AAVAdvances in Acoustics and Vibration1687-627X1687-6261Hindawi Publishing Corporation98436110.1155/2010/984361984361Research ArticleFree and Forced Vibrations of Elastically Connected StructuresKellyS. GrahamLiewK. M.Department of Mechanical Engineering, The University of AkronAkron, Oh 44235USAuakron.edu201028112010201021102010111120102010Copyright © 2010 S. Graham Kelly.This is an open access article distributed under the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

A general theory for the free and forced responses of n elastically connected parallel structures is developed. It is shown that if the stiffness operator for an individual structure is self-adjoint with respect to an inner product defined for Ck[0,1], then the stiffness operator for the set of elastically connected structures is self-adjoint with respect to an inner product defined on U=Rn×Ck[0,1]. This leads to the definition of energy inner products defined on U. When a normal mode solution is used to develop the free response, it is shown that the natural frequencies are the square roots of the eigenvalues of an operator that is self-adjoint with respect to the energy inner product. The completeness of the eigenvectors in W is used to develop a forced response. Special cases are considered. When the individual stiffness operators are proportional, the problem for the natural frequencies and mode shapes reduces to a matrix eigenvalue problem, and it is shown that for each spatial mode there is a set of n intramodal mode shapes. When the structures are identical, uniform, or nonuniform, the differential equations are uncoupled through diagonalization of a coupling stiffness matrix. The most general case requires an iterative solution.

1. Introduction

The general theory for the free and forced response of strings, shafts, beams, and axially loaded beams is well documented . Investigators have examined the free and forced response of elastically connected strings [9, 10], Euler-Bernoulli beams , and Timoshenko beams . These analyses focused on a pair of elastically connected structures using a normal-mode solution for the free response and a modal analysis for the forced response. Each of these papers uses a normal-mode solution or a modal analysis specific to the problem to obtain a solution.

Ru  proposed that model for multiwalled carbon nanotubes to be modeled by elastically connected structures with the elastic layers representing interatomic vanDer Waals forces. Ru  proposed a model of concentric beams connected by elastic layers to model buckling of carbon nanotubes and elastic shell models [17, 18]. Yoon et al.  and Li and Chou  modeled free vibrations of multiwalled nanotubes by a series of concentric elastically connected Euler-Bernoulli beams, while Yoon et al. [21, 22] modeled nanotubes as concentric Timoshenko beams connected by an elastic layer. Xu et al.  modeled the nonliearity of the vanDer Waals forces. Elishakoff and Pentaras  gave approximate formulas for the natural frequencies of double-walled nanotubes noting that if developed from the eigenvalue relation, the computations can be compuitationally intensive and difficult.

Kelly and Srinivas  developed a Rayleigh-Ritz method for elastically connected stretched structures.

This paper develops a general theory within which a finite set of parallel structures connected by elastic layers of a Winkler type can be analyzed. The theory shows that the determination of the natural frequencies for uniform parallel structures such as shafts and Euler-Bernoulli beams can be reduced to matrix eigenvalue problems. The general theory is also used to develop a modal analysis for forced response of a set of parallel structures.

2. Problem Formulation

The problem considered is that of n structural elements in parallel but connected by elastic layers. Each elastic layer is modeled by a Winkler foundation, a layer of distributed stiffness across the span of the element. For generality, it is assumed that the outermost structures are connected to fixed foundations through elastic layers, as illustrated in Figure 1. Each layer has a uniform stiffness per unit length kii=0,1,2,,n.

Set of n elastically connected parallel structures.

Let wi(x) represent the displacement of the ith structure. If isolated from the system, the nondimensional differential equation governing the time-dependent motion of this structure is written as Liwi(x)+Miẅi=Gi(x,t), where Li is the stiffness operator for the element, Mi is an inertia operator for the element, and Gi(x,t) is the force per unit length acting on the structure which includes the forces from the elastic layer as well as any externally applied forces. The stiffness operator is a differential operator of order k  (k=2 for strings and shafts, k=4 for Euler-Bernoulli beams), where the inertia operator is a function of the independent variable x.

Each structure has the same end supports, and therefore their differential equations are subject to the same boundary conditions. Let S be the subspace of Ck[0,1] defined by the boundary conditions; all elements in S satisfy all boundary conditions.

The external forces acing on the ith structure areGi(x,t)={-λ0w1-λ1(w1-w2)+F1(x,t)i=1,-λi-1(wi-wi-1)-λi(wi-wi+1)+Fi(x,t)i=2,3,,n-1,-λn-1(wn-wn-1)-λnwn+Fn(x,t)i=n, where λi,i=0,1,,n are the nondimensional stiffness coefficients connecting the ith and i plus first structures. Substitution of (2) into (1) leads to a coupled set of differential equations which are written in a matrix form as (K+Kc)W+MẄ=F, where W=[w1(x,t)w2(x,t)wn(x,t)]T, F=[F1(x,t)F2(x,t)Fn(x,t)]T, K is an n×n diagonal operator matrix with k  i,i  =Li, M is an n×n diagonal mass matrix with mi,i=Mi, and Kc is a tridiagonal n×n stiffness coupling matrix with (kc)i,i-1=-λi-1i=2,3,,(kc)i,i=λi-1+λii=1,2,,n,(kc)i,i+1=-λii=1,2,,n-1. The vector W is an element of the vector space U=S×Rn; an element of U is an n-dimensional vector, whose elements all belong to S.

3. General Theory

Let f(x) and g(x) be arbitrary elements of S. A standard inner product on S is defined as (f,g)S=01f(x)g(x)dx.

If the stiffness operator is self-adjoint [(Lif,g)S=(f,Lig)S] and positive definite [(Lif,f)s0 and (Lif,f)s=0 if and only if f=0] with respect to the standard inner product, then a potential energy inner product is defined as(f,g)Li=(Lif,g)S.

Clearly each Mi is positive definite and self-adjoint with respect to the standard inner product, and thus a kinetic energy inner product can be defined as(f,g)Mi=01(Mif)gdx, for any f and g in S.

The standard inner product on U is defined as(f,g)U=01gTfdx, for any f and g in U. It is easy to show that M is self-adjoint with respect to the inner product of (8) and a kinetic energy inner product on U is defined as (f,g)M=(Mf,g)U.

Define K̂=K+Kc. Since K is self-adjoint with respect to the standard inner product on S and Kc is a symmetric matrix, it can be shown that K̂ is self-adjoint with respect to the standard inner product on U. The positive definiteness of K̂ with respect to the standard inner product on U is determined by considering (K̂f,f)U=(Kf,f)U+(Kcf,f)U. If Kc is a positive definite matrix with respect to the standard inner product on Rn, then (Kcf,f)U0 and (Kcf,f)U=0 if and only if f=0. If each of the operators Li,i=1,2,,n is positive definite with respect to the standard inner product on Ck[0,1] then (Kf,f)U0 and (Kf,f)U=0 if and only if f=0. Thus, K̂ is positive definite with respect to the standard inner product on U if either Kc is a positive definite matrix with respect to the standard inner product on Rn or each of the operators Li,i=1,2,,n is positive definite with respect to the standard inner product on Ck[0,1]. Under either of these conditions, a potential energy inner product is defined on U by(f,g)K=(Kf,g)U. The operator K̂ is not positive definite only when the structures are unrestrained and λ0=0 and λn=0.

Define D=M-1K̂. It is possible to show that D is self-adjoint with respect to the kinetic energy inner product of (9) and the potential energy inner product of (10). If K̂ is positive definite with respect to the standard inner product, then D is positive definite with respect to both inner products.

4. Free Response

First consider the free response of the structures, F=0. A normal-mode solution is assumed asW=weiωt, where ω is a natural frequency and w=[w1(x)w2(x)w3(x)wn-1(x)wn(x)]T is a vector of mode shapes corresponding to that natural frequency. Substitution of (11) into (3) leads toM-1(K+Kc)w=ω2w, where the partial derivatives have been replaced by ordinary derivatives in the definition of K. From (12), it is clear that the natural frequencies are the square roots of the eigenvalues of D=M-1(K+Kc), and the mode shape vectors are the corresponding eigenvectors.

It is well known  that eigenvalues of a self-adjoint operator are all real and that eigenvectors corresponding to distinct eigenvalues are orthogonal with respect to the inner product for which the operator is self-adjoint. Thus, if ωi and ωj are distinct natural frequencies with corresponding mode shape vectors wi and wj, respectively, then(wi,wj)M=0,(wi,wj)K=0.

The mode shape vectors can be normalized by requiring(wi,wi)M=1, which then leads to (wi,wi)K=ωi2.

5. Forced Response

Since D is a self-adjoint operator, its eigenvectors, the mode shape vectors, can be shown to be complete in U. An expansion theorem then implies that for any f in U there exists coefficients αi, i=1,2,, such that f=iαiwi, where wi are the normalized mode shape vectors, and the summation is carried out over all modes. For a given f in U, the coefficients are calculated by αi=(f,wi)M.

Let W(t) represent the response due to the force vector F(t). Since W must be in U, the expansion theorem may be applied at any t, leading to W(x,t)=ici(t)wi(x). Substitution of (18) into (3) results in ic̈i(t)Mwi+ici(t)Kwi=F(x,t). Taking the standard inner product on U of both sides of (26) with wj(x) for an arbitrary j=1,2,,n and using mode-shape orthogonality properties of (13)–(15) leads to an uncoupled set of differential equations of the form c̈j(t)+ωj2cj(t)=(F,wj)U. A convolution integral solution of (20) iscj(t)=1ωj0t(F(x,τ),wj(x))Usin[ωj(t-τ)]dτ=1ωj0t01wjT(x)F(x,τ)sin[ωj(t-τ)]dxdτ.

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A special case occurs when the structures are uniform and operators for the individual structural elements are proportional to one another, Li=μiL1. In this case, the component of the stiffness operator due to the elasticity of the structural elements becomesK=AL1=[μ100000μ200000μ300000μ400000μn]L1.

The system may be nondimensionalized such that M1=1. Then the differential equation governing the free response of the first structural element, if isolated from the remainder of the system, isΦ̈1+L1Φ1=0. A normal-mode solution of (32) of the form Φ(x,t)=ϕ(x)eiδt leads to the eigenvalue-eigenvector problem L1ϕ=δ2ϕ. There are an infinite, but countable, number of natural frequencies for the system of (25), δ1,δ2,,δk-1,δk,δk+1, with corresponding mode shapes ϕ1,ϕ2,,ϕk-1,ϕk,ϕk+1, The mode shapes are normalized by requiring (ϕi,ϕi)S=1.

Assume a solution to (12) of the form w=aϕk(x), where a is an n×1 vector of constants. Substitution of (26) into (12) using (24) and (25) leads to M-1(δk2A+Kc)a=ω2a.

Equation (27) implies that, for this special case, the determination of the natural frequencies of the set of elastically connected structures is reduced to the determination of the eigenvalues of the matrix M-1(δk2A+Kc). For each k there are n natural frequencies. The natural frequencies can thus be indexed as ωk,j for k=1,2, and j=1,2,,n. The corresponding mode shapes are written as wk,j(x)=ak,jϕk(x), where ak,j is the eigenvector of M-1(δk2A+Kc) corresponding to the natural frequency ωk,j.

The function ϕk(x) represents the kth mode shape of a single structure. For each k, there are n natural frequencies and n corresponding mode shapes. Such a set of mode shapes, which are referred to as intramodal modes, have the same spatial behavior, but their dependence across the structures varies. Two mode shapes wk,j(x) and wp,q(x) for which pk are referred to as intermodal modes.

Note that since M and δk2A+K are both symmetric matrices, the matrix M-1(δk2A+Kc) is self-adjoint with respect to a kinetic energy inner product. The intramodal mode shapes satisfy an orthogonality condition on Rn given by (ak,i,ak,j)Mn=(ak,j)TMak,i=0for  ij. All mode shapes satisfy the orthogonality condition on U,(wk,i,wl,j)M=01(al,j)TMak,iϕk(x)ϕl(x)dx=0when  kl  or  ij.

Consider the case when Kc is singular (λ0=0 and λn=0). Then zero is the smallest eigenvalue of Kc and b is its corresponding eigenvector, b=c, where c is a normalization constant. Suppose that M-1A=I, then (27) becomes (δk2I+Kc)a=ω2a. Note that the solution of (31) corresponds to ω=δk and a=b. Thus, for this special case, the lowest natural frequency for each set of intramodal modes is equal to the natural frequency for the spatial mode of one structural element and its corresponding mode shape is the null space of Kc.

To examine the most general case when Kc is singular, take the standard inner product for Rn of both sides of (27) with b. Using properties of inner products and noting that A, Kc, and M are symmetric leads to(a,(δk2A-ω2M)b)Rn=0. Application of the orthogonality condition of (32) leads toi=1nai(δk2μi-ω2βi)=0.

The expansion theorem is used to assume a forced response of the formW(x,t)=k=1j=1nck,j(t)ak,jϕk(x). Use of (34) in (3) leads to differential equations of the formc̈k,j+ωk,j2ck,j=(F,ak,jϕk(x))U.

A special case occurs when uniform structures are identical such that A=I and M=I. Then, (27) becomes(δk2I+Kc)a=ω2a. Equation (36) can be rewritten asKca=(ω2-δk2)a. Let κj  j=1,2,,n be the eigenvalues of Kc. Then, ωk,j2=(δk2+κj). In addition, since the same set of eigenvalues is used to calculate the natural frequencies for each intramodal sets, the intramodal mode shape vectors are the same for each set and are the eigenvectors of Kc, that is ak,j=al,j=vjk,l=1,2,,j=1,2,,n, where vjj=1,2,,n are the eigenvectors of Kc.

7. Nonuniform Structures

The case when one or more of the structures is nonuniform is more difficult in that the differential equations of (12) have variable coefficients. Consider the case when the structures are nonuniform, but identical. In this case, the coupled set of differential equations can be written asILw+Kcw-ω2IMw=0, where L and M are the stiffness and inertia operators, respectively, for any of the structures.

Let κ1,κ2,,κn be the eigenvalues of Kc and let z1,z2,,zn be their corresponding eigenvectors normalized with respect to the standard inner product on Rn  (ziTzi=1). Let P be the matrix whose columns are the normalized eigenvectors. Since Kc is symmetric, it can be shown that PTP=I and PTKcP=Δ, where Δ is a diagonal matrix with Δi,j=κi. Defining w=Pq and substituting into (40) and then premultiplying by PT leads to ILq+Δq-ω2IMq=0. Equation (42) represents a set of uncoupled differential equations, each of the form Lqj+κjqj-ω2Mqj=0j=1,2,,n. Equation (43) represents, along with appropriate homogeneous boundary conditions, an eigenvalue problem to determine the natural frequencies. For each j, there are an infinite, but countable, number of natural frequencies. Thus, the natural frequencies can be indexed by ωj,kj=1,2,,n  k=1,2,.

It still may not be possible to solve (43) in closed form; however, it is now known how to index the natural frequencies. For a set of identical structures, there are an infinite number of natural frequencies corresponding to each eigenvalue of Kc. The term intramodal is not appropriate for this set of natural frequencies as they do not correspond to the same mode. Indeed, there are not necessarily intramodal frequencies for the nonuniform case.

If Kc is singular and thus has zero as its lowest eigenvalue, then (43) shows that one set of natural frequencies is identical to the natural frequencies of the individual structures.

8. Examples8.1. Shafts

Consider n concentric shafts of equal length connected by elastic layers. The stiffness and inertia operators for uniform shafts are, respectively,Liθi=-μi2θix2,Miθi=βiθi.

The stiffness operators are of the form of those considered in Section 6. Hence, the natural frequencies and mode shapes can be determined from solving a matrix eigenvalue problem. If all shafts are made of the same material then βi=μi leading to M-1K=I, and thus if Kc is singular then the lowest natural frequency for each set of intramodal modes is the same as the modal natural frequency for the innermost shaft, and the mode shapes corresponding to each frequency are the eigenvector of Kc corresponding to its zero eigenvalue times, the spatial mode shape for the shaft.

The eigenvalue problem for the innermost shaft is ϕ′′(x)+δ2ϕ(x)=0, subject to appropriate boundary conditions. The values for δkk=1,2,,5 and the corresponding normalized mode shapes ϕk(x) are listed in Table 1 for various end conditions.

First five sets of intramodal natural frequencies of four elastically connected fixed free shafts, ω(k,j).

 j k 1 2 3 4 5 1 1.5708 4.7124 7.8540 10.9956 14.1372 2 1.5763 4.7142 7.8551 10.9964 14.1378 3 1.5965 4.7210 7.8592 10.9993 14.1400 4 1.8811 4.8247 7.9219 11.0442 14.1750

The matrix eigenvalue problem for a set of intramodal frequencies is of the form Ka=ω2Ma, where K is a tridiagonal matrix whose elements areki,i=μiδ2+λi-1+λii=1,2,,n,ki,i-1=-λi-1i=2,3,,n,ki,i+1=-λii=1,2,,n-1, and M is a diagonal matrix with mi,i=βi.

8.2. Natural Frequencies of Four Concentric Fixed-Free Shafts

As a numerical example, consider four concentric fixed-free shafts connected by layers of torsional stiffness. Solving (45) subject to ϕk(0)=0 and ϕk(1)=0, δk=(2k-1)π2ϕk(x)=2sin[(2k-1)π2x]. Each shaft is made of the same material. The inner shaft is solid of radius r. The outer shafts are each of thickness r. The thickness of each elastic layer is negligible. This leads to μ1=1μ2=15μ3=65,  μ4=175,  β1=1β2=15β3=65,  β4=175. The torsional stiffness of each elastic layer is the same and is taken such that λi=1. The inner shaft is solid, thus, λ0=0. The outer radius of the outer shaft is unrestrained from rotation, hence, λ4=0. The matrix eigenvalue problems become[1+((2k-1)π2)2-100-12+15((2k-1)π2)2-100-12+65((2k-1)π2)2-100-11+175((2k-1)π2)2][ak,1ak,2ak,3ak,4]=ω2[ak,1ak,2ak,3ak,4].The natural frequencies for k=1,2,,5 are given in Table 1. Since Kc is singular and M-1Kb=I, the lowest natural frequency in each intramodal set is δk. Each mode shape in a set of intramodal mode shapes corresponds to the same spatial mode ϕk(x). The difference in intramodal mode shapes is in the relative magnitude and signs of the displacements of the individual shafts. The normalized mode shapes of Figure 2 correspond to the first mode shape in the intramodal set for the first spatial mode and illustrate the mode in which the shafts rotate as if they are rigidly connected. The mode shapes of Figure 3 correspond to the third intramodal mode for the first spatial mode and illustrate that when the rotations of the first, second, and fourth shafts are counterclockwise, the rotation of the third shaft is clockwise. Figures 4 and 5 illustrate mode shapes corresponding to the third spatial mode. All mode shapes in the intramodal set for this mode have two nodes across the length of the shaft. Note that for the second intramodal frequency there is a change in the direction of rotation of the shafts between the third and fourth shafts. Thus, there is a cylindrical surface of nodes between these shafts. There are two changes in the direction of rotation for the third intramodal frequency, between the second and third shafts and between the third and fourth shafts, leading to two cylindrical surfaces of nodes.

Set of intramodal mode shapes of elastically connected fixed-free torsional shafts with k=1 and j=1. The mode shapes correspond to rigid-body motion across the set of shafts.

Set of intramodal mode shapes of elastically connected fixed-free torsional shafts with k=1 and j=3.

Set of intramodal mode shapes of elastically connected fixed-free torsional shafts with k=3 and j=2.

Set of intramodal mode shapes of elastically connected fixed-free torsional shafts with k=3 and j=3.

8.3. Forced Response of Four Concentric Shafts

Suppose that the midspan of the outer shaft is subject to a constant torque, T0, such that the nondimensional applied torques are T1(x,t)=T2(x,t)=T3(x,t)=0,T4(x,t)=δ(x-1/2). The forced response of the system is calculated by using a convolution integral solution of the form of (21) leading to ck,j(t)=1ωk,j0t01ak,jTF(x,τ)ϕk(x)sin[ωk,j(t-τ)]dxdτ=(ak,j)4ϕk(1/2)ωk,j2[1-cos(ωk,jt)]. Substitution of (50) into (18) leads to W(x,t)=k=1(j=14(ak,j)4ωk,j2ak,j[1-cos(ωk,jt)])ϕk(12)ϕk(x). Equation (51) is evaluated leading to the time dependence of the response at x=1/2 and x=1 illustrated in Figures 6 and 7.

Forced response of elastically connected torsional shafts at x=0.5 due to constant concentrated torque applied to outer shaft at x=0.5.

Forced response of elastically connected torsional shafts at x=1 due to constant concentrated torque applied to outer shaft at x=0.5.

8.4. Nonuniform Shafts

Now consider the same set of shafts, except that each has a taper, such that the differential equation for the innermost shaft when isolated from the system isddx[(1-0.1x)2dθdx]+ω2(1-0.1x)2θ=0. Along with the boundary for a fixed-free shaft, (52) has a Bessel function solution leading to the characteristic equation for the shaft’s natural frequencies asj0(9ω)yo(10ω)-y0(9ω)j0(10ω)=0, where jn(x) and yn(x) are spherical Bessel functions of the first and second kinds of order n and argument x. The mode shape (nonnormalized) corresponding to a natural frequency ωk isθk(x)=jo[10ωk(1-0.1x)]-j0[9ωk]y0[10ωk(1-0.1x)].

The differential equations for the elastically coupled shafts becomeddx[(1-0.1x)2ddx][θ1θ2θ3θ4]-[1-100-12-100-12-100-11][θ1θ2θ3θ4]+ω2×(1-0.1x)2[θ1θ2θ3θ4]=. Even though the shafts are not identical, the differential equations in (55) may still be decoupled because the stiffness and inertia matrices are the same. The eigenvalues of M-1Kc are κ1=0,κ2=0.0172,κ3=0.0815, and κ4=1.0712. The corresponding matrix of eigenvectors isP=[0.06250.11910.22920.96400.06250.11710.2106-0.06860.06250.0848-0.06540.00100.0625-0.04220.0049-5.44×10-6]. The columns of P have been normalized such that PTMP=I and PTKcP=Δ. Following the same procedure as in the derivation of (43), the uncoupled differential equations becomeddx[(1-0.1x)2dq1dx]+ω2(1-0.1x)2q1=0,ddx[(1-0.1x)2dq2dx]+[ω2(1-0.1x)2-0.0172]q2=0,ddx[(1-0.1x)2dq3dx]+[ω2(1-0.1x)2-0.0815]q3=0,ddx[(1-0.1x)2dq4dx]+[ω2(1-0.1x)2-1.0712]q4=0. The solutions of (57) areq1(x)=C1j0[ω(1-0.1x)]+C2y0[ω(1-0.1x)],q2(x)=C1j0.904[ω(1-0.1x)]+C2y0.904[ω(1-0.1x)],q3(x)=C1j2.398[ω(1-0.1x)]+C2y2.398[ω(1-0.1x)],q4(x)=C1j9.862[ω(1-0.1x)]+C2y9.862[ω(1-0.1x)]. The characteristic equations to determine the natural frequencies arej0(10ωk,1)y1(9ωk,1)-y0(10ω1)j1(9ωk,1)=0,j0,904(10ωk,2)[0.904y0.904(9ωk,2)9ωk,2-y1.904(9ωk,2)]-y0,904(10ωk,2)[0.904j0.904(9ωk,2)9ωk,2-j1.904(9ωk,2)]=0,j2.398(10ωk,3)[2.398y2.398(9ωk,3)9ωk,3-y3.398(9ωk,3)]-y2.398(10ωk,3)[2.398j2.398(9ωk,3)9ωk,3-j3.398(9ωk,3)]=0,j9.862(10ωk,4)[9.862y9.862(9ωk,4)9ωk,4-y10.862(9ωk,4)]-y9.862(10ωk,4)[9.862j9.862(9ωk,4)9ωk,4-j10.962(9ωk,4)]=0. The characteristic equation for the first set of frequencies is identical to (53). The first five frequencies for each j are given in Table 2.

First five sets of frequencies for set of linearly tapered shafts.

 j k 1 2 3 4 5 1 1.639 4.736 7.868 11.006 14.145 2 1.645 4.738 7.869 11.007 14.146 3 1.667 4.745 7.874 11.010 14.148 4 1.981 4.861 7.944 11.010 14.187
8.5. Euler-Bernoulli Beams

Consider a set of n parallel Euler-Bernoulli beams connected by elastic layers. For uniform beams, the mass and stiffness operators are Li=μi(4/x4) and Mi=βi. The differential eigenvalue problem for the first beam in the set is d4ϕdx4-δ2ϕ=0. The transverse displacements of Euler-Bernoulli beams is another example of the special case discussed in Section 6.

If the beams are identical (μi=1 and βi=1) and if Kc is singular, then the lowest natural frequency for the kth set of intramodal modes is δk with the mode shape, such that each beam has the same displacement and the springs are unstrectched. Otherwise, the mode shape for the lowest natural frequency of each intramodal set satisfies the orthogonality condition of (31).

As a numerical example, consider a set of five fixed-free elastically connected Euler-Bernoulli beams. The solution of Equation (60) subject to the boundary conditions ϕ(0)=0, ϕ(0)=0, ϕ′′(1)=0, and ϕ′′′(1)=0 leads to ϕk(x)=cosh(δkx)-cos(δkx)-cos(δk)+cosh(δk)sin(δk)+sinh(δk)×[sinh(δkx)-sin(δkx)], where δk is the kth solution of cos(δk)cosh(δk)=-1. Numerical values used in the computations are μ1=1, μ2=2, μ3=0.5, μ4=1, μ5=0.25, β1=1, β2=1.5, β3=0.75, β4=1, β5=0.5, λ0=0, λ1=100, λ2=50, λ3=50, λ4=20, and λ5=0

Using these numerical values, the matrix eigenvalue problem for a set of intramodal frequencies and mode shapes is[δk2+100-100000-1002δk2+150-50000-500.5δk2+100-50000-50δk2+70-20000-200.25δk2+20]×[a1a2a3a4a5]=ω2[1000001.5000000.75000001000000.5][a1a2a3a4a5].

The sets of intramodal frequencies are listed in Table 3. Recall that δk is the natural frequency of the first beam. The natural frequency of an Euler-Bernoulli beam increases with stiffness. Since the first beam is stiffer than several other beams in the set, some intramodal frequencies are lower than δk. The mode shapes for k=2 and j=4 illustrated in Figure 8 show one spatial node and three cylindrical surfaces of nodes.

First four sets of intramodal natural frequencies for a set of five elastically connected fixed-fixed Euler-Bernoulli beams.

 j δk 3.5100 22.0300 61.7000 120.90 1 3.4630 16.7195 44.0798 85.7222 2 6.6167 20.6044 51.6490 99.3829 3 8.5529 23.4906 62.2901 121.1925 4 12.7319 24.1708 62.4633 121.3072 5 14.8317 28.0564 71.9859 139.9667

Intramodal mode shapes for a set of five elastically connected Euler-Bernoulli beams.

9. Conclusion

A general theory is developed for the free and forced response of elastically connected structures. The following has been shown.

The general problem can be formulated using the vector space U=Rn×S, where S is a subspace of Ck[0,1] defined by the system’s end conditions.

If the differential stiffness operator for a single structure is self-adjoint with respect to a standard inner product on S, then the general stiffness operator is self-adjoint with respect to a standard inner product on U.

Kinetic and potential energy inner products are defined on both S and U.

A normal-mode solution for the free response leads to the formulation of an eigenvalue problem defined for a matrix of operators.

The operator is self-adjoint with respect to the energy inner products leading to the development of an orthogonality condition.

The expansion theorem is used to develop a modal analysis for the forced response.

The case where the structures are uniform and the individual stiffness operators are proportional is a special case in which the determination of natural frequencies and mode shapes can be reduced to eigenvalue problems for matrices on Rn.

When the stiffness operators are proportional, the natural frequencies and mode shapes are indexed with two indices, the first representing the spatial mode shape, the second representing the intramodal mode shapes.

If the uniform structures are identical, then a simple formula can be derived for the sets of intramodal natural frequencies using the eigenvalues of the coupling stiffness matrix. The intramodal mode shapes for each spatial mode are the eigenvectors of the coupling stiffness martrix.

An iterative solution must be applied to determine the natural frequencies for the most general case of the uniform structure.

The differential equations for the coupling of identical structures, uniform, or nonuniform can be uncoupled through diagonalization of the coupling stiffness matrix.

Elastically connected uniform strings and elastically connected uniform concentric shafts are applications in which the stiffness operators are proportional.

The differential equations for the concentric shafts, even though they are not identical, can be decoupled when each individual stiffness operator is the same as the individual mass operator.

The individual stiffness operators for uniform Euler-Bernoulli beams are proportional implying that their natural frequencies can be indexed as an infinite number of sets of intramodal frequencies.

The general method is applied here only for undamped systems. However, it can be applied to certain damped systems as well. If the structures are undamped but the Winkler layers have viscous damping, the same δk and ϕk for the undamped system may be used, but an eigenvalue problem is obtained involving complex numbers. If the structures are damped but the Winkler layers are undamped, the choice of δk and ϕk is modified to include viscous damping, but again a complex eigenvalue problem is obtained. If the entire system (both the individual structures and the Winkler layers) is subject to proportional damping, the eigenvalues and the eigenvectors of the undamped system can be used to uncouple the forced vibrations equations.

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