and ${\mathbf{D}}_{f}$ are listed as follows:
where $j=$ 1, 2, 3, 4.
2.5. Equation of natural frequencies for the plate/shell coupled structures
Taking advantage of the unidirectional wave form solutions of matrix form for the thin plate obtained in Subsection 2.2 and for the OCCS obtained in Subsection 2.4, the MRRM is introduced to derive the equation of the natural frequencies of the plate/shell coupled structures.
Firstly, the plate/shell coupled structure is discretized into basic components such as flat plates and OCCSs. A dual local coordinate system is established for each of the components at both of the ends, where the cartesian coordinate for a flat plate and the circular cylindrical coordinate for an OCCS. Then, the local scattering matrix is derived according to the continuity conditions of the displacements and the equilibrium conditions of the internal forces and moments at each of the joints of the plate/shell coupled structure. Meanwhile, the local phase matrix is obtained according to the inherent relations of the harmonic waves in the dual local coordinate system. With all of the local scattering matrices and all of the local phase matrices respectively assembled into a global scattering matrix and a global phase matrix, the global scattering equation and the global phase equation are obtained. When keeping the two global amplitude vectors of the arriving wave the same, the two global amplitude vectors of the departing wave contain the same scalar state variables arranged in different sequential orders. Therefore, a permutation equation can be obtained from the relation between the two global amplitude vectors of the departing wave. Subsequently, the reverberationray matrix can be obtained from the simultaneous equations of the global scattering equation, the global phase equation and the permutation equation. Finally, the equation of the natural frequencies can be derived by equating the determinant of the coefficient matrix of the global amplitude vector of the departing wave to zero.
With the derivation procedure mentioned above, the equations of the natural frequencies of the three plate/shell coupled structures including a boxtype structure, a racetrack cylindrical shell and a ship hull structure will be obtained in the following discussions in this subsection. However, since the three independent derivation procedures are quite similar to each other, for simplicity, one of them will be taken as an example and the other two will be omitted. Note that the solutions obtained in Subsection 2.2 are applied for flat plate components while the solutions obtained in Subsection 2.4 are applied for OCCS components.
Next, the derivation for equation of the natural frequencies of the racetrack cylindrical shell shown in Fig. 3, is presented as follows.
Fig. 3Dual local coordinate systems for the racetrack cylindrical shell
2.5.1. Scattering matrix
The continuity conditions for the plate and the OCCS at Node Line 1 are presented as follows:
which can be rewritten in matrix form as:
where ${\mathbf{T}}_{d}^{1}=\mathrm{d}\mathrm{i}\mathrm{a}\mathrm{g}\left\{1111\right\}$ denotes the transfer matrix corresponding to the displacement vector at Node Line 1. The variables with superscript $IJ$ ($I$, $J=$ 14) indicate the physical quantities of the substructure bounded by Node Line $I$ and Node Line $J$, this will not be mentioned repeatedly in the following discussions.
Since the local coordinates are established at the node lines, the phase matrix turns to be a unit matrix at the node lines. Taking into account of this condition and substituting Eqs. (26) and (71) into Eq. (90) yields:
Meanwhile, the equilibrium conditions for the plate and the OCCS at Node Line 1 are presented as:
which can be rewritten in matrix form as:
where ${\mathbf{T}}_{f}^{1}={\mathbf{T}}_{d}^{1}$ denotes the transfer matrix corresponding to the force vector at Node Line 1.
Substitution of Eqs. (34) and (82) into Eq. (93) results in:
Eqs. (91) and (94) can be combined and expressed in a single matrix form as:
where ${\mathbf{d}}^{1}={\left\{{\left({\mathbf{d}}^{14}\right)}^{T}{\left({\mathbf{d}}^{12}\right)}^{T}\right\}}^{T}$ and ${\mathbf{a}}^{1}={\left\{{\left({\mathbf{a}}^{14}\right)}^{T}{\left({\mathbf{a}}^{12}\right)}^{T}\right\}}^{T}$ are amplitude coefficients of the departing wave and arriving wave at Node Line 1, respectively. ${\mathbf{S}}^{1}$, the scattering matrix at Node Line 1, is defined as:
In the same manner, the scattering relations and scattering matrices for the rest node lines can be obtained. For simplicity, the derivation procedures are omitted and the results are given straightforwardly as follows.
The scattering relations at an arbitrary node line $J$ can be presented as:
where ${\mathbf{d}}^{J}={\left\{{\left({\mathbf{d}}^{JI}\right)}^{T}{\left({\mathbf{d}}^{JK}\right)}^{T}\right\}}^{T}$ and ${\mathbf{a}}^{J}={\left\{{\left({\mathbf{a}}^{JI}\right)}^{T}{\left({\mathbf{a}}^{JK}\right)}^{T}\right\}}^{T}$ are amplitude coefficients of the departing wave and arriving wave at Node Line $J$, respectively. ${\mathbf{S}}^{J}$, the scattering matrix at Node Line $J$, is defined as:
in which $J=$ 2, 3, 4. As $J=$ 2 and 3, $I=J1$ and $K=J+1$. However, as $J=$ 4, $I=$ 3 and $K=$ 1.
Assembling all of the local scattering equations for Node Line 14 by stacking ${\mathbf{d}}^{1}$${\mathbf{d}}^{4}$ and ${\mathbf{a}}^{1}$${\mathbf{a}}^{4}$ into two column vectors $\mathbf{d}$ and $\mathbf{a}$, the global scattering equation can be obtained as follows:
where $\mathbf{d}$ and $\mathbf{a}$ are global amplitude vectors of the departing wave and the arriving wave, and $\mathbf{S}$ is the global scattering matrix. They are presented in detail as follows:
2.5.2. Phase matrix
The phase relations of harmonic waves in the dual coordinate system provide additional equations for solving the unknown amplitude vectors. Note that the departing wave from Node Line 1 (3) is exactly the arriving wave to Node Line 2 (4), and vice versa. Therefore, the amplitudes of the departing wave and the arriving wave differ with each other by a phase factor.
With the employment of the solutions for the plate obtained in Subsection 2.2, the relations between the amplitudes of the departing wave and the arriving wave between Node Line 1 (3) and Node Line 2 (4) are presented as:
where $IJ=$ 12 or 34 and $JI=$ 21 or 43.
Similarly, with the employment of the solutions for the OCCS obtained in Subsection 2.4, the relations between the amplitudes of the departing wave and the arriving wave between Node Line 2 (4) and Node Line 3 (1) are presented as:
where $IJ=$ 23 or 41 and $JI=$ 32 or 14.
Assembling all of the local phase equations defined by Eqs. (103)(106) results in the global phase equation:
where the physical significance and expression of $\mathbf{a}$ are the same as the one presented in Eq. (99). ${\mathbf{d}}^{*}$ is a rearranged global amplitude vector of the departing wave, and $\mathbf{P}$ is the global phase matrix. They are presented in detail as follows:
where ${\theta}_{0}$ denotes the included angle of the OCCS, and ${L}_{y}$ represents the length of the plate in y direction.
2.5.3. Permutation matrix
A comparison of the global amplitude vectors of the departing wave $\mathbf{d}$ and ${\mathbf{d}}^{*}$ indicates that the two amplitude vectors contain the same scalar state variables arranged in different sequential orders. The relation between $\mathbf{d}$ and ${\mathbf{d}}^{*}$ is:
where $\mathbf{U}$ is the permutation matrix, which is presented in detail as follows:
in which ${0}_{4}$ and ${\mathbf{I}}_{4}$ are respectively zero matrix and unit matrix of fourth order.
2.5.4. Reverberationray matrix and the equation of natural frequencies
Substitution of Eqs. (107) and (110) into Eq. (99) yields:
where $\mathbf{I}$ is a unit matrix of 32nd order, and $\mathbf{R}=\mathbf{S}\mathbf{P}\mathbf{U}$ is defined as the reverberationray matrix of the racetrack cylindrical shell.
To obtain a nontrivial solution of the global amplitude vector of the departing wave, the determinant of $\left(\mathbf{I}\mathbf{R}\right)$ must be zero, namely:
which is the equation of natural frequencies of the racetrack cylindrical shell.
2.6. Searching algorithm for natural frequencies
As the equation of natural frequencies of the plate/shell coupled structure is obtained, the problem at hand is to solve the equation for the natural frequencies. It is obvious that the left hand side of Eq. (113) is a function of frequency, and the natural frequencies are zeros of the function. Unfortunately, for most of the frequencies, the function values are complex numbers. Therefore, finding the zeros of the function needs to search the common zeros of the real part and the imaginary part of the function. A good idea is to search the zeros or the minimal values of the absolute value of the function. This simple approach is adopted in this paper and the golden section search algorithm is introduced to determine the natural frequencies of the plate/shell coupled structure. The procedure for determining the natural frequencies of the plate/shell coupled structure is same to the one for free vibration analysis of open and closed circular cylindrical shell by MRRM presented in [5254].
3. Results and discussions
In this section, free vibration analysis of plate/shell coupled structures with two opposite edges simply supported is presented to verify the validity and accuracy of the present method. The natural frequencies of the three plate/shell coupled structures are calculated by MRRM and by FEM, and the comparison results are presented in tabular form. It is particularly pointed out that all the results obtained by FEM in the following discussions are calculated with the commercial software ANSYS.
3.1. The boxtype structure
Studies on the free vibration of the boxtype structure, as shown in Fig. 4, with two opposite edges simply supported are conducted in this subsection. The material properties of the boxtype structure are: Young’s modulus $E=$ 2.1×10^{11} Pa, Poisson’s ratio $\mu =$ 0.3, mass density $\rho =$ 7800 kg m^{−3}, and the geometrical parameters of the boxtype structure are: ${L}_{x}=$ 10 m, ${L}_{y1}={L}_{y2}=$ 4 m, $h=$ 0.01 m. The results for natural frequencies of the boxtype structure obtained by MRRM are compared with those obtained by FEM in Table 1, where $m$ denotes the axial mode number, $n$ represents the mode number in the circumferential direction, and PError indicates the percentage error between the results obtained by MRRM and FEM.
Fig. 4Geometry and notations of a boxtype structure
Table 1Comparison of natural frequencies obtained by MRRM and FEM for the boxtype structure
Mode number  Natural frequencies and percentage errors  
$n$  $m$  1  2  3  4  5  6  7  8 
1  FEM  1.787  2.526  3.758  5.484  7.703  10.416  13.621  17.320 
MRRM  1.788087  2.528058  3.761279  5.487782  7.707571  10.420647  13.627011  17.326663  
PError  0.06 %  0.08 %  0.09 %  0.07 %  0.06 %  0.04 %  0.04 %  0.04 %  
2  FEM  2.596  3.206  4.305  5.926  8.067  10.723  13.886  17.552 
MRRM  2.597011  3.207152  4.307759  5.928857  8.070860  10.726757  13.890259  17.556971  
PError  0.04 %  0.04 %  0.06 %  0.05 %  0.05 %  0.04 %  0.03 %  0.03 %  
3  FEM  3.635  4.116  5.053  6.523  8.548  11.117  14.217  17.835 
MRRM  3.633189  4.116464  5.053479  6.524894  8.550082  11.119675  14.219630  17.838747  
PError  0.05 %  0.01 %  0.01 %  0.03 %  0.02 %  0.02 %  0.02 %  0.02 %  
4  FEM  6.411  7.149  8.380  10.104  12.321  15.032  18.236  21.934 
MRRM  6.412601  7.152497  8.385686  10.112165  12.331930  15.044982  18.251319  21.950942  
PError  0.02 %  0.05 %  0.07 %  0.08 %  0.09 %  0.09 %  0.08 %  0.08 %  
5  FEM  8.015  8.662  9.768  11.357  13.445  16.039  19.142  22.752 
MRRM  8.012426  8.664084  9.772218  11.363087  13.452738  16.048779  19.153304  22.765631  
PError  0.03 %  0.02 %  0.04 %  0.05 %  0.06 %  0.06 %  0.06 %  0.06 %  
6  FEM  9.798  10.384  11.374  12.824  14.768  17.224  20.203  23.706 
MRRM  9.817337  10.387455  11.377389  12.827979  14.772323  17.230424  20.210830  23.714801  
PError  0.20 %  0.03 %  0.03 %  0.03 %  0.03 %  0.04 %  0.04 %  0.04 %  
7  FEM  14.118  14.855  16.085  17.807  20.023  22.732  25.934  29.629 
MRRM  14.115270  14.858683  16.092494  17.819161  20.038987  22.752050  25.958374  29.657972  
PError  0.02 %  0.02 %  0.05 %  0.07 %  0.08 %  0.09 %  0.09 %  0.10 %  
8  FEM  16.484  17.169  18.300  19.901  21.984  24.560  27.633  31.209 
MRRM  16.503369  17.175778  18.307301  19.910498  21.997010  24.575867  27.653055  31.231947  
PError  0.12 %  0.04 %  0.04 %  0.05 %  0.06 %  0.06 %  0.07 %  0.07 % 
It can be found from Table 1 that, the natural frequencies obtained by MRRM and FEM agree well with each other. The maximum of the percentage errors is 0.20 %, and most of the percentage errors are no larger than 0.10 %. The small discrepancies in the results should be attributed to the approximation of FEM. Therefore, it indicates that MRRM is validate and of high precision for free vibration analysis of plate coupled structures such as the boxtype structure.
3.2. The racetrack cylindrical shell
Consider an isotropic, racetrack cylindrical shell composed of thin plates and OCCSs with axial length ${L}_{x}=$ 10 m, uniform thickness $h=$ 0.01 m, circumferential length of the plate ${L}_{y}=$ 4 m, middle surface radius of the OCCS $R=$ 2 m, as shown in Fig. 5. The material properties of the racetrack cylindrical shell are the same as those defined for the boxtype structure. The results for the natural frequencies of the racetrack cylindrical shell obtained by MRRM are compared with those obtained by FEM in Table 2, in which the notations are of the same meaning as those in Table 1.
Table 2 shows that the natural frequencies obtained by MRRM and FEM agree well with each other. The maximum of the percentage errors is 1.34%, and most of the percentage errors are no larger than 1.00 %. The difference between the results obtained by MRRM and FEM may be caused by the approximation of FEM and the different shell theories adopted by FEM and this paper. Therefore, it indicates that MRRM is validate for free vibration analysis of plateshell coupled structures such as the racetrack cylindrical shell, and the results are of high precision.
Table 2Comparison of natural frequencies obtained by MRRM and FEM for the racetrack cylindrical shell
Mode number  Natural frequencies and percentage errors  
$n$  $m$  1  2  3  4  5  6  7  8 
1  FEM  2.151  2.953  4.090  5.715  7.854  10.508  13.670  17.336 
MRRM  2.151035  2.957795  4.099758  5.727604  7.869848  10.525766  13.689947  17.357884  
PError  0.00 %  0.16 %  0.24 %  0.22 %  0.20 %  0.17 %  0.15 %  0.13 %  
2  FEM  5.383  6.984  8.461  10.232  12.415  15.064  18.200  21.832 
MRRM  5.456198  7.031901  8.506268  10.281368  12.471784  15.128039  18.271764  21.911835  
PError  1.34 %  0.68 %  0.53 %  0.48 %  0.46 %  0.42 %  0.39 %  0.36 %  
3  FEM  5.444  13.162  15.027  17.038  19.382  22.133  25.328  28.989 
MRRM  5.514263  13.195591  15.093706  17.133938  19.500229  22.270141  25.483314  29.162474  
PError  1.27 %  0.25 %  0.44 %  0.56 %  0.61 %  0.62 %  0.61 %  0.59 %  
4  FEM  10.458  20.291  23.281  25.846  28.524  31.492  34.835  38.598 
MRRM  10.460565  20.468996  23.452468  26.038717  28.741440  31.738403  35.110916  38.903664  
PError  0.02 %  0.87 %  0.73 %  0.74 %  0.76 %  0.78 %  0.79 %  0.79 %  
5  FEM  10.684  20.334  33.382  36.659  39.751  43.017  46.583  50.513 
MRRM  10.687016  20.525417  33.560942  36.926488  40.084378  43.403142  47.016813  50.994129  
PError  0.03 %  0.93 %  0.53 %  0.72 %  0.83 %  0.89 %  0.92 %  0.94 %  
6  FEM  13.803  27.932  33.447  48.481  52.618  56.447  60.381  64.577 
MRRM  13.927559  27.962842  33.633721  48.880722  53.107713  57.012575  61.015628  65.278488  
PError  0.89 %  0.11 %  0.56 %  0.82 %  0.92 %  0.99 %  1.04 %  1.07 %  
7  FEM  15.378  28.755  42.048  48.529  66.309  71.352  75.927  80.554 
MRRM  15.492372  28.837613  42.200925  48.948598  66.815182  72.081826  76.777865  81.507231  
PError  0.74 %  0.29 %  0.36 %  0.86 %  0.76 %  1.01 %  1.11 %  1.17 %  
8  FEM  15.955  30.057  42.961  57.665  72.451  84.556  92.219  97.881 
MRRM  15.956436  30.175983  43.324195  57.735824  72.588720  84.589523  93.167426  99.063747  
PError  0.01 %  0.39 %  0.84 %  0.12 %  0.19 %  0.04 %  1.02 %  1.19 % 
Fig. 5Geometry and notations of a racetrack cylindrical shell
Fig. 6Geometry and notations of a ship hull structure
3.3. The ship hull structure
In this subsection, the natural frequencies of the ship hull structure, as shown in Fig. 6, with two opposite edges simply supported are calculated. The material properties of the ship hull structure are the same as those defined for the boxtype structure. The geometrical parameters are: ${L}_{x}=$ 10 m, ${L}_{y1}=$ 2 m, ${L}_{y2}=$ 4m, $R=$ 2 m and $h=$ 0.01 m. The results for the natural frequencies of the boxtype structure obtained by MRRM are compared with those obtained by FEM in Table 3, where the notations are of the same meaning as those in Tables 1 and 2.
Table 3Comparison of natural frequencies obtained by MRRM and FEM for the ship hull structure
Mode number  Natural frequencies and percentage errors  
$n$  $m$  1  2  3  4  5  6  7  8 
1  FEM  2.562  5.713  6.575  8.847  13.875  16.349  16.377  20.440 
MRRM  2.574674  5.803015  6.651149  8.818441  13.925635  16.342406  17.035003  20.525130  
PError  0.49 %  1.55 %  1.14 %  0.32 %  0.36 %  0.04 %  3.86 %  0.41 %  
2  FEM  3.217  7.027  8.241  9.973  16.702  21.496  22.775  28.663 
MRRM  3.222129  7.072771  8.300801  9.991594  16.703286  21.734920  23.026966  27.617563  
PError  0.16 %  0.65 %  0.72 %  0.19 %  0.01 %  1.10 %  1.09 %  3.79 %  
3  FEM  4.318  8.407  9.681  11.209  18.061  24.346  25.892  30.900 
MRRM  4.322056  8.441562  9.737088  11.237360  18.073568  24.512647  26.124280  31.037877  
PError  0.09 %  0.41 %  0.58 %  0.25 %  0.07 %  0.68 %  0.89 %  0.44 %  
4  FEM  5.934  10.159  11.372  12.757  19.754  26.778  28.435  32.886 
MRRM  5.937682  10.193966  11.432079  12.791735  19.773413  26.937417  28.661965  33.011183  
PError  0.06 %  0.34 %  0.53 %  0.27 %  0.10 %  0.59 %  0.79 %  0.38 %  
5  FEM  8.071  12.362  13.460  14.723  21.882  29.349  31.044  35.145 
MRRM  8.075624  12.400127  13.527708  14.764371  21.906555  29.520271  31.286598  35.270387  
PError  0.06 %  0.31 %  0.50 %  0.28 %  0.11 %  0.58 %  0.78 %  0.36 %  
6  FEM  10.725  15.043  16.010  17.168  24.480  32.245  33.918  37.775 
MRRM  10.729149  15.085841  16.086390  17.215377  24.509296  32.435677  34.185502  37.908113  
PError  0.04 %  0.28 %  0.47 %  0.28 %  0.12 %  0.59 %  0.78 %  0.35 %  
7  FEM  13.886  18.212  19.052  20.122  27.566  35.541  37.152  40.826 
MRRM  13.891362  18.261305  19.137747  20.174461  27.599471  35.755749  37.449705  40.969879  
PError  0.04 %  0.27 %  0.45 %  0.26 %  0.12 %  0.60 %  0.79 %  0.35 %  
8  FEM  17.552  21.874  22.599  23.598  31.149  39.276  40.800  44.330 
MRRM  17.557387  21.930540  22.694437  23.653557  31.186065  39.516188  41.130030  44.484581  
PError  0.03 %  0.26 %  0.42 %  0.23 %  0.12 %  0.61 %  0.80 %  0.35 % 
It can be observed from Table 3 that, the natural frequencies obtained by MRRM and FEM agree well with each other. Except for certain mode numbers, the maximum of the percentage errors is 1.55 %, and most of the percentage errors are no larger than 1.00 %. The difference between the results obtained by MRRM and FEM may be caused by the approximation of FEM and the different shell theories adopted by FEM and this paper. Therefore, it indicates that MRRM is applicable for free vibration analysis of plateshell coupled structures such as the ship hull structure, and the results are of high precision.
4. Conclusions
This paper presents a semianalytical solution procedure and accurate calculation results for plate/shell coupled structures with two opposite edges simply supported. The validity and applicability of the MRRM for free vibration analysis of plate/shell coupled structures are verified. It has been proved that the results obtained by MRRM are in excellent agreement with those obtained by FEM and that high precision of MRRM has been shown. MRRM is advantageous in its simple and uniform formulation as well as accurate results for dynamic response analysis of coupled structures.
It should be pointed out that the method of reverberationray matrix only applies to plate/shell coupled structures with two opposite edges simplysupported at present. This limitation may be breached by replacing the standing wave form solution with the combination of the trigonometric function and the tangent or cotangent functions. The effects of the structural parameters, boundary conditions and connection forms on the vibration characteristics of the plate/shell coupled structures will be discussed in the subsequent researches.
References

Irie T., Yamada G., Kobayash Y. Free vibration of a cantilever folded plate. The Journal of the Acoustical Society of America, Vol. 76, Issue 6, 1984, p. 17431748.

Eterovic A. L., Godoy L. A. An exact strip method for folded plate structures. Computers and Structures, Vol. 32, Issue 2, 1989, p. 263276.

Maleki S. Compound strip method for box girders and folded plates. Computers and Structures, Vol. 40, Issue 3, 1991, p. 527538.

Liu W. H., Huang C. C. Vibration analysis of folded plates. Journal of Sound and Vibration, Vol. 157, Issue 1, 1992, p. 123137.

Hinton E., Özakça M., Rao N. V. R. Free vibration analysis and shape optimization of variable thickness plates, prismatic folded plates and curved shells. Part 1: finite strip formulation. Journal of Sound and Vibration, Vol. 181, Issue 4, 1995, p. 553566.

Niyogi A. G., Laha M. K., Sinha P. K. Finite element vibration analysis of laminated composite folded plate structures. Shock and Vibration, Vol. 6, Issues 56, 1999, p. 273283.

Samanta A., Mukhopadhyay M. Finite element static and dynamic analyses of folded plates. Engineering Structures, Vol. 21, Issue 3, 1999, p. 277287.

Lai Y., Yu W. Analytical solution for forced vibration of a simplysupported Vshaped folded plate roof. ThinWalled Structures, Vol. 40, Issue 3, 2002, p. 215223.

Peng L. X., Kitipornchai S., Liew K. M. Free Vibration Analysis of Folded Plate Structures by the FSDT Meshfree Method. Computational Mechanics, Vol. 39, Issue 6, 2007, p. 799814.

Cuschieri J. M. Structural powerflow analysis using a mobility approach of an Lshaped plate. The Journal of the Acoustical Society of America, Vol. 87, Issue 3, 1990, p. 11591165.

Cuschieri J. M. Parametric analysis of the power flow on an Lshaped plate using a mobility power flow approach. The Journal of the Acoustical Society of America, Vol. 91, Issue 5, 1992, p. 26862695.

Wang Z. H., Xing J. T., Price W. G. An investigation of power flow characteristics of Lshaped plates adopting a substructure approach. Journal of Sound and Vibration, Vol. 250, Issue 4, 2002, p. 627648.

Kessissoglou N. J. Power transmission in Lshaped plates including flexural and inplane vibration. The Journal of the Acoustical Society of America, Vol. 115, Issue 3, 2004, p. 11571169.

Chen X., Sheng M. Research on vibration characteristics of Lshaped plate using a mobility power flow approach. Journal of Marine Science and Application, Vol. 6, Issue 3, 2007, p. 1216.

Liu C., Li F., Huang W. Active vibration control of finite Lshaped beam with travelling wave approach. Acta Mechanica Solida Sinica, Vol. 23, Issue 5, 2010, p. 377385.

Lin T. R., Tan C. C., Yan C., Hargreaves D. Vibration of Lshaped plates under a deterministic force or moment excitation: a case of statistical energy analysis application. Journal of Sound and Vibration, Vol. 330, Issue 20, 2011, p. 47804797.

Liu C., Li F., Tang L., Huang W. Vibration control of the finite Lshaped beam structures based on the active and reactive power flow. Science China Physics, Mechanics and Astronomy, Vol. 54, Issue 2, 2011, p. 310319.

Liu C., Li F., Liang T., Huang W. Early short time transient response of finite Lshaped Mindlin plate. Wave Motion, Vol. 48, Issue 5, 2011, p. 371391.

Liu C., Li F., Liang T., Huang W. The wave and vibratory power transmission in a finite Lshaped Mindlin plate with two simply supported opposite edges. Acta Mechanica Sinica, Vol. 27, Issue 5, 2011, p. 785795.

Lin Y., Wu W., Li X., Gan J. Study on transmission law of flexural waves through Lshaped steelaluminum joints. The Twentysecond International Offshore and Polar Engineering Conference, International Society of Offshore and Polar Engineers, 2012.

Dickinson S., Warburton G. Vibration of boxtype structures. Journal of Mechanical Engineering Science, Vol. 9, Issue 4, 1967, p. 325338.

Popplewell N. The vibration of a boxtype structure I. Natural frequencies and normal modes. Journal of Sound and Vibration, Vol. 14, Issue 3, 1971, p. 357365.

Handa K. Analysis of inplane vibration of boxtype structures by a finite element method. Journal of Sound and Vibration, Vol. 21, Issue 1, 1972, p. 107114.

Fulford R. A., Petersson B. A. T. Estimation of vibrational power in builtup systems involving boxlike structures. Part 1: uniform force distribution. Journal of Sound and Vibration, Vol. 232, Issue 5, 2000, p. 877895.

Fulford R. A., Petersson B. A. T. Estimation of vibrational power in builtup systems involving boxlike structures. Part 2: infinite topplate and circular geometry. Journal of Sound and Vibration, Vol. 232, Issue 5, 2000, p. 897915.

Lee S. Y., Wooh S. C. Finite element vibration analysis of composite box structures using the high order plate theory. Journal of Sound and Vibration, Vol. 277, Issues 45, 2004, p. 801814.

Lin T. R., Pan J. Vibration characteristics of a boxtype structure. Journal of Vibration and Acoustics, Vol. 131, Issue 3, 2009, p. 19.

Lin T. R., Pan J. Sound radiation characteristics of a boxtype structure. Journal of Sound and Vibration, Vol. 325, Issues 45, 2009, p. 835851.

Chen Y., Jin G., Zhu M., Liu Z., Du J., Li W. Vibration behaviors of a boxtype structure built up by plates and energy transmission through the structure. Journal of Sound and Vibration, Vol. 331, Issue 4, 2012, p. 849867.

Yamada G., Irie T., Tamiya T. Free vibration of a circular cylindrical doubleshell system closed by end plates. Journal of Sound and Vibration, Vol. 108, Issue 2, 1986, p. 297304.

Schlesinger A. Transmission of elastic waves from a cylinder to an attached flat plate. Journal of Sound and Vibration, Vol. 186, Issue 5, 1995, p. 761780.

Tso Y. K., Hansen C. H. Wave propagation through cylinder/plate junctions. Journal of Sound and Vibration, Vol. 186, Issue 3, 1995, p. 447461.

Stanley A. J., Ganesan N. Frequency response of shellplate combinations. Computers and Structures, Vol. 59, Issue 6, 1996, p. 10831094.

Tso Y. K., Hansen C. H. An investigation of the coupling loss factor for a cylinder/plate structure. Journal of Sound and Vibration, Vol. 199, Issue 4, 1997, p. 629643.

Wu J. H., Chen H. L., An W. B. A method to predict sound radiation from a plateended cylindrical shell excited by an external force. Journal of Sound and Vibration, Vol. 237, Issue 5, 2000, p. 739803.

Wang Z. H., Xing J. T., Price W. G. A study of power flow in a coupled platecylindrical shell system. Journal of Sound and Vibration, Vol. 271, Issues 35, 2004, p. 863882.

Liang S., Chen H. L. The natural vibration of a conical shell with an annular end plate. Journal of Sound and Vibration, Vol. 294, Issue 45, 2006, p.927943.

Liang S., Chen H. L., Chen T., Wang M. Y. The natural vibration of a symmetric crossply laminated composite conicalplate shell. Composite Structures, Vol. 80, Issue 2, 2007, p. 265278.

Efraim E., Eisenberger M. Exact vibration frequencies of segmented axisymmetric shells. ThinWalled Structures, Vol. 44, Issue 3, 2006, p. 281289.

Caresta M., Kessissoglou N. J. Free vibrational characteristics of isotropic coupled cylindricalconical shells. Journal of Sound and Vibration, Vol. 329, Issue 6, 2010, p. 733751.

Kang J. H. Threedimensional vibration analysis of joined thick conicalCylindrical shells of revolution with variable thickness. Journal of Sound and Vibration, Vol. 331, Issue 18, 2012, p. 41874198.

Qu Y., Chen Y., Long X., Hua H., Meng G. A modified variational approach for vibration analysis of ringstiffened conicalcylindrical shell combinations. European Journal of Mechanics – A/Solids, Vol. 37, 2013, p. 200215.

Qu Y., Wu S., Chen Y., Hua H. Vibration analysis of ringstiffened conicalcylindricalspherical shells based on a modified variational approach. International Journal of Mechanical Sciences, Vol. 69, 2013, p. 7284.

Shakouri M., Kouchakzadeh M. A. Stability analysis of joined isotropic conical shells under axial compression. ThinWalled Structures, Vol. 72, 2013, p. 2027.

Wu S., Qu Y., Hua H. Vibration characteristics of a sphericalcylindricalspherical shell by a domain decomposition method. Mechanics Research Communications, Vol. 49, 2013, p. 1726.

Kouchakzadeh M. A., Shakouri M. Free vibration analysis of joined crossply laminated conical shells. International Journal of Mechanical Sciences, Vol. 78, 2014, p. 118125.

Ma X., Jin G., Xiong Y., Liu Z. Free and forced vibration analysis of coupled conicalcylindrical shells with arbitrary boundary conditions. International Journal of Mechanical Sciences, Vol. 88, 2014, p. 122137.

Shakouri M., Kouchakzadeh M. A. Free vibration analysis of joined conical shells: analytical and experimental study. ThinWalled Structures, Vol. 85, 2014, p. 350358.

Su Z., Jin G., Ye T. Threedimensional vibration analysis of thick functionally graded conical, cylindrical shell and annular plate structures with arbitrary elastic restraints. Composite Structures, Vol. 118, 2014, p. 432447.

Xie X., Zheng H., Jin G. Integrated orthogonal polynomials based spectral collocation method for vibration analysis of coupled laminated shell structures. International Journal of Mechanical Sciences, Vol. 98, 2015, p. 132143.

Tang D., Pang F., Wang Q., Yao X. Research on transmission loss of power flow through a finite Vshaped plate. Journal of Vibration Engineering, Vol. 29, Issue 1, 2016, p. 112122.

Tang D., Wu G., Yao X., Wang C. Free vibration analysis of circular cylindrical shells with arbitrary boundary conditions by the method of reverberationray matrix. Shock and Vibration, 2016.

Yao X., Tang D., Pang F., Li S. Exact free vibration analysis of open circular cylindrical shells by the method of reverberation ray matrix. Journal of Zhejiang University – Science A (Applied Physics and Engineering), Vol. 17, Issue 4, 2016, p. 295316.

Tang D., Sun L., Yao X., Yue X. Free vibration analysis of open circular cylindrical shells by the method of reverberationray matrix. Advances in Mechanical Engineering, Vol. 8, Issue 3, 2016, p. 121.
About this article
This work is financially supported by National Natural Science Foundation of China (Grant Nos. 51479041, 51279038). The authors would like to express their profound thanks for the financial support and sincerely thank Miss Jingjing Yu for the scientific discussions and suggestions and the anonymous reviewers for the critical and constructive comments on this paper.