Abstract
With fuel consumption, the fuel container system vibrates with decreasing mass, which is a typical variable mass system. This paper investigates the dynamic characteristics of flexible rectangular fluid containers with decreasing liquid. The dynamic equations of the container with time varying liquid are derived by combining finite element method (FEM) and virtual mass method (VMM). Free vibration states of the variable mass system are mainly investigated. The vibration signals are decomposed using ChoiWalliam Distribution, and the energy density spectrum is given by time frequency analysis. Results show that decrease of the liquid of the system induces increase of the vibration frequencies of the system, and generates an additional negative damping causing the vibration decay slowly. It is found that the additional damping is proportional to rate of mass change. The additional negative damping can cause the system vibrate with increasing amplitudes while the negative damping plays the dominant role rather than the structural damping.
1. Introduction
The liquid storage containers are widely used to store vital products, such as petroleum products for transport facilities and aerospace industries (wing containers of aircraft, liquid propelled launchers, etc.). The dynamic behaviors of the fluidstructure coupled systems are important for design and safe use of the liquid storage container. In the previous study of vibration of containers containing liquid [1], the container wall and foundation were assumed to be rigid. The impulsive pressure resulting from the liquid mass accelerating due to the ground acceleration was taken into consideration in the study. Moiseev [2] constructed normal mode functions by integral equations in terms of Green’s function of the second kind. Chu [3] generalized Moiseev’s method by employing a perturbation technique using the characteristic functions to determine the subharmonic response of an arbitrary tank to an oscillatory axial excitation. In aerospace industries, NASA investigated the dynamic behavior of liquids in moving containers in 1966 [4].
For simplicity, the effect of liquid on the structure dynamics was considered as a set of simple resonators (massspring or pendulum) [57]. It is simple to estimate the low natural frequencies of the coupling system by this method. Hence in the aerospace domain, this approach is still widely used. The influence of motion of propellant on the control properties of launch vehicles were investigated utilizing this simplified model [811]. For more accurate results, many approximation methods were developed. J. D. Wang et al. [1213] studied sloshing response of liquid in a rigid cylindrical container with multiple annual rigid baffles subjected to lateral excitations by a semianalytical approach based on the mode superposition method. In E. Askari’ study [14], the interaction between the fluid and the structure was taken into account applying the Rayleigh quotient and Galerkin method. S. Rebouillat and D. Liksonov [15] introduced recent studies devoted to the problem of modeling the solidfluid interaction in partially filled liquid containers. M. Eswarana et al. [16] investigated the difference between baffled and unbaffled containers, and numerical simulation and experimental validation were both given. J. R. Shao et al. [17] developed an improved analytical method to model the liquid structure coupling system. R. Belakroum et al. [18] presented a numerical partitioned strongly coupled approach to the vibration in containers partially filled with liquid, the obtained numerical results agreed with analytical results. O. Andrianarison and R. Ohayon [19] derived the basic equations describing the interactions between acoustic and gravity waves and constructed appropriate variational formulations. For fluid rectangular containers, the RayleighRitz method was useful for modelling the fluidstructural coupling system [20, 21].
Many numerical methods were used in analyzing the dynamic behavior of in liquidfilled containers. M. Sauer [22] utilized the adaptive SPH approach to solve impact problems on fluidfilled containers with complete rupture of the container, obtaining a good result. Xiang Xu et al. [23] developed a novel matrix method for a liquidfilled circular cylindrical shell with partially constrained layer damping. The finite element method was utilized to study the dynamic interaction between the liquid and elastic containerbaffle system [2426]. P. Pal et al. [27] dealt with the numerical and experimental studies of motion of liquid in partially filled prismatic containers subjected to external excitation by meshless local PetrovGalerkin (MLPG) method.
Actually, the flexible fluid containers with time varying fluid are complicated systems. On the one hand, they are classical fluidstructure coupling system; on the other hand they are time varying systems. During transportation of fluid containers, impacts take place inevitably because of various causes. Recently, baffles are widely used to suppress the fluid sloshing in fluid containers transported by trucks. However, the time varying liquid affects the dynamic characteristics of the systems significantly. Especially for rockets, the ratio between the propellant mass and the initial mass of the vehicle can reach 95 %. Liquid propelled launchers or containers with fluid transported are greatly influenced by the inertial forces of internal fluid. These forces are proportional to the acceleration of wetted surface, and can consequently be represented by fictive fluid mass that vibrates along with the walls of the containers. This fictive mass is called “hydrodynamic added mass” [2830], which in addition to the structure mass represents “virtual mass” [3134]. In dynamic analysis of liquid propelled launchers or containers with fluid transported, virtual mass method has been widely used to model the liquidstructure coupling system for its simplicity and practicability [3234]. Hence for liquid propelled launchers or containers with fluid transported by a truck, the sloshing of the fluid and the free surface influence the dynamic characteristics of the system little, while the most important role of the liquid is the additional mass of the system.
In previous studies, the dynamic behaviors of containers with constant liquid were usually analyzed, however the liquid in the containers are usually consumables. The main purpose of this study is to organize an efficient numerical procedure to analyze the effect of time varying liquid on the response of fluid storage containers, considering the coupling of the elastic container and time varying liquid. To take into account the fluidstructure interaction phenomenon, the liquid region is modeled by virtual mass method. The coupled equations for the fluidstructure modes are derived based on FEM and VMM. The main effect of time varying mass is investigated based on time frequency analysis. Finally, some conclusions are briefly summarized.
2. Basic equations
Consider a generic internal fluidstructure system shown in Fig. 1. The flexible container with the fluid in it is viewed as a coupled hydroelastic system. A rectangular tank with four flexible vertical walls of uniform thickness ${t}_{w}$ and a horizontal rigid bottom is partially filled with incompressible and nonviscous liquid of depth $h\left(t\right)$, as shown in Fig. 1. $h\left(t\right)$ is a function of time $t$. The side lengths and height of the tank are $a$, $b$ and $c$, respectively. The walls of the tank are considered as thin plates made of linearly elastic, homogeneous and isotropic material and are assumed to perform transverse bending deflection but no inplane deformation. The motion of the liquid is assumed to be frictionless and irrotational so that the velocity distribution of the liquid may be represented as a gradient of the velocity potential.
Fig. 1A rectangular container partially filled with liquid
If $P$ is the pressure within the fluid and ${\mathbf{u}}_{F}$ is the vector of fluid displacements, then its dynamic behavior is governed by the Euler’s equation:
The boundary conditions must be satisfied:
a) At liquidstructure interface:
b) At liquid free surface:
However, if free surface wave of the liquid is ignored, we have $P=\text{0}$.
c) At the bottom of the container:
where ${\mathbf{u}}_{\mathrm{s}}$ – the vector of wetted structural displacements, $\rho $ – density of fluid, $g$ – the value of the acceleration acting normal to the free surface, $\mathbf{n}$ – outward unit normal to the fluid surface.
The fluid displacement ${\mathbf{u}}_{F}$ can be expressed in terms of a potential function $\varphi $ as follows:
Because the fluid is incompressible and irrotational, the following functions are derived:
Substituting Eq. (5) into Eq. (1) and integrating gives the following differential equation for the fluid:
The following is a brief overview of the virtual mass approach in MSC. Nastran [34]. The Helmholtz method used by MSC. Nastran solves Laplaces equation by distributing a set of sources over the outer boundary, each producing a simple solution to the differential equation. By matching the assumed known boundary motions to the effective motion caused by the sources, we can solve a linear matrix equation for the magnitude of the sources. The values of the sources determine the effective pressures and, thereby, the forces on the grid points. Combining all of these steps into a matrix equation results in a virtual mass matrix as derived below. If ${\sigma}_{j}$ is the value of a point source of fluid (units are volume flow rate per area) located at location ${\mathbf{r}}_{j}$, and is assumed acting over an area ${A}_{j}$, the potential $\varphi \left({\mathbf{r}}_{i}\right)$ at any other point ${\mathbf{r}}_{i}$ is:
where ${\mathbf{e}}_{ij}$ is the unit vector in the direction from point $j$ to point $i$. Note that the gradient of the vector ${\dot{\mathit{u}}}_{i}$ is the potential function which satisfies Laplace’s equation on a term by term basis. The other set of necessary equations are the pressures ${P}_{i}$ at any point $i$ in terms of the density $\rho $, sources and geometry, namely:
The results of integrating Eq. (9) and Eq. (10) over the finite element surfaces are collected respectively in two matrices $\mathbf{\chi}$ and $\mathbf{\Lambda}$, where:
where $\mathbf{F}$ are the forces at the grid points. The matrix $\mathbf{\Lambda}$ is obtained by integrating Eq. (11).
An additional area integration is necessary to convert the pressures to forces. A mass matrix may now be defined using Eq. (11) and Eq. (12) as:
where the virtual fluid mass matrix ${\mathbf{M}}_{s}^{f}$ is:
For a container filling with constant liquid, the potential energy and kinetic energy can be written as:
where ${\mathbf{M}}_{\mathrm{s}}$ and ${\mathbf{K}}_{\mathrm{s}}$ are the mass matrix and stiff matrix of the elastic structure. For a fluidstructure coupling system with variable mass in a container of spacecraft, the mass of the system decreases in flight. The effect of the fuel on the stiffness of the structure can be neglected. Hence the potential energy for the time varying mass system is defined as:
The dynamic equations of the variable mass system can be derived according to Lagrange equation:
Substituting Eqs. (16)(17) into Eq. (18), the dynamic matrix equations can be written as:
where ${\mathbf{F}}_{\mathrm{s}}$ is the external force vector, and ${\dot{\mathbf{M}}}_{s}^{f}$ is derivative of additional mass. From Eq. (19), an additional damping is induced by the change of the mass. If the mass increases, an additional positive damping is induced, while the mass decreases, an additional negative damping is induced. It should be noted that the additional negative damping may cause instability of the vibration. Let $\mathbf{C}$ is the Rayleigh damping matrix, which is the form of:
Rayleigh damping coefficients can be determined by the first two natural frequencies:
where ${\omega}_{1}$ and ${\omega}_{2}$ are frequencies, ${\zeta}_{1}$ and ${\zeta}_{2}$ are damping ratios.
The final dynamic matrix equations can be expressed as Case I:
For the sake of contrast, a special Case II is also considered:
For this case, the additional damping is not taken into account. Therefore, the effect of the additional damping on the dynamic characteristics of the container can be verified.
3. Numerical procedure
A calculation procedure based on MATLAB is used for modal analysis, dynamic analysis and signal processing, in conjunction with PATRANNASTRAN, which is widely used for linear structural and vibration analysis of aircrafts. Fig. 2 shows a procedure for dynamic analysis and signal processing of a container with time varying mass. The virtual fluid mass can be directly calculated based on the PATRANNASTRAN virtual mass calculation program. However, as PATRANNASTRAN is basically a linear analysis code for time invariant system, useless results are often obtained for time varying models. MATLAB is more suitable for nonlinear calculation, but complicated fluid modeling and largescale calculation are necessary to calculate the impulsive mass using fluidstructure interaction analysis for this container with time varying mass. Based on this consideration, only virtual mass are to be calculated by PATRANNASTRAN. The obtained virtual mass data are converted into a MATLAB code to carry out the dynamic analysis. The response acceleration and displacement are calculated using Newmark direct integration method. After that, time frequency analysis is developed to study the characteristic of the time varying mass system. The vibration signals are decomposed by ChoiWalliam distribution, therefore time frequency features of the time varying mass system are obtained.
Fig. 2The flow chart of the calculation
4. Comparison study
In this section, a comparison study is first given to determine the validity of the theoretical formulation in computing dynamic responses of the flexible storage tank. The dimensionless natural frequencies ($\mathrm{\Omega}$) of wall vibration are expressed as:
where $D$, ${t}_{w}$ and ${\rho}_{w}$ are the flexural rigidity, the wall thickness and density of the tank wall, respectively. $w$ is the circular natural frequency of the container and:
where $E$ and $\nu $ are the Young’s modulus and Poisson’s ratio of the tank wall. The first six dimensionless natural frequencies of wall vibration (impulsive modes) of the empty cubic tank ($a=b=c$) and a cubic tank fully filled with liquid which has a mass ratio of the liquid to the tankwall equal to 5 are presented in Table 1 and compared with those presented by Zhou and Liu [20] and S. Hashemi [21]. A mesh grid of 20×20 is used for each wall surface. Good agreements are observed for all cases, as shown in Table 1. The parameter effects on eigen frequencies of the liquidtank system are investigated for a cubic tank. The effect of the liquid depth on the first six bulging eigen frequencies is given in Table 2 when the tank is partially filled with the liquid. One can see that the bulging eigen frequencies decrease with the increase of the liquid depth.
Fig. 3Finite element model of a filled cubic tank
Table 1Comparison on the first six impulsive frequencies of the empty and filled cubic tank
Frequency  Empty cubic tank  Cubic tank filled with liquid  
Present study  S. Hashemi [21]  Zhou [20]  Present study  S. Hashemi [21]  Zhou [20]  
${\mathrm{\Omega}}_{1}$  17.509  17.619  17.632  8.318  8.212  8.48 
${\mathrm{\Omega}}_{2}$  35.910  36.049  36.045  18.615  18.011  18.175 
${\mathrm{\Omega}}_{3}$  51.775  52.076  52.133  26.112  26.742  26.212 
${\mathrm{\Omega}}_{4}$  70.988  71.389  71.208  39.451  38.484  37.494 
${\mathrm{\Omega}}_{5}$  74.040  74.377  74.576  42.721  42.441  42.931 
${\mathrm{\Omega}}_{6}$  106.203  106.297  107.531  59.025  60.470  58.016 
Table 2The effect of the ratio of liquiddepth to tankheight on the first two bulging eigenfrequencies
$h\left(t\right)/c$  ${\mathrm{\Omega}}_{1}$  ${\mathrm{\Omega}}_{2}$  
Present study  Zhou [21]  Present study  Zhou [21]  
0.5  15.845  15.956  25.920  24.974 
0.6  14.056  14.146  24.593  23.544 
0.7  12.267  12.358  24.339  22.998 
0.8  10.691  10.865  23.461  22.134 
0.9  9.467  9.636  21.627  20.489 
The convergence study on the fundamental impulsive frequencies with respect to the term number of the mesh grids is shown in Table 3. It is seen that the impulsive frequencies monotonically decrease with the increase of the number of the mesh grids. The results show rapid convergence to the present method. The present method is now applied to investigate dynamic response characteristics of 3D rectangular fluid storage tanks.
Table 3The convergence study on the fundamental frequencies with respect to number of mesh grids
Fundamental dimensionless natural frequency  Mesh grids of the tank wall  
5×5  10×10  15×15  20×20  
Empty cubic tank  18.735  17.742  17.569  17.50 
Cubic tank filled with liquid  9.196  8.495  8.366  8.318 
5. Numerical example
Since rectangular tanks are used most often for the wettype storage of spent fuel assemblies, a typical dimension for those tanks is selected for the next example [24]. The material properties are: the density ${\rho}_{w}=\text{1200}\text{}\text{kg/}{\text{m}}^{\text{3}}$; the Young’s modulus $E=\text{2.9\xd7}{\text{10}}^{\text{9}}\text{}\text{N/}{\text{m}}^{\text{2}}$; and the Poisson’s ratio $\nu =\text{0.35}$. The model is designed according to the following geometric specifications: the height $c=\text{0.9 m;}$ the wall thickness ${t}_{w}=\text{0.035 m;}$ the length $a=\text{2.2 m,}$ and the width, $b=\text{1.15 m}$. The tank is assumed to be fixed to the ground ideally, and to have 1 % structural damping. The effect of the liquid in the tank on the stiffness of the structure is small, however the liquid influences the total mass of the structure much. Hence, the vibration frequency will get bigger as the liuqid decreases. The finite element model and the bulging mode shapes of the first three order of the rectangular container shown in Fig. 4 and Fig. 5. From the mode shapes, deformation at Node 389 in $y$ direction is large, hence the responses of Node 389 in $y$ direction is monitored. Table 4 gives the natural frequencies of the containers with different ratio of liquiddepth to tankheight. The natural frequency of the system is related directly to the liquiddepth. As introduced above, the value of the frequency get lowered remarkably resulting from a large ratio of liquiddepth to tankheight. From Table 4, the frequency of the first order bulging vibration increases from 6.696 Hz to 16.933 Hz, as the ratio of liquiddepth to tankheight decreases from 1 to 0.1.
Fig. 4Finite element model and mode 1st of the tank
a) Finite element model
b) Mode 1
Fig. 5Mode shapes of the tank
a) Mode 2
b) Mode 3
Table 4The first three coupling eigenfrequencies for different ratio of liquiddepth to tankheight
$h\left(t\right)/c$  ${f}_{1}\text{}\text{/}\text{}\text{Hz}$  ${f}_{2}\text{}\text{/}\text{}\text{Hz}$  ${f}_{3}\text{}\text{/}\text{}\text{Hz}$ 
0.1  16.933  17.750  29.442 
0.2  16.922  17.139  29.416 
0.3  16.845  17.658  29.247 
0.4  16.545  17.345  28.631 
0.5  15.754  16.524  27.125 
0.6  14.282  15.018  24.604 
0.7  12.319  13.042  21.588 
0.8  10.288  11.011  18.657 
0.9  8.470  9.186  16.039 
1  6.696  7.642  13.738 
The vibration responses of the container in $y$ direction with decreasing liquid are calculated using Newmark direct integration method. In Fig. 6(a) and Fig. 6(b), the displacement responses and frequency responses of the container with decreasing liquid are illustrated, and the responses for Case I and Case II are compared. In Fig. 6(a), from the two curves, it is found for Case I, a negative damping is induced resulting from the decreasing mass. The negative damping influences the vibration significantly, and the amplitude of the vibration increases. For Case II, the structural damping causes the vibration decay. The displacement signals are processed by Fourier transformation, as shown in the right one. From the frequency response curves, it is observed that the vibration frequency of the system consists of a series of frequencies. The frequency ranges from about 7 Hz to 17 Hz, coincided with the frequency listed in Table 4. The range of the frequency can be determined by the range of the mass of the system. From the results shown in Fig. 6 and Fig. 7, the influence of the negative damping is remarkable so that the additional damping cannot be neglected in the analysis of containers with time varying mass.
Fig. 6Response curves of Node 389 in y direction
a) Displacement response curves
b) Frequency response curves
6. The effect of the rate of mass change
As depicted in Eq. (19), the induced damping is proportional to the rate of mass change. In other words, if the mass of the system decreases fast, a large negative damping is generated. In this section, the liquid decreases in 1 s, meaning that the rate of mass change increases by 4 times, hence a large negative damping is induced. From the displacement curves in Fig. 7, one can find that for Case I the negative damping induces increase of the vibration amplitude, and for Case II, the structural damping influences the vibration much, which causes decay of the vibration. From the frequency response curves, the vibration frequency increases from about 7 Hz to 17 Hz. Compared with curves shown in Fig. 6, the amplitude of the vibration increases more resulting from the additional negative damping. In Fig. 6(b), the maximum amplitude is 1.2 mm, while in Fig. 7(b) the maximum amplitude is 1.8 mm. In other words, the influence of the negative damping must be taken into account in the analysis of container with decreasing mass.
Fig. 7Response curves of Node 389 in y direction
a) Displacement response curves
b) Frequency response curves
7. Time frequency analysis
With change of mass of the system, the frequency of the system is changing as well. The classical Fourier Transformation cannot give enough information of the vibration signals in timefrequency domain [35]. In this section, the time frequency analysis is used to analyze the timedomain signals. It has been shown that the class of timefrequency energy distributions have the following general expression [36]:
where $x\left(s\right)$ is the time signal, ${x}^{*}\left(s\right)$ is its complex conjugate and $\phi (\xi ,\tau )$ is a twodimensional function called the parameterisation or kernel function, representative of the particular distribution. This class of distributions, known as the Cohen’s class, uses a bilinear transformation that depends on two variables: time and frequency. Depending on the specific kernel function used for the bilinear transformation, various different distributions have been proposed.
Choi and Williams succeeded in devising a distribution for which the spurious values are minimal. It diminishes the effects of the crossterms within the range while retaining the desirable properties [35]. There are many papers demonstrating effectiveness of CWD [3638]. The ChoiWilliams distribution (CWD) is a member of Cohen’s class with kernel:
where $\sigma $ is a constant scaling factor to suppress the crossterms. By substituting this kernel into Eq. (26), the CWD is obtained:
Fig. 8 shows the ChoiWilliams distribution of vibration signals displayed in Fig. 6(a). As the mass of the system decreases, the vibration frequency increases. The change of the vibration frequency versus time is also given by ChoiWilliams distribution, which cannot be found employing classical Fast Fourier Transformation. In Fig. 8(a), the power spectral density of the vibration response for Case I is displayed. It is observed that the decrease of the mass induces a negative damping, causing increase of the energy of the system. Comparing the contour in Fig. 8(a) and Fig. 8(b), we can find that the energy density in Fig. 8(a) is larger than that in Fig. 8(b). The reason is that the negative damping induced by the decreasing mass has a great influence on vibration of the container. The results show that CWD is an efficient tool to decompose nonstationary signal.
The change of the vibration frequency versus time can also be obtained bases on ChoiWilliams distribution, as shown in Fig. 9. It can be found that the vibration frequency increases from 7 Hz to 17 Hz, coincided with that listed in Table 4. Using classical Fast Fourier Transformation, the range of the vibration frequency can be obtained, but the change of the vibration frequency versus time cannot be given.
Similarly, Fig. 10 and Fig. 11 show the ChoiWilliams distribution of vibration signals displayed in Fig. 7. In Fig. 10, one can also find that the additional negative damping induces increase of the energy density, hence the energy density for Case I is much larger than that for Case II as shown in Fig. 10(b). The change of the vibration frequency versus time can also be obtained bases on ChoiWilliams distribution, as shown in Fig. 11. It can also be found that the vibration frequency increases from 7 Hz to 17 Hz in 0 s1 s, coincided with that listed in Table 4.
Fig. 8ChoiWilliams distribution of vibration signals for Case I with different damping
a) Case I
b) Case II
Fig. 9ChoiWilliams distribution of vibration signals for Case i with different damping
a) Case I
b) Case II
Fig. 10ChoiWilliams distribution of vibration signals for Case I with different damping
a) Case I
b) Case II
Fig. 11ChoiWilliams distribution of vibration signals for Case i with different damping
a) Case I
b) Case II
8. Conclusions
In this paper, free vibration states of an elastic rectangular container with time varying mass were discussed. The decease of the internal liquid (like fuels) was taken into account in the numerical study of container structures. The fluidstructure coupling was analyzed as a time varying mass system. A numerical procedure was constructed to solve vibration problems of this time varying mass system, using MATLAB in conjunction with PATRANNASTRAN. Conclusively, in this study, the influence of the fluid sloshing is neglected, and the effect of the deceasing fluid on flexible containers is mainly investigated. From the engineering point of view, the following conclusions have been drawn:
1) The vibration frequencies increases with the decrease of the mass, and the range of the frequency can be determined by the range of the mass. From the dynamic equations, a negative damping is induced as the mass of the system decreases. The negative damping is an unstable parameter. In the vibration, the negative damping can slow down the decay amplitude. If the effect of the negative damping exceeds the effect of the structural damping, the amplitude of the vibration increases. In the simulation of vibration of containers with time varying mass, the negative damping cannot be neglected. The negative damping is proportional to rate of mass change. If the mass of the system changes fast, a big negative damping is induced, and the big negative damping can cause instability of vibration of the system.
2) For the time varying mass system, the time frequency feature is obtained using ChoiWilliams distribution. The energy density spectrum and the change of the frequency versus time are displayed. ChoiWilliams distribution is a useful tool to analyze nonstationary signals of time varying mass system.
3) In the study, the influence of slosh of the liquid the vibration of the system is little comparing with the influence of the time varying mass. The proposed method can be used to solve similar problems in engineering.
4) Further work intends to study dynamic behavior of containers with fuel consumption under different load cases. More work is still necessary in order to fully exploit the coupling between the fluid sloshing and time varying mass. As the decreasing mass induces an additional negative damping, stability analysis for time varying mass system is to be developed. In addition, an experiment will be provided for verification in detail.
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About this article
The authors of the paper gratefully acknowledge the financial support from National Natural Science Foundation (Grant No. 11302160 and Grant No. 11102147) of China.