Research Article  Open Access
Liu Jun, Huang ZhiQiang, Li Gang, Li ShiDe, Liu Xian, Zhang Kai, "Dynamic Characteristics Analysis of a Seismic VibratorGround Coupling System", Shock and Vibration, vol. 2017, Article ID 2670218, 12 pages, 2017. https://doi.org/10.1155/2017/2670218
Dynamic Characteristics Analysis of a Seismic VibratorGround Coupling System
Abstract
A mathematical model to describe the seismic vibrator and ground coupling is proposed based on dynamic analyses of the vibrator baseplate structure and properties of soil. This mathematical model is solved using Newmark’s method. It produces reasonable results and the trend of amplitude envelope follows the experimental data. With this model, characterization of the vibratorground coupling is performed. Modal analysis study shows that the motions between the vibrator reaction mass and baseplate experience a combination of a firstorder vibration mode and a secondorder vibration mode. The corresponding frequencies of two vibration modes are located at 2.77 Hz and 488.7 Hz, respectively. As the frequency goes up, the motion behavior of the reaction mass and baseplate is dominated by the secondorder vibration mode. It is realized that the secondorder vibration mode is responsible for the phase difference between the input sweep signal and the output force signal. Furthermore, the impact of the coupling system parameters on the vibrator output force is also investigated. It is observed that the vibrator output force decreases as the sweep frequency increases. The weight ratio of the reaction mass and the baseplate has an impact on the vibrator output force.
1. Introduction
Due to efficient operation and being environment friendly, seismic vibrators (Figure 1) have become a main energy source on land in oil and gas exploration. With extension of seismic acquisition frequency bandwidth towards low and high ends, seismic vibrators face more challenges to produce a predictable and invariant output force signal. To achieve this goal, it is very necessary to understand the impact of key parameters in the vibratorground system on the vibrator output force [1].
Chapman et al. [2] presented a new design of the vibrator actuator to improve the vibrator output force in the frequency range of 200 Hz. They stressed that the weight and stiffness of the vibrator baseplate played a big role in extension of the vibrator output force bandwidth towards 200 Hz. Sallas and Weber [3] used a massspringdamper model to describe the vibrator system and the coupled ground system. Under rigidbody assumptions of reaction mass and baseplate, they theoretically demonstrated that a vibrator output force could be estimated using a weightedsum method that is summing reaction mass and baseplate acceleration weighted by their respective masses. This weightedsum ground force estimate provided a mean to approach the vibratorground force. However, this weightedsum ground force estimate was proved to be valid in a narrow frequency bandwidth by many researchers [4–8].
Wei [9, 10] demonstrated that the frequency bandwidth of the vibrator output force was constrained by hydraulic and mechanical limits of the vibrator system. Based on experimental testing results, he derived formulas that can be used to estimate the maximum fundamental force from the vibrator at low frequencies. Subsequently, Wei [11] simulated the dynamic behavior of the vibrator baseplate as well as its coupled ground using finite element analysis tools. Then, he [11] concluded that the vibrator performance was significantly affected by the baseplate stiffness.
To improve the vibrator force output towards a wide frequency range, one way is to design broadband sweeps using nonlinear sweep design techniques [12, 13]. The other way is to improve the performance of fundamental components in the vibrator mechanical and hydraulic system [14, 15]. Although improvements of individual components in the vibrator system are seen [5–16], the theoretical progress on the seismic vibrator design moves very slowly. This is because the dynamics between the vibrator and ground coupling are unclear.
The purpose of this paper is to present a mathematical model that is used to simulate the vibratorground coupling. Next, an impact of vibrator system parameters on the vibratorground coupling is investigated. Last, an impact of these parameters on the quality of the vibrator force output is performed.
2. Ground Stiffness and Damping
2.1. Surface Displacement Field of a HalfInfinite Ground Space under Concentrated Vertical Harmonic Disturbances
The displacement field of a halfinfinite ground space under a concentrated vertical harmonic disturbance force (Figure 2) can be written as follows [17, 18]:with boundary conditionswhere , , volume strain , and rotational component refer to Table 3 and, here, , , and are the density, elastic modulus, and Poisson’s ratio of the ground, respectively. and are displacement components and and are stress components.
The surface displacement field of the halfinfinite space can be derived as follows:where and are, respectively, zerothorder and firstorder Bessel functions of the first kind and , , , , , , and refer to Table 3.
The second formula in (3) yields the surface displacement component of the halfinfinite space under a vertical disturbance as where is the dimensionless frequency and and are the functions of . The specific expression refers to Table 3.
2.2. Wave Displacement Field of a Circular Baseplate under a Simple Harmonic Load
Using the wave model for a dynamic half space presented in Section 2.1, the displacement field of an infinite half space with a circular surface under a simple harmonic load can be derived as shown in Figure 3.
A vibrational coupling model of a circular baseplate and the ground is shown in Figure 3, where , , , and denote, respectively, the displacement components along the horizontal axis , the vertical axis , and the direction . The displacements produced by a body wave are given as follows [19, 20]:where and refer to Table 3.
The stress field of the circular baseplateground system under the action of a harmonic disturbance can be obtained from (5). To determine the vertical displacement component of the circular baseplate with , the distribution of the reaction force imposed on the bottom of the baseplate needs to be determined.
The force beneath the baseplate is determined by the stress distribution pattern beneath the baseplate. Three stress distribution patterns, which are (a) a static stiffness, (b) a uniform distribution, and (c) a parabolic distribution, are assumed (Figure 4). The stress boundary based on this three stress distribution patterns is obtained. The specific expression of refers to Table 3.
(a)
(b)
(c)
The displacement at the center of the baseplate will be equal to the displacement of the ground, :where and are the dimensionless frequencies. For each assumption of the stress distribution and refer to Table 3.
Redefining and , (6) can be written asThat is,where is the driving signal, is the response, and is the impedance function of the vibrational system.
Equation (7) is the dynamic model of the infinite half space where circular surface is under a uniform stress .
2.3. Dynamic Model of BaseplateGround Vibrational System
For the baseplateground system, if nonlinear factors are neglected, the system vibration can be treated as a superposition vibration under a vertical disturbance force (Figure 5).
According to (8), the reaction force on the bottom of the baseplate can be written as
Therefore, the differential equation describing the dynamics of the baseplate can be written as [19]That is,where and are, respectively, the mass and the displacement response of the baseplate, is the amplitude of the external force, and is the phase difference.
For harmonic vibrations, and and (11) can be written asUnder the assumptions (12) can be simplified towhere and are the stiffness and damping . They are functions of the disturbance frequency . Expressions for and are given in Table 1 for the different modes of vibration.

3. Dynamic Model of Seismic VibratorGround Coupling
3.1. Vibrational Coupling Model
A simple seismic vibratorground coupling system is shown in Figure 6 [3].
The dynamic equations for the coupling system can be written aswhere and are, respectively, the masses of the reaction mass and the baseplate, and are, respectively, the stiffness and damping of the ground, and and are, respectively, the stiffness and damping of the hydraulic oil. is the sweep signal and can be written as follows for a linear signal: where is the amplitude, is the sweep length, is the time, and are the start frequency and end frequency, and is the sweep bandwidth.
3.2. Solution of the Model
Equation (15) is solved using Newmark’s integral method, with the displacement and velocity at instant being expressed, respectively, aswhere is the time interval. and are parameters related to the precision and stability, and the values are 0.5 and 0.25, respectively.
Using (17a) and (17b), the velocity and acceleration at instant can be expressed approximately by the responses at instant and the displacement at instant :
Substitution of (18a) and (18b) into (15) yields the following equation:
The system responses, including displacements, velocities, and acceleration, at instant can be obtained by substituting the solution of (19) into (18a) and (18b).
3.3. Verification of the Model
To verify the validity of the dynamic model, the force output of a real seismic vibratorground system is calculated and compared with the experiment data. The experiment was taken with an AHVIV model362 vibrator positioned on load cells using 5 Hz to 105 Hz in a 20second linear sweep. A linear sweep was used to run a vibrator equipped with an advanced actuator which carries a modified hydraulic system and a new stiffer baseplate that is 2.5 times stiffer than the standard vibrator baseplate [21, 22]. This vibrator in Figure 7 was placed on load cells from which the true vibratorground force can be measured. All load cells were firmly mounted on a concrete pad and the target force was set at 42000 lbs or 186816 N [14]. In the experiment, it is assumed that the ground is homogeneous. Key parameters in the model and the experiment are listed in Table 2, and the parameters of the sweep signal in (16) are kN, Hz, Hz, and s.


The timehistory response of the output force from the model is compared with experimental results in Figure 8. Both model data and experimental results show that the vibrator output force decreases as the sweep frequency goes up. At very high frequencies, the torque motor current on the pilot servovalve reaches its limit of 40 mA (milliampere). Often times, the instability of the servovalve system occurs once the torque motor current is saturated and causes the poor vibrator performance [23]. The amplitude envelope in Figure 9 produced by model data shows a similarity to the amplitude envelope of experimental data although some discrepancy is still seen. This shows that the model is reasonable and can be used to estimate the impact of parameters in vibratorground system on the vibrator output force.
(a)
(b)
4. Dynamic Characteristics Analysis
4.1. Modal Analysis
The characteristic equation for the dynamic equations (15) of the coupling system can be written as
The natural frequency of the system can be obtained by solving this characteristic equation and is found to be
The vibrational modal matrix of the coupling system can be written as
If a firstorder vibration system is used to describe the vibrator actuator motion at low frequencies, the reaction mass and baseplate motions can be depicted as a synchronous vibration where the natural frequency is at 2.77 Hz (Figure 10(a)). The displacement amplitude of the reaction mass is much greater than the baseplate. This means that contribution to the vibrator output force from the baseplate motion can be neglected and decoupling hardly occurs.
(a)
(b)
As the frequency goes up, a secondorder vibration system is capable of describing the reaction mass and baseplate motions where the natural frequency is at 488.7 Hz (Figure 10(b)). The displacement amplitude of the baseplate becomes greater than the reaction mass. As the vibration frequency approaches the natural frequency, the phase difference of the reaction mass and baseplate becomes large and goes to 180 degrees apart. This outphase relationship will possibly cause the decoupling. It is important that the sweep frequency should not be close to this natural frequency.
4.2. Effects of System Parameters on Vibrational Output
The effects of reaction (inertial) mass and baseplate mass on the vibrational output are shown in Figures 11 and 12 which illustrate the effects of the masses of the reaction mass and baseplate on the vibrator output force, respectively. It can be seen from Figures 11 and 12 that an increase of the reaction mass results in an increase of the vibrator force amplitude while an increase of the mass of the baseplate results in a decrease of the vibrator force amplitude. These findings provide guidelines for the design of the vibrator actuator. It is authors’ opinion that the mass ratio of the reaction mass and the baseplate should be kept at 3 : 1.
Figure 13 displays how the elastic modulus of the ground affects the vibrator output force. It can be observed that increasing the elastic modulus of the ground will improve the vibrator output force. Increasing the elastic modulus means that the ground becomes stiff. Although it helps improve the amplitude of the vibrator output force, it reduces the forceenergy that is transmitted to the deep ground. For example, a vibrator vibrating on a piece of limestone will produce an improved force amplitude. However, the forceenergy radiated into deep ground will become poor.
4.3. Effects of System Parameters on Phase Difference
To obtain a vibrational output force of high quality, it is necessary that the phase difference between the sweep signal and the vibrational output be as small as possible. To investigate the influence of the system parameters on the phase difference, the following harmonic signal is used to replace the actual linear sweep signal shown in (16):
Figure 14 shows the vibrational output response under a harmonic sweep signal. It can be seen that, with increasing frequency of the sweep signal, the phase difference becomes larger.
(a)
(b)
To further investigate the influence of the frequency of the sweep signal on the phase difference, the load vector of the system dynamic (15) is replaced by the following expression:
The steadystate solution of (15) can be assumed to bewhere and are the displacement state solutions for the vibrating plate and hammer and , , , and are undetermined coefficients.
Substituting , , and and , , and into (15) and separating the sine and cosine functions yield the following equations:
The undetermined coefficients , , , and , and thus and , can be determined by solving (26a), (26b), (26c), and (26d).
The force transmitted to the ground through the baseplate can be written as
Substituting into this equation then gives where and are, respectively, the amplitude of the vibrational output and the phase difference between the sweep signal and the vibrational output.
Figure 15 shows that the effect of the reaction mass on the phase difference is significant, with greater reaction mass leading to a bigger phase difference, but, as shown in Figure 16, there is no obvious effect of baseplate mass on the phase difference.
Figure 17 shows that there is an obvious influence of baseplate area on phase difference, with greater baseplate area leading to a smaller phase difference. It can be seen from Figure 18 that the greater the ground stiffness, the greater the phase difference.
5. Conclusions
The dynamic characteristics of a seismic vibratorground coupling system have been investigated by analyzing the system’s vibrational model and the influence of the main system parameters on the vibrational output and the phase difference between the input sweep signal and the vibrational output. The following conclusions can be drawn from the results obtained:(1)The main vibrational modes of the coupling system are the first and secondorder modes with frequencies 2.77 Hz and 488.7 Hz, the latter being reverse movement of the reaction mass and baseplate will lead to an output phase difference of 180 degrees. This outphase relationship will possibly cause the decoupling. Therefore, in an actual vibrator design, the frequency of the sweep signal should be lower than the secondorder frequency of the system.(2)The design of the two most important components of the coupling system, namely, the baseplate and the reaction mass, should aim at improving the amplitude of the vibrational output and reducing the phase difference. It is found in this paper that an increase of the reaction mass results in an increase of the vibrator force amplitude while an increase of the mass of the baseplate results in a decrease of the vibrator force amplitude. These findings provide guidelines for the design of the vibrator actuator. The mass ratio of the reaction mass and the baseplate should be kept at 3 : 1.(3)The physical properties of the ground will affect the natural frequency of the coupling system and ultimately the quality of the output signal. Therefore, the sweep signal should be determined according to the physical properties of the ground and the modal analysis of the coupled system. That is, the frequency of the sweeping signal should be kept away from the frequency of secondorder vibration of the vibratorground system.
Conflicts of Interest
The authors declare that there are no conflicts of interest regarding the publication of this paper.
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Copyright © 2017 Liu Jun et al. 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.