An analytical model for dynamic simulation of the composite rotor with internal damping
Yongsheng Ren^{1} , Xingqi Zhang^{2} , Yanghang Liu^{3} , Xiulong Chen^{4}
^{1, 2, 3, 4}College of Mechanical and Electronic Engineering, Shandong University of Science and Technology, Qindao, 266590, China
^{1}Corresponding author
Journal of Vibroengineering, Vol. 16, Issue 8, 2014, p. 40024016.
Received 3 September 2014; received in revised form 4 November 2014; accepted 19 November 2014; published 30 December 2014
JVE Conferences
A theoretical model for the dynamics of composite rotor is presented. The composite shaft that includes rigid disks and is supported on rigid bearings is considered as a thinwalled EulerBernoulli beam. Internal damping of the composite shaft is taken into account. The equations of motion are derived using the thinwalled composite beam theory based on variational asymptotic method and Hamilton’s principle. The internal damping of shaft is introduced by adopting the multiscale damping analysis method. Galerkin’s method is used to discretize and solve the governing equations. To demonstrate the validity of the present model, the convergence of the method is examined and the results are compared with those available in the literature. Numerical study shows the effect of design parameters on the natural frequencies, critical rotating speeds and instability thresholds of composite shaft. In addition, the free vibration responses due to the initial perturbations and the forced responses to unbalance for composite shaft are also presented.
Keywords: composite shaft, internal damping, vibration and stability.
1. Introduction
The rotating composite shaft is the primary component for helicopter power transmission and automotive drive system. This can be attributed to the improved properties of composites, such as high strength and stiffness as well as lightweight, compared to metals. On the other hand, composites have excellent damping characteristics, which make them play an important role on damping vibration of laminated plate and shelllike nonrotating structures [1]. Unfortunately, as the rotating speed is in the supercritical range, such damping may cause whirl instabilities of composite shaft. In the field of rotor dynamics, the damping in the composite shaft is called internal damping, which is also called rotating damping. Bucciarelli [2] demonstrated that increasing internal damping may reduce the instability threshold, whereas, the addition of external damping can improve rotor stability. Internal damping characteristics is very important for dynamical stability of composite shaft. Moreover, the destabilizing effect of internal damping can be easily influenced by the arrangement of the different composite layers: fiber orientation angles and number of layers. With the increasing demand for high speed composite rotors, the study of composite shaft based on accurate prediction of internal damping is becoming fundamental as it will provide more reasonable and safe rotor design.
The study of rotating shaft with internal damping was first presented by Kimball [3]. His’s results exhibited that internal damping causes shaft whirling above the first critical speed. Gunter [4] illustrated how above the first critical speed, internal viscous damping leads to instability in the rotating shaft. Vance and Lee [5] studied the stability of high speed rotors with internal viscous damping. The model they used included a single unbalance disk connected with a flexible shaft. The stability was predicted using mathematical methods and verified by numerical solutions of the equation of motion. Melanson and Zu [6] performed vibration analysis of an internally damped rotating shaft, modeled using Timoshenko shaft theory. Explicit analytical expressions for the complex frequencies and mode shapes were presented and the stability threshold was also determined. Montagnier and Hochard [7] studied dynamic instability of supercritical driveshaft with internal damping based on EulerBernoulli shaft model. The model taken into account the effect of translational and rotatory inertia. Simon and Flowers [8] presented an investigation of using adaptive control method to suppress instability and synchronous disturbance of a flexible shaft with internal viscous damping. However, all these foregoing literatures deal with only rotating metal shaft.
In recent years, dynamic analysis of the composite shaft with internal damping has also received some attention [911]. Singh and Gupta [9] presented the rotordynamic formulations which are applicable to the analysis of the composite shaft with internal damping, obtained by using a layerwise beam theory. Kim et al. [10] conducted the forced vibration analysis of a rotating tapered composite shaft. The stability of a rotating shaft driven through a universal joint has been investigated by Mazzei and Scott [11]. In the above studies, an equivalent viscous damping coefficient was introduced to account for the effect of composite internal damping, but no internal damping analysis has been proposed. To our best knowledge, very few models which incorporate damping estimations are available for predicting the dynamic behavior of composite shaft. Sino et al. [12] presented a dynamic analysis of composite rotating shaft. Internal damping was introduced by the complex constitutive relation of a viscoelastic composite.
The internal damping of composite shaft mainly comes from the dissipation of energy in the materials. Saravanos et al. [13] presented a hollow tubular beam finite element model to predict the damping properties of anisotropic thinwalled closedsection structures. However, this model is only applicable to nonrotating composite beams or blades.
An analytical model of rotating composite thinwalled shaft including internal damping is proposed. This model is based on the composite thinwalled beam theory, referred to as variational asymptotically method (VAM) by Berdichevsky et al. [14]. The internal damping of shaft is introduced via the multiscale damping mechanics [15]. The flexible composite shaft is assumed supporting on on rigid bearings and containing of the rigid disks mounted on it. The equations of motion of the composite shaft are derived by Hamilton’s principle. Galerkin’s method is used then to discretize and solve the governing equations. The natural frequencies, critical rotating speeds, instability thresholds are obtained through numerical simulations. The effect of the ply angle is then assessed. The validity of the model is proved by comparing the results with those in literatures and convergence examination. Finally, the transient free vibration responses due to the initial perturbations and the steady state forced responses to unbalance are also obtained by using the timeintegration method and the mode superposition method, respectively.
2. Composite shaft
The slender thinwalled composite shaft subjected to a rotation along its longitudinal $x$axis at a constant rotating speed $\mathrm{\Omega}$ is shown in Fig. 1. The length, wall thickness and radius of the shaft are denoted by $L$, $h$ and $r$, respectively. To describe the motion of the shaft the following coordinate systems are used: (I) $(X,Y,Z)\mathrm{}$is inertial reference system whose origin$\mathrm{}O$ is located in the geometric center, (II) $(x,y,z)$ is rotating reference system with the common origin$\mathrm{}{\rm O}$. $(I,J,K)$ and $(i,j,k)$ denote the unit vectors of the reference system $(X,Y,Z)$ and $(x,y,z)$, respectively. In addition, a local coordinate system $(\varsigma ,s,x)$ is used, where $\varsigma $ and $s$ are measured along the direction normal and tangent to the middle surface, respectively.
The variation of the strain energy of the crosssection can be expressed as:
where ${\mathbf{\epsilon}}_{x,y}^{T}=\left\{\begin{array}{ccc}{\gamma}_{11}& {\gamma}_{12}& {2\gamma}_{12}\end{array}\right\}$, ${\gamma}_{\alpha \beta}$$\left(\alpha ,\beta =\text{1,2}\right)$ are the inplane tensorial strain components, $A$ is the crosssectional area of the shaft and ${\stackrel{}{\mathbf{Q}}}_{ij}$ equivalent offaxis stiffness matrix of a composite ply with respect to the system $osx\varsigma $.
The variation of the dissipated energy of the crosssection can be expressed as:
where ${\stackrel{}{\mathit{\eta}}}_{ij}$ is equivalent offaxis damping matrix of a composite ply with respect to the system $osx\varsigma $.
Fig. 1. Geometry and coordinate systems of composite thinwalled shaft
Further, the variation of the kinetic energy of the crosssection with rotating motion is:
where $\mathrm{d}\mathrm{i}\mathrm{a}\mathrm{g}\left(\rho \right)\mathrm{}$is a diagonal matrix with components equal to the mass density $\rho $ of a ply.
The position, velocity and acceleration vectors for the deformed shaft are described as:
$\dot{\mathbf{R}}=\left[{\dot{u}}_{2}\mathrm{\Omega}\left(z+{u}_{3}\right)\right]\mathbf{i}+\left[{\dot{u}}_{3}+\mathrm{\Omega}\left(y+{u}_{2}\right)\right]\mathbf{j}+{\dot{u}}_{1}\mathbf{k},$
$\ddot{\mathbf{R}}=[{\ddot{u}}_{2}2\mathrm{\Omega}{\dot{u}}_{3}{\mathrm{\Omega}}^{2}(y+{u}_{2}\left)\right]\mathbf{i}+[{\ddot{u}}_{3}+2\mathrm{\Omega}{\dot{u}}_{2}{\mathrm{\Omega}}^{2}(z+{u}_{3}\left)\right]\mathbf{j}+{\ddot{u}}_{1}\mathbf{k}$,
where ${u}_{1}$, ${u}_{2}$ and ${u}_{3}$ are the displacements of any point on the the crosssection of the shaft in the $x$, $y$ and $z$ directions, dot indicates the time derivative of the variable.
2.1. Equivalent crosssection stiffness matrix
To derive the strain energy and dissipated energy of the crosssection, the composite thinwalled beam theory proposed by Berdichevsky et al. [14] is employed in the following.
For the case of no internal pressure acting on the shaft, Eq. (1) can be simplified by using free hoop stress resultant assumption as:
where $\oint (\cdot )$ denotes the integral around the loop of the midline crosssection, and the reduced axial, coupling and shear stiffness $A$, $B$ and $C\mathrm{}$can be written as:
$C\left(s\right)=4\left[{A}_{66}\frac{{A}_{26}^{2}}{{A}_{22}}\right],{A}_{ij}=\sum _{k=1}^{N}{\stackrel{}{Q}}_{ij}^{\left(k\right)}\left({z}_{k}{z}_{k1}\right),\mathrm{}\mathrm{}\mathrm{}i,\mathrm{}j=1,\mathrm{}2,\mathrm{}6.$
In Eq. (5), $\mathrm{\delta}{U}_{s}$ can also be expressed with respect to the $u$, $v$, $w$ and $\phi $ as:
where $\mathrm{\Delta}$ is 4×1 column matrix of kinematic variables defined as ${\mathrm{\Delta}}^{T}=(\begin{array}{cccc}{u}^{\text{'}}& {\phi}^{\text{'}}& {w}^{\text{'}\text{'}}& {v}^{\text{'}\text{'}})\end{array}$, and $\mathbf{K}\mathrm{}$ is 4×4 symmetric stiffness matrix. Its components ${k}_{ij}$ are given in Ref. [14].
2.2. Equivalent crosssection damping matrix
Similar to the derivation of the previous crosssection stiffness formulations, the variation of the dissipated energy of the crosssection [15]:
where:
${B}_{d}={A}_{d16}+{A}_{d61}+2\left(\frac{{A}_{12}{A}_{26}}{{A}_{22}^{2}}\right){A}_{d22}\left(\frac{{A}_{26}}{{A}_{22}}\right)\left({A}_{d12}+{A}_{d21}\right)\left(\frac{{A}_{12}}{{A}_{22}}\right)\left({A}_{d26}+{A}_{d62}\right),$
${\mathrm{C}}_{d}=4\left[{A}_{d66}+\left(\frac{{A}_{26}^{2}}{{A}_{22}^{2}}\right){A}_{d22}\left(\frac{{A}_{26}}{{A}_{22}}\right)({A}_{d26}+{A}_{d62})\right],$
${A}_{dij}={\int}_{h/2}^{h/2}{\stackrel{}{\psi}}_{il}{\stackrel{}{Q}}_{lj}d\varsigma =2\sum _{k=1}^{N/2}{\stackrel{}{\psi}}_{il}^{k}{\stackrel{}{Q}}_{lj}^{k}\left({h}_{k}{h}_{k1}\right),i,j,l=1,2,6.$
The variation of the dissipated energy also be expressed in terms of the kinematic variables as:
where $\mathbf{C}$ is 4×4 symmetric damping matrix.
The formulation of its components ${c}_{ij}$is analogous to stiffness components ${k}_{ij}$, but the terms $A$, $B$ and $C$ in Eq. (6) should be replaced by the terms ${A}_{d}$, ${B}_{d}$ and ${C}_{d}$ in Eq. (9), respectively.
2.3. Equivalent crosssection damping matrix
Based on the displacement expressions presented in Ref. [14] and in view of Eqs. (3) and (4), the variation of the kinetic energy of the crosssection can be obtained as:
where:
${I}_{3}={b}_{1}\left(\ddot{w}+2\mathrm{\Omega}\dot{v}{\mathrm{\Omega}}^{2}w\right)+{b}_{2}\left(\ddot{\phi}{\mathrm{\Omega}}^{2}\phi \right){b}_{3}\left(2\mathrm{\Omega}\dot{\phi}+{\mathrm{\Omega}}^{2}\right),$
${I}_{4}={b}_{2}\left(\ddot{w}+2\mathrm{\Omega}\dot{v}{\mathrm{\Omega}}^{2}w\right){b}_{3}\left(\ddot{v}2\mathrm{\Omega}\dot{w}{\mathrm{\Omega}}^{2}v\right)+\left({b}_{4}+{b}_{5}\right)\left(\ddot{\phi}{\mathrm{\Omega}}^{2}\phi \right),$
${b}_{1}={\int}_{A}^{}\rho dsd\varsigma ,{b}_{2}={\int}_{A}^{}\rho ydsd\varsigma ,{b}_{3}={\int}_{A}^{}\rho zdsd\varsigma ,$
${b}_{4}={\int}_{A}^{}\rho {y}^{2}dsd\varsigma ,{b}_{5}={\int}_{A}^{}\rho {z}^{2}dsd\varsigma .$
3. Rigid disks
3.1. Kinetic energy of rigid disks
According to Eqs. (11) and (12) the expression of variation of the kinetic energy of the rigid disks fixed to the shaft is written as:
in which:
${I}_{3l}^{D}={b}_{1l}^{D}\left(\ddot{w}+2\mathrm{\Omega}\dot{v}{\mathrm{\Omega}}^{2}w\right)+{b}_{2l}^{D}\left(\ddot{\phi}{\mathrm{\Omega}}^{2}\phi \right){b}_{3l}^{D}\left(2\mathrm{\Omega}\dot{\phi}+{\mathrm{\Omega}}^{2}\right),$
${I}_{4l}^{D}={b}_{2l}^{D}\left(\ddot{w}+2\mathrm{\Omega}\dot{v}{\mathrm{\Omega}}^{2}w\right){b}_{3l}^{D}\left(\ddot{v}2\mathrm{\Omega}\dot{w}{\mathrm{\Omega}}^{2}v\right)+\left({b}_{4l}^{D}+{b}_{5l}^{D}\right)\left(\ddot{\phi}{\mathrm{\Omega}}^{2}\phi \right),$
${b}_{1l}^{D}={\int}_{A}^{}{\rho}_{l}^{D}dsd\varsigma ,{b}_{2l}^{D}={\int}_{A}^{}{\rho}_{l}^{D}ydsd\varsigma ,{b}_{3l}^{D}={\int}_{A}^{}{\rho}_{l}^{D}zdsd\varsigma ,$
${b}_{4l}^{D}={\int}_{A}^{}{\rho}_{l}^{D}{y}^{2}dsd\varsigma ,{b}_{5l}^{D}={\int}_{A}^{}{\rho}_{l}^{D}{z}^{2}dsd\varsigma ,$
where, ${\rho}_{l}^{D}$ denotes the mass density of the disk $l$. The symbol $\mathrm{\Delta}(x{x}_{Dl})$ denotes Dirac delta function, ${N}_{D}$ the number of the disks mounted on the shaft and ${x}_{Dl}$ the location of the $l$th disk.
3.2. Work of external forces
The external forces include the centrifugal force resulting from the unbalanced rigid disks. The virtual work done by the centrifugal force of rigid disks can be given by:
where, ${p}_{x}^{u}$, ${p}_{y}^{u}$ and ${p}_{z}^{u}$ denote centrifugal force per unit length, ${G}_{x}^{u}$ denotes centrifugal moment per unit length. They are given by:
${p}_{z}^{u}={\mathrm{\Omega}}^{2}\mathrm{sin}\mathrm{\Omega}t\sum _{l=1}^{{N}_{D}}{b}_{1l}^{D}{e}_{l}\mathrm{\Delta}\left(x{x}_{Dl}\right),{G}_{x}^{u}=0,$
where, the mass unbalance associated with disks is assumed as the concentrated masses ${b}_{1l}^{D}={\iint}_{A}^{}{\rho}_{l}^{D}dA$ at points with small distances of eccentricity ${e}_{l}$.
4. Equations of motion and approximate solution method
The equations of motion of the composite rotor can be derived by employing the following Hamilton’s principle:
In order to find the approximate solution, the quantities $u\left(x,t\right)$, $v\left(x,t\right)$, $w(x,t)$ and $\phi (x,t)$ are assumed in the form:
$v\left(x,t\right)=\sum _{j=1}^{N}{X}_{jv}\left(t\right){\psi}_{j}\left(x\right),w\left(x,t\right)=\sum _{j=1}^{N}{X}_{jw}\left(t\right){\psi}_{j}\left(x\right),$
where ${\alpha}_{j}\left(x\right)$, ${\theta}_{j}\left(x\right)$ and ${\psi}_{j}\left(x\right)$ are mode shape functions which fulfill all the boundary conditions of the composite shaft, ${X}_{ju}\left(t\right)$, ${X}_{j\phi}\left(t\right)$, ${X}_{jv}\left(t\right)$ and ${X}_{jw}\left(t\right)$ are the generalized coordinates.
Substituting Eq. (18) into the governing equations of motion, and applying Galerkin’s procedure, the following governing equations in matrix form can be found:
where ${X}^{T}=\left({X}_{1u}\left(t\right),\dots ,{X}_{Nu}\left(t\right),{X}_{1\phi}\left(t\right),\dots ,{X}_{N\phi}\left(t\right),{X}_{1w}\left(t\right),\dots ,{X}_{Nw}\left(t\right),{X}_{1v}\left(t\right),\dots ,{X}_{Nv}\left(t\right)\right)\text{,}$$\mathbf{M}$ is the mass matrix, $\mathbf{C}$ is the damping matrix, $\mathbf{G}$ is the gyroscopic matrix, $\mathbf{K}$ the stiffness matrix. $\mathbf{H}$ is the rotating softening matrix, $\mathbf{f}$ is the generalized force vector. Their detailed expressions are follows:
where:
${I}_{ij}={\int}_{0}^{L}{\theta}_{i}{\alpha}_{j}^{\text{'}\text{'}}dx,{J}_{ij}={\int}_{0}^{L}{\theta}_{i}{\theta}_{j}^{\text{'}\text{'}}dx,{K}_{ij}={\int}_{0}^{L}{\theta}_{i}{\psi}_{j}^{\text{'}\text{'}\text{'}}dx,{L}_{ij}={\int}_{0}^{L}{\theta}_{i}{\theta}_{j}dx,$
${M}_{ij}={\int}_{0}^{L}{\theta}_{i}{\psi}_{j}dx,{N}_{ij}={\int}_{0}^{L}{\psi}_{i}{\alpha}_{j}^{\text{'}\text{'}\text{'}}dx,{O}_{ij}={\int}_{0}^{L}{\psi}_{i}{\theta}_{j}^{\text{'}\text{'}\text{'}}dx,{P}_{ij}={\int}_{0}^{L}{\psi}_{i}{\psi}_{j}^{\text{'}\text{'}\text{'}\text{'}}dx,$
${Q}_{ij}={\int}_{0}^{L}{\psi}_{i}{\theta}_{j}dx,{R}_{ij}={\int}_{0}^{L}{\psi}_{i}{\psi}_{j}dx,{S}_{ij}={\int}_{0}^{L}{\psi}_{i}{\psi}_{j}^{\text{'}\text{'}}dx,{R}_{ij}^{Dl}={\int}_{0}^{L}{\psi}_{i}{\psi}_{j}\mathrm{\Delta}\left(x{x}_{Dl}\right)dx,$
${S}_{ij}^{Dl}={\int}_{0}^{L}{\psi}_{i}{\psi}_{j}^{\text{'}\text{'}}\mathrm{\Delta}\left(x{x}_{Dl}\right)dx,{B}_{i}^{Dl}={\int}_{0}^{L}{\psi}_{i}\mathrm{\Delta}\left(x{x}_{Dl}\right)dx,i,j=1,\dots ,N.$
The solution for free vibration of Eq. (19) can be written in the form $\left\{X\right\}=\left\{{X}_{0}\right\}{e}^{\lambda t}$, from Eq. (19), one can obtain the following characteristic equation:
Complex eigenvalue $\lambda $ can be expressed in the form:
The damping natural frequency or whirl frequency of the system is the imaginary part $\omega $, whereas its real part $\sigma $ gives the decay or growth of the amplitude of vibration. A negative value of $\sigma $ indicates a stable motion, whereas a positive value indicates an unstable motion, growing exponentially in time.
It should be pointed out that the govering Eq. (19) is derived based on the rotating frame $Oxyz$. In addition, for a CUS (Circumferentially Uniform Stiffness) configuration [16], Eq. (19) involving in terms of displacements can be split into two independent equation systems associated with both flapping bendingsweeping bending and extensiontwist motions. The latter is important to helicopter blades. However, we will focus on the flapping bendingsweeping bending coupling only and give all numerical examples with reference to the inertial frame $OXYZ$ in this paper, the resulting govering equation are not shown for the sake of simplicity.
5. Numerical results
5.1. Model verifications
In order to examine the influence of the number of mode shape functions used in the solution of the equation on the accuracy of the results, the numerical calculations are performed by considering the shaft made of graphiteepoxy whose elastic characteristics are listed in Table 1. The shaft has rectangular crosssection of fixed width $a=$ 0.32 m, wall thickness $h=$0.01016 m and length $L=$4.5952 m, whose layup is ${\left[\theta \right]}_{16}$ with clampedfree boundary conditions.
Table 1. Mechanical properties of composite material [13]
$\rho $ (kg/m^{3})

${E}_{11}$ (GPa)

${E}_{22}$ (GPa)

${G}_{12}$ (GPa)

${\upsilon}_{12}$

${\eta}_{l1}$ (%)

${\eta}_{l2}$ (%)

${\eta}_{l6}$ (%)

1672

25.8

8.7

3.5

0.34

0.65

2.34

2.89

The numerical results of natural frequencies and modal dampings are shown in Tables 2 and 3 for an increasing number of mode shape functions. From these tables, it can be seen that to obtain the accurate results of the first two natural frequencies and dampings of flapping, sweeping and torsional mode no more than five mode shape functions are required. This indicates clearly that the convergence of the present model is quite good.
A comparison of predictions using the present model with those obtained in Ref. [13] are also shown in Table 2 and Table 3. A perfect agreement of numerical results with those in Ref. [13] can be seen.
Table 2. Modal frequencies and damping of cantilever composite box beam: $L/a=$ 14.36, $a/b=$ 5, [90]_{16}
Mode

Natural frequency (Hz)

Damping (%)


Present

Ref. [13]

Present

Ref. [13]


$N=$ 1

$N=$ 3

$N=$ 5

$N=$ 1

$N=$ 3

$N=$ 5


First flapping

1.80

1.81

1.81

1.80

2.39

2.38

2.38

2.35

Second flapping

–

11.36

11.36

11.5

–

2.37

2.37

2.35

First sweeping

5.67

6.20

6.20

6.50

3.03

2.53

2.53

2.35

Second sweeping

–

39.21

39.21

39.7

–

2.36

2.36

2.37

First torsional

37.83

37.84

37.84

37.7

2.89

2.89

2.89

2.89

Second torsional

112.90

113.01

113.01

113.3

2.82

2.92

2.92

2.89

Table 3. Modal frequencies and damping of cantilever composite box beam: $L/a=$ 14.36, $a/b=$5, [0]_{16}
Mode

Natural frequency (Hz)

Damping (%)


Present

Ref. [13]

Present

Ref. [13]


$N=$ 1

$N=$ 3

$N=$ 5

$N=$ 1

$N=$ 3

$N=$ 5


First flapping

3.12

3.13

3.13

3.10

0.64

0.65

0.65

0.65

Second flapping

–

19.02

19.02

19.8

–

0.69

0.69

0.67

First sweeping

10.80

10.81

10.81

11.0

0.68

0.68

0.68

0.68

Second sweeping

–

68.38

68.38

65.6

–

0.85

0.85

0.90

First torsional

37.83

37.84

37.84

37.7

2.89

2.89

2.89

2.89

Second torsional

112.91

113.01

113.01

113.3

2.33

2.92

2.92

2.89

Table 4. Comparison of the nondimensional natural frequencies of a rotating cantilever composite shaft
${\mathrm{\Omega}}^{*}$

${\omega}_{1}^{*}$

${\omega}_{2}^{*}$

${\omega}_{3}^{*}$

${\omega}_{4}^{*}$

${\omega}_{5}^{*}$

${\omega}_{6}^{*}$

0

3.5160

3.5160

22.0345

22.0345

61.6973

61.6973

Ref. [17]

3.5160

3.5160

22.0340

22.0340

61.6970

61.6970

2

1.5160

5.5160

20.0345

24.0345

59.6973

69.6973

Ref. [17]

1.5160

5.5160

20.0340

24.0340

59.6970

63.6970

3.5

0.0160

7.0160

18.5345

25.5345

58.1973

65.1973

Ref. [17]

0.0000

7.0160

18.5340

25.5340

58.1970

65.1970

4

–

7.5160

18.0345

26.0345

57.6973

65.6973

Ref. [17]

–

7.5160

18.0340

26.0340

57.6970

65.6970

8

–

11.5160

14.0345

3.30345

53.6973

69.6973

Ref. [17]

–

11.5160

14.0340

30.0340

53.6970

69.6970

The nondimensional natural frequencies of a rotating cantilever composite shaft obtained using the present model together with those obtained in Ref. [17] are shown in Table 4 for different nondimensional rotating speeds. A perfect agreement of numerical results with those in Ref. [17] can be seen.
5.2. Results and discussion
In the following numerical example, the composite rotor system is a composite shaft with one rigid disk supported by two rigid bearings at the ends, the disk is at the midpoint of the shaft. The material properties of the composite shaft are those in Table 1. The stacking sequence is ${[\pm \theta ]}_{5}$. The length, the radius and the wall thickness of the shaft are assumed to be $L=$ 2.47 m, $r=$ 0.0135 m and $h=$ 0.001321 m, respectively whereas those of a uniform rigid disk are ${b}_{1l}^{D}=$2.4364 kg, $e=$5×10^{5} m, ${b}_{4l}^{D}={b}_{5l}^{D}=$0.1901 kg/m^{2}. The rotor system is shown in Fig. 2.
Fig. 2. A composite shaftrigid disk with rigid bearings
Fig. 3 shows the variation of the first two flexural natural frequencies vs. rotating speed for various ply angles, where, 1F denotes the first forward mode whereas 2F denotes the second forward mode. The synchronous whirl line $\omega =\mathrm{\Omega}$ is also presented in the same figure which is the wellknown Campbell diagram. The intersection point of the line $\omega =\mathrm{\Omega}$ and the flexural natural frequency curves is referred to as the critical speed. The motion corresponding to positive flexural natural frequencies are in forward mode. The backward mode branch of the Campbell diagram with negative value of frequencies is on the lower half $\omega \mathrm{\Omega}$plane and is symmetrical with respect to the $\mathrm{\Omega}$ axes. The curves corresponding to negative frequencies are not shown in this figure for the sake of simplicity.
Fig. 3. The Campbell diagram of a composite rotor for various ply angles
From Fig. 3, it is seen that in the inertial frame, natural frequencies of the rotating composite rotor become function of rotating speed as flapping bending and sweeping bending motions are coupled due to the presence of internal damping (there is no the Corilois effect in the inertial frame). In fact, natural frequencies of a rotating shaft are constant at all rotating speeds in the inertial frame if the influence of internal damping is not considered. The reason is that for the rotating shaft of circular section, the displacement and acceleration vectors and the elastic restoring force always have same direction [18]. As seen in Fig. 3, the effect of the ply angle and rotating speed appear to be more significant for the highermode ones.
Fig. 4 shows the variation of the first two flexural dampings vs. rotating speed for various ply angles. It can be seen clearly that as the rotating speed is increased, the dampings of the backward modes decrease and remain negative for all rotating speed, so the backward modes are stable. From the results of Fig. 4, it can be also observed that the dampings corresponding to the forward modes are negative at low rotating speed and increase with increasing rotating speed, and at certain value of rotating speed the dampings vanish and then become positive. Transformation of dampings from negative to positive values marks the onset of unstable motion. The rotating speed corresponding to zero damping is the threshold of instability of the composite rotor. From Fig. 4 it seems that the effect of the ply angle on the damping of the mode 2 is more significant.
Fig. 4. The decay plot of a composite rotor for various ply angles
a) Mode 1
b) Mode 2
Fig. 5. Root locus of a composite rotor with internal damping
Fig. 5 shows the root locus plot of the complex eigenvalue as a function of rotating speed for various ply angles. For brevity, only the first mode is represented. In the figure the rotating speed is varied from $\mathrm{\Omega}=$0.0 to $\mathrm{\Omega}=$200 rpm in increments of 8.0 rpm. It is readily seen that the curve crosses the imaginary axis into the right hand side of the locus (the unstable region) with the increase of the rotating speed. It is also important to note that in the stable region, the decrease of ply angle will exert a stabilizing effect on the composite rotor, as seen by the curves with lower ply angles being further into the negative region of the locus. In contrast, in the unstable region, the decrease of ply angle promotes further instability, shown by the curves with lower ply angles being further into the right hand side of the locus.
Fig. 6 shows the effect of ply angle on the first critical rotating speed. It can be seen that as the ply angle increases, the critical rotating speed decreases and the critical speed is maximum at $\theta =$0°.
Fig. 7 shows the effect of ply angle on the threshold of instability for the flexural mode. It is evident that the general effect of the ply angle on the threshold of instability is similar to that associated with the critical rotating speeds. By comparing Fig. 6 with Fig. 7, it may be noted that for the same ply angle the threshold of instability is larger than the critical rotating speed. This implies that the onset of instability always occurs after the critical rotating speed.
Fig. 6. The variation of the first critical speed with ply angle
Fig. 7. The variation of thresholds of instability with ply angle
Fig. 8. The time history of deflections and the trajectories of the geometric center of the rigid disk ($\mathrm{\Omega}=$20 rpm)
a)
b)
c)
The transient responses for the geometric center of the rigid disk due to the initial perturbations are calculated by integrating Eq. (20) (with $\left\{f\right\}=\text{0}$) using the fourth order RungeKutta method. Figs. 810 present the time history of deflections and the trajectories of the geometric center of the rigid disk for rotating speeds 20 rpm, 27.89 rpm and 32.5 rpm, respectively. In the simulation, the ply angle $\theta =$45°, and the initial conditions ${X}_{jv}\left(0\right)=$ 10^{5}, ${X}_{jw}\left(0\right)=$10^{5}, ${\dot{X}}_{jv}\left(0\right)=\text{0}$ and ${\dot{X}}_{jw}\left(0\right)=\text{0}$ are specified. Fig. 8 shows that at rotating speed below threshold of instability the oscillation in the two direction can be damped out by internal damping. Fig. 9 shows that when rotating speed is equal to the threshold of instability a whirling response with constant amplitude occurs as internal damping equals to zero in this case. It can also be noted from Fig. 10 that the forward mode of a supercritical composite rotor is unstable since the internal damping becomes positive in this case.
Fig. 9. The time history of deflections and the trajectories of the geometric center of the rigid disk ($\mathrm{\Omega}=\mathrm{}$27.89 rpm)
a)
b)
c)
Finally, the steady state response of the composite rotor under the action of unbalance rigid disk is studied. The effect of internal damping on the steady state response is illustrated. In the simulation, the mode superposition method is used to obtain frequency response plots. Fig. 11 shows the two frequency response plots, the plot in dashed line corresponding to the case in which the internal damping are considered and the solid line corresponding to the case in which they are not. As can be seen from Fig. 11, internal damping tends to reduce resonance peak amplitude at critical speeds. This shows that the effect of internal damping on the the forced response is the same as external damping. Fig. 12 shows the effect of ply angle on unbalance steady state response with internal damping. It can be observed that the resonance peak amplitude of the lower mode is much stronger affected by internal damping, on the other hand, the effect of internal damping on the resonance frequency of the higher mode is more significant.
Fig. 10. The time history of deflections and the trajectories of the geometric center of the rigid disk ($\mathrm{\Omega}=\mathrm{}$32.5 rpm)
a)
b)
c)
Fig. 11. The frequency response of composite rotor with and without internal damping
6. Conclusions
A model was presented for the study of the dynamical behavior of the composite rotor with internal damping. The presented model was used to predict the natural frequencies, critical rotating speeds, instability thresholds, the transient free vibration responses and the steady state unbalance responses. Theoretical solutions of the free vibration of the system were determined by applying Galerkin’s method. From the present analysis and the numerical results, the following main conclusions were drawn:
1) The developed model provides means of predicting of the dynamic behavior of rotating composite rotors with internal damping.
Fig. 12. The frequency response of composite rotor with various internal damping
2) The ply angle affect the steady state unbalance responses and instability behavior of composite rotors significantly.
3) There is an obvious decrease in the first critical rotating speed and instability threshold as the ply angle is increased. And the critical rotating speed and threshold of instability have their maximum values at $\theta =$0°.
4) The onset of instability always occurs after the critical rotating speed.
Acknowledgements
The research is funded by the National Natural Science Foundation of China (Grant Nos. 11272190), Shandong Provincial Natural Science Foundation of China (Grant Nos. ZR2011EEM031) and Graduate Innovation Project of Shandong University of Science and Technology of China (Grant Nos. YC130210).
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