 Original Article
 Open access
 Published:
Vibration Characteristics of Rotor System with Loose Disc Caused by the Insufficient Interference Force
Chinese Journal of Mechanical Engineering volumeÂ 35, ArticleÂ number:Â 70 (2022)
Abstract
The rotating parts looseness is one of the common failures in rotating machinery. The current researches of looseness fault mainly focus on nonrotating components. However, the looseness fault of discshaft system, which is the main work part in the rotor system, is almost ignored. Here, a dynamic model of the rotor system with loose disc caused by the insufficient interference force is proposed based on the contact model of discshaft system with the microscopic surface topography, the vibration characteristics of the system are analyzed and discussed by the number simulation, and verified by the experiment. The results show that the speed of the shaft, the contact stiffness, the clearance between the disc and shaft, the damping of the disc and the rotational damping have an influence on the rotation state of the disc. When the rotation speed of the disc and the shaft are same, the collision frequency is mainly composed of one frequency multiplication component and very weak high frequency multiplication components. When the rotation speed of the disc and the shaft is close, the vibration of the disc occurs a beat vibration phenomenon in the horizontal direction. Simultaneously, a periodical similar beat vibration phenomenon also occurs in the waveform of the discshaft displacement difference. The collision frequency is mainly composed of a low frequency and a weak high frequency component. When the rotation speed of the disc and the shaft has great difference, the collision frequency is mainly composed of one frequency multiplication, a few weak high frequency multiplication components and a few low frequency multiplication component. With the reduction of the relative speed of the disc, the trajectory of the disc changes from circleshape to inner eightshape, and then to circleshape. In the inner eightshape, the inner ring first gradually becomes smaller and then gradually becomes larger, and the outer ring is still getting smaller. The obtained research results in this paper has important theoretical value for the diagnosis of the rotor system with the loose disc.
1 Introduction
Looseness fault of rotating parts, which is one of the common failures in the rotor system, has a serious influence on the normal operation of the rotor system, and leads to the reduction of mechanical power, or even machine malfunction. Generally, mechanical looseness includes pedestal looseness, base looseness and so on. For a doubledisc singlespan rotor system with pedestal looseness fault, Ma et al. [1] used three dimensional spectrums and shaft center trajectory to analyze the influence of no loose bolt stiffness, looseness gap, and rotational speed on the dynamic characteristics of the rotor system. Sun et al. [2] proposed a finite element model which can characterize complex structures, and verified the effectiveness of the proposed model by comparing critical speed and vibration mode. Zhang et al. [3] proposed a rotor model with nonlinear oil film force, and solved this mode by the numerical method of RungeKutta, and obtained frequency characteristics of the rotor system. Lu et al. [4] proposed a coupling model of the rotor system with looseness and rubimpact faults, and discussed the influence of the looseness stiffness and the clearance between rotor and stator on the dynamic characteristics of the rotor system based on the nonlinear finite element method and the contact theory. Based on nonlinearity measure, Mian et al. [5] proposed an evaluation method of the pedestal looseness in bearingrotor system with piecewiselinear stiffness, damping and nonlinear elastic force under constant rotational speed at constant speed. In Ref. [6,7,8,9], the corresponding model was proposed, and the dynamic characteristics of the proposed model were analyzed. Liu [10] proposed a dynamic model of loosenessrubbing coupling fault in the rotorbearing system with dualdisk and threesupport. Wang [11] proposed a model of rotorsupportcasing in the rotor system with support looseness fault. Yu [12] proposed a dynamic model of rotorbearing system with loose support. Chen [13] proposed a dynamics model in a rotorball bearingstator coupling system with rotor unbalanceloose coupling fault. For the complex movement of a single loose bearing pedestal rotorbearing system, Zhang [14] proposed a nonlinear nonsteady oilfilm force model with short bearings. Cao et al. [15] proposed a nonlinear dynamic equation of rotorbearing system with looseness fault between the pedestal and the casing. Xu et al. [16] proposed a dynamics model of dual rotor system with loose bearing pedestal based on the dual rotor structure of experiment rig. Yang et al. [17] analyzed the vibration features of the rotor system with respect to the effects of geometrical nonlinearity, rotorstator rub and pedestal looseness, and revealed the change rules of resonant characteristic and rub region under different loose stiffness. Cao et al. [18] proposed a vibration differential equation with piecewiselinear stiffness by considering the system with oneside support looseness and the rotor with unbalanced extraneous exciting force. Cao et al. [19] proposed a dynamic model of rotorbearingpedestal system to investigate the vibration characteristics and stability due to fit clearance.
However, the above researches ignore the looseness fault in the discshaft system. Because the working environment of the disc in the rotor system is complex, and the working time is also long, the looseness of the discshaft easily occurs. When the interference between the disc and shaft is insufficient, the relative sliding of discshaft will be easily caused. This relative sliding has an influence on the dynamic characteristics of the rotor system with a loose disc. At present, some progress has been made in the looseness of rotating parts of rotor system [20,21,22,23,24,25]. Behzad [20, 21] proposed the model of the looseness of discshaft system with large clearance based on the hypothesis that the rotating speed of the disc is constant, and the disc and the shaft are always in contact. However this model ignored the collision and friction of discshaft, and assumed that the disc and the shaft rotated at the same speed. Obviously, the engineering practice does not meet these requirements. In practice, the discshaft looseness in the rotor system generally also occurs when the clearance is small, even no clearance [22]. In our works [23, 24], a dynamic model of rotor system with the clearance between the shaft and the disc is proposed in Ref. [23],and a rotor system model with discshaft looseness with nonsteadystate oil film force is proposed in Ref. [24]. Wei [25, 26] studied the discshaft looseness fault in rotor system caused by the excessive speed of the shaft. However these research results in discshaft system do not consider the influence of the discshaft looseness fault on the motion state of the rotor system with the insufficient interference force.
Therefore, the motion differential equations of the rotor system with loose disc caused by the insufficient interference force is proposed based on the contact model of the discshaft system with the microscopic surface morphology, and the numerical simulation and experiment test have been completed. The influence factors of the motion state of the disc, the vibration characteristics of the disc, the variation of the displacement difference of discshaft, and the trajectory of the disc. These obtained outcomes provide theoretical support for the fault diagnosis of the discshaft system with slight loose disc in engineering practice.
2 Mathematical Model
2.1 Contact Model of DiscShaft System with the Microscopic Surface Morphology
Figure 1 is a schematic diagram of discshaft contact [27]. In Figure 1, the outer ring represents the reference plane of the inner diameter of the disc and the inner ring represents the cylindrical surface of the outer diameter of the shaft. The coordinates of the center of the shaft, which is represented by o_{1,} is the point (x_{1}, y_{1}). The symbol r_{1} is the radius of the shaft. The coordinates of the center of the disc, which represented by o_{2}, is the point (x_{2},y_{2}). The symbol r_{2} is the radius of the disc. The extension line of o_{1}o_{2} intersects with the shaft and the disc at point P and point Q respectively. The line o_{1}Q rotated around the center of the shaft by an angle, which is represented by Ï† counterclockwise, and intersects with the shaft and the disc at points B and C respectively.
In the triangle \(\Delta o_{1} o_{2} C\), the line \(\overline{{o_{1} o_{2} }}\) is the relative distance r of discshaft, and \(\overline{{o_{2} C}} = r_{2}\), According to cosine theorem, o_{2}C can be written as
Due to \(\cos \angle o_{2} o_{1} C =  \cos \varphi\), Eq. (1) can be simplified as
Then, at any angle \(\phi\), the clearance between the disc and shaft is as follows
When the height of microconvex, which is represented by z, is higher than the clearance between the disc and shaft, which is represented by \(\Delta d\left( \phi \right)\), The normal force generated by the microconvex can be obtained [24]
where Î³_{1} is the radius of the disc, Î³_{2} is the radius of the shaft, Î³ is the radius of the microconvex, E_{1} is the elastic modulus of the disc, E_{2} is the elastic modulus of the shaft, and E is the elastic modulus of the microconvex. Suppose that the normal distribution function of z on the contact surface is \(\phi \left( z \right)\), the mean of normal force generated by a single microconvex at any angle \(\varphi\) is expressed as
Let \(\eta\) is the density of the microconvex, \(r_{1} d\left( \varphi \right)\) is the length of the microconvex, then the number of microconvex is \(\eta lr_{1} \text{d}\varphi\) on the microarea segment with a width of l. The total force on the entire contact surface can be obtained by integrating Eq. (7)
Since the contact force is symmetric with respect to \(o_{1} o_{2}\), Eq. (8) can be simplified as follows
The radial force generated by a single microconvex at any angle Ï† is \(\Delta F_{r} = \Delta F\cos \varphi\), then the mean of radial force is \(\Delta \overline{{F_{r} }} \left( \varphi \right) = \Delta \overline{F\left( \varphi \right)} \cos \varphi\). Therefore, the total radial force generated over the entire circumference can be expressed as follows
Suppose that the contact of each microconvex meets to the conditions of friction law of Coulomb in the process of contact of discshaft, let \(\mu\) is the friction coefficient at the contact point, and \(\Delta \omega = \omega_{1}  \omega_{2}\), where Ï‰_{1} is the rotation speed of the disc, Ï‰_{2} is the speed of the shaft. Then the tangential force generated by a single microconvex at any angle \(\varphi\) is \(\Delta F_{t} = {\text{sgn}} \left( {\Delta \omega } \right)\Delta F\cos \varphi\), then the mean of the tangential force is \(\Delta \overline{{F_{t} }} \left( \varphi \right) = {\text{sgn}} \left( {\Delta \omega } \right)\Delta \overline{F} \left( \varphi \right)\cos \varphi\). the total tangential force generated over the entire circumference can be expressed as
where \({\text{sgn}} \left( {\Delta \omega } \right)\) is the step function, which is expressed as
The frictional force \(F_{f}\) generated by the entire contact surface of discshaft can be expressed as follows:
The components of the force from the disc on the shaft in two coordinate directions are respectively expressed as follows.
2.2 Motion Equations of the Rotor System with Loose Disc
Figure 2 shows a rotor system with concentrated mass and rigid support. This system is composed of a rotor and a loose disc. From Ref. [20], the influence of gyroscopic effect on the dynamic response of the system is very little, for simplicity, the gyroscopic effect is ignored. The dynamic response of the rotor system can be obtained by integrating the equation of motion.
The vibration equation of the shaft is expressed as follows
where \(m_{1}\), \(c_{1}\) and \(k\) are the mass, damping coefficient and stiffness coefficient of the shaft respectively, \(e_{1}\) and \(\omega\) are the eccentric distance and the rotation speed of the shaft respectively. \(f_{x}\) and \(f_{y}\) are the components of the force from the disc on the shaft in the xdirection and the ydirection respectively.
The vibration equations of the disc are is expressed as follows
where \(m_{2}\) is the mass of the disc , \(c_{2}\) is the vibration damping coefficient of the disc, \(e_{2}\) is the eccentric distance of the disc.
The rotational vibration equation of the disc is expressed as follows
where \(j\) is the moment of inertia of the disc, \(c_{3}\) is the rotational damping coefficient of the disc.
For the convenience of calculation and analysis, dimensionless to the equations [15,16,17], The dimensionless equation of motion can be written as
where Ï„=Ï‰t, X_{i}=x_{i}/Ïƒ, Y_{i}=y_{i}/Ïƒ, \(\dot{X}\)_{i}=dx_{i}/dÏ„, \(\dot{Y}\)_{i}=dy_{i}/dÏ„, \(\ddot{X}\)_{i}=d\(\dot{X}\)_{i}/dÏ„, \(\ddot{Y}\)_{i}=d \(\dot{Y}\)_{i}/dÏ„, i=1,2, \(\omega_{{0}} = \sqrt {{k \mathord{\left/ {\vphantom {k {\left( {m_{1} + m_{2} } \right)}}} \right. \kern\nulldelimiterspace} {\left( {m_{1} + m_{2} } \right)}}}\), Î©=Ï‰/Ï‰_{0}, E_{1}=e_{1}/Ïƒ, E_{2}=e_{2}/Ïƒ, Î¾_{1}=c_{1}/m_{1}Ï‰_{0}, Î¾_{2}=c_{2}/m_{2}Ï‰_{0}, Î¾_{3}=c_{3}/m_{3}Ïƒ_{2}, K=k/m_{1}Ï‰_{0}^{2}, G=g/Ï‰_{0}^{2}Ïƒ, F_{x}=f_{x}/m_{1}Ï‰_{0}^{2}Ïƒ, F_{y}=f_{y}/m_{1}Ï‰_{0}^{2}Ïƒ, J=j/m_{1}Ïƒ_{2}, S=m_{2}/m_{1,} C=4/3lÎ·r_{1}Eâ€™Ïƒâˆš2ÏƒÎ³/Ï€. Ïƒ is the root mean square of the microconvex distribution on contact surface of discshaft.
Due to the nonlinear of Eq. (18), its analytical solution is hardly obtained. Here the fourthorder RungeKutta [28,29,30] method is used to solve Eq. (18). In order to ensure the convergence of the solution and reduce the calculation error, the time step of the solution is set to \({{\uppi } \mathord{\left/ {\vphantom {{\uppi } {640}}} \right. \kern\nulldelimiterspace} {640}}\).
3 Numerical Analysis
In this model, the parameters is \(\omega_{0} = 500\), \(E_{1} = 1\), \(E_{2} = 4\), \(G = {{98} \mathord{\left/ {\vphantom {{98} {25}}} \right. \kern\nulldelimiterspace} {25}}\), \(s = {1 \mathord{\left/ {\vphantom {1 3}} \right. \kern\nulldelimiterspace} 3}\), \(J = {1 \mathord{\left/ {\vphantom {1 3}} \right. \kern\nulldelimiterspace} 3} \times 10^{7}\), \(\sigma = 10^{{{  }5}}\), the initial condition is that the disc is stationary.
3.1 Analysis of Motion State of the Disc
When the discshaft looseness fault occurs, the speed of the disc is no longer the same as speed of the shaft, and the motion state will inevitably change, this is also the unique characteristic of this fault. From Eqs. (10), (11), (13) and (18), the speed of the shaft Î©, contact stiffness C, clearance H, damping of the disc Î¾_{2} and rotational damping Î¾_{3} have an influence on the motion characteristics of the discshaft system. Therefore, the motion state of the disc is discussed mainly from the above several factors.
3.1.1 Influence of Shaft Speed
The change of relative speed of the disc \(\theta^{\prime}_{1}\) at different rotation speed of the shaft Î© is shown in Figure 3. The change of relative speed of the disc \(\theta^{\prime}_{1}\) with speed of the shaft Î© is shown in Figure 4. From Figure 3a, when Î©â‰¤1, the shaft speed of the disc is basically the same as that of the shaft. From Figure 3b, when \(1 < {\varOmega} \le 2.6\), The disc is no longer at the same speed as the shaft, \(\theta^{\prime}_{1}\) decreases with the increase of \({\varOmega}\). From Figure 3c, when \(2.6 < {\varOmega} \le 5.8\), the rotation speed of the disc is no longer stable and in a surge state. From Figure 3d, when \(5.8 < {\varOmega}\), the rotation speed of the disc returns to a stable state again, and the relative speed of the disc \(\theta^{\prime}_{1}\) continues to decrease with the increase of \({\varOmega}\). From Figure 4, when \({\varOmega}\) is low, the disc and the shaft basically rotate at the same speed. With the increasing of \({\varOmega}\), \(\theta^{\prime}_{1}\) decreases gradually.
3.1.2 Influence of the Contact Stiffness of DiscShaft
Due to the contact stiffness of discshaft \(C = {4 \mathord{\left/ {\vphantom {4 3}} \right. \kern\nulldelimiterspace} 3}l\eta r_{1} E^{^{\prime}} \sigma \sqrt {{{2\sigma \gamma } \mathord{\left/ {\vphantom {{2\sigma \gamma } \pi }} \right. \kern\nulldelimiterspace} \pi }}\), i.e., there is a linear relationship between \(C\) and the contact stiffness of discshaft. Therefore, the influence of the elastic modulus on the motion state of the disc is known by discussing the influence of \(C\) on the motion state of the disc. Figure 5 is the change of relative speed of the disc \(\theta^{\prime}_{1}\) under the different C.
When \(C \le 4 \times 10^{6} \,\text{N/m}\), \(\theta^{\prime}_{1}\) gradually increased with the increase of C, which is seen in Figure 5a. When \(4 \times 10^{6} \,\text{N/m} < C \le 6 \times 10^{7} \,\text{N/m}\), the rotational speed of the disc is in the state of surge, and \(\theta^{\prime}_{1}\) gradually increases with the increases of C, which is seen in Figure 5b. When \(6 \times 10^{7} \,\text{N/m} < C \le 3 \times 10^{8} \,\text{N/m}\), the rotation speed of the disc recovers to the stable state again, and \(\theta^{\prime}_{1}\) continues to increase with the increases of C, which is seen in Figure 5c. When \(C> 3 \times 10^{8} \,\,\text{N/m}\), the rotation speed of the disc is the same as that of the shaft. With the increase of C, the rotation state of the disc is more stable. The increase of C is the increase of the contact stiffness, also is the increase of the interference force and the reduction of the looseness fault of the discshaft, therefore the rotation speed of the disc is getting more stable.
3.1.3 Influence of the Clearance of the DiscShaft
Let H is the clearance of the discshaft. Figure 6 shows the change of relative speed of the disc \(\theta^{\prime}_{1}\) under the different clearance of the discshaft H. From Figure 6, \(\theta^{\prime}_{1}\) gradually decreases with the increases of H. The reason is that the interference force of discshaft system gradually decreases with the increases of H. The influence of the interference force is consistent with the rotation speed of the shaft and the contact stiffness of discshaft.
3.1.4 Influence of Damping
Figure 7 is the change of \(\theta^{\prime}_{1}\) at \(\xi_{1} = 1/3\). Firstly, the influence of \(\xi_{1}\) is analyzed. From Figure 7, under the condition of \(\xi_{2} = 2\) and \(\xi_{3} = {1 \mathord{\left/ {\vphantom {1 3}} \right. \kern\nulldelimiterspace} 3} \times 10^{9}\), With the change of \(\xi_{1}\), the simulation result demonstrates that the rotation state of the disc is almost constant, that is, the motion state of the disc is not sensitive to the change of \(\xi_{1}\).
Next, the influence of \(\xi_{2}\) is analyzed, under the condition of \(\xi_{1} = {1 \mathord{\left/ {\vphantom {1 3}} \right. \kern\nulldelimiterspace} 3}\) and \(\xi_{3} = {1 \mathord{\left/ {\vphantom {1 3}} \right. \kern\nulldelimiterspace} 3} \times 10^{9}\), With the change of \(\xi_{2}\), the simulation result shows that when \(\xi_{2} \ge 12\), \(\theta^{\prime}_{1}\) is always stable at 0.61. However, when \(\xi_{2} \le 12\), the vibration amplitude of the disc becomes larger and larger, the rubimpact of the discshaft also gets more and more intense. The rotation speed of the disc is sharply increased from a certain value, which is lower than the rotational speed of the shaft, to fluctuation at the rotational speed of the shaft, and the sharp increase of rotating speed and serious vibration of the disc occur simultaneously. The serious change of the discshaft rubbing causes the frictional force to change drastically, thus causes \(\theta^{\prime}_{1}\) to fluctuate at \(\theta^{\prime}_{1}\)=1. Figure 8 shows that the change of motion state of the disc at \(\xi_{2} = 0.1\). From Figure 8, the amplitude of the disc is particularly large at this time, and is still increasing, this is obviously abnormal.
Finally, the influence of \(\xi_{3}\) is analyzed under the conditions of \(\xi_{1} = {1 \mathord{\left/ {\vphantom {1 3}} \right. \kern\nulldelimiterspace} 3}\) and \(\xi_{2} = 2\). Figure 9 shows the change of \(\theta^{\prime}_{1}\) with \(\xi_{3}\). The simulation result shows that the relative speed of the disc \(\theta^{\prime}_{1}\) gradually decreases with the increase of \(\xi_{3}\). When \(\xi_{3}\) is smaller, \(\theta^{\prime}_{1}\) decreased rapidly, and when \(\xi_{3}\) is larger, \(\theta^{\prime}_{1}\) decreased slowly. The bigger the rotational damping is, the lower the rotating speed of the disc is.
3.2 TimeFrequency Characteristic Analysis
Figure 10 shows that the timefrequency characteristics of the discshaft when the rotation speed of the disc is the same as that of the shaft. From Figure 10b, the spectrum of X_{2}, i.e., in the horizon direction, is mainly composed of one frequency component. However, there is a process of collision between the disc and the shaft, and the frequency components of the discshaft collision should occur in the spectrum. The discshaft displacement difference R is a parameter that directly describes the process of the discshaft collision. From Figure 10d, one frequency multiplication is its main component in the collision frequency, and the amplitude of other frequency components are extremely weak compared with that in the X_{2} spectrum, therefore the collision frequency of the discshaft is not found in the spectrum of X_{2}.
Figure 11 shows that the timefrequency characteristics of the disc in the xdirection and the timefrequency characteristics of the discshaft displacement difference R when \(\theta^{\prime}_{1} = 0.86\). From Figure 11b, the rotating frequency component of the disc and the rotating frequency component of the shaft the spectrum are both in the spectrum of X_{2}. The collision frequency component also doesnâ€™t occur in the spectrum of X_{2}. Since the rotating frequency component of the disc and the rotating frequency component of the shaft are close, the beat vibration of the disc occurs in the waveform of X_{2}, which is seen in Figure 11a. The periodic similar beat vibration also occurs in the waveform of the displacement difference of the discshaft. The result shows that the displacement difference of the discshaft has a causal relationship with the vibration of X_{2}. From Figure 11d, the spectrum of R is composed of a few low frequency multiplication components and a few weak high frequency components. Similarly, because the amplitude of the collision frequency component is very small, the collision frequency component is also not found in the X_{2} spectrum.
Figure 12 shows that the timefrequency characteristics of the disc in the xdirection and the timefrequency characteristics of the discshaft displacement difference R when \(\theta^{\prime}_{1} = 0.42\). From Figure 12b, the rotating frequency component of the disc and the rotating frequency component of the shaft the spectrum are both in the spectrum of X_{2}. The collision frequency component also doesnâ€™t occur in the spectrum of X_{2}. From Figure 12d, the spectrum of R is composed of one multiplication component. A few weak low frequency multiplication components and a few weak high frequency components also occur in the spectrum of R. Similarly, because the amplitude of the collision frequency component is very small, the collision frequency component is also not found in Figure 12b.
3.3 Motion Trajectory
Figure 13 shows the trajectory of the disc under the different \(\theta^{\prime}_{1}\). From Figure 13, With the decreases of \(\theta^{\prime}_{1}\), the trajectory of the disc changes from circleshape to inner eightshape, and then return to circleshape. In the inner eightshape, the inner ring first gradually becomes smaller and then gradually becomes larger, the outer ring is still getting smaller. The appearance of the inner eightshape is caused by the difference of \(\theta^{\prime}_{1}\).
The reason why the outer ring has been getting smaller is that the centrifugal force is getting smaller due to the smaller and smaller relative speed of the discshaft \(\theta^{\prime}_{1}\).
4 Experimental Research
The experimental system is composed of ZT3 rotor test bench, Bentley displacement sensor and MULLERBBM sampling system. Figure 14 is the ZT3 rotor test bench. The ZT3 rotor test bench consists of the power output system and the rotor system with loose disc. The flexible coupling is used as the demarcation point, the right half part of demarcation point is the power output system, which consists of motor, joint coupling, shaft, bearing seat, flexible coupling and phase bonder. The right side of the flexible coupling is connected to a shaft with a length of 320Â mm, and the left side of the flexible coupling is connected to a shaft with a length of 500Â mm. Both ends of the shaft are supported by the sliding bearings. In order to ensure the accuracy of the experimental results of the rotor system with loose disc, the flexible coupling is used to make the power output system only output torque, not horizontal or vertical vibration. The phase bonder is used to measure the rotational frequency of the shaft by an eddy current sensor. The motor in the experiment is a DC motor with the output power of 250Â W. The speed controller is used to realize stepless speed in the range of 0âˆ’10000Â r / min. The left half part of the demarcation point is a rotor system with loose disc, which consists of a rotating shaft, two bearing seats, and a disc. The diameter of the shaft is 9.5 mm, the mass of the disc is 0.612 kg, the outer diameter of the disc is 76.2 mm, and the distance between the centers of the two bearing seats is 422 mm.
This experiment is used to analyze the vibration characteristics of the rotor system with loose disc. Figure 15 is the structure of the loose disc, which consists of the inner ring and the outer ring. These two parts are assembled together by a cone surface, and the pressure is provided by the screw thread. When the screw thread rotates clockwise, the inner ring and the outer ring are extruded by the cone surface contact, so that the diameter of the inner ring reduces. Therefore, the screw thread is only turned clockwise to fix the disc on the shaft, then turned counterclockwise, thus the clearance can be generated between the disc and the shaft.
4.1 Analysis of the Spectrum
Figure 16 shows that the waveform and spectrum in the xdirection when the disc and the shaft is at the same speed. Figure 17 shows that the waveform and spectrum in the xdirection when the speed of the disc and the speed of the shaft is close. Figure 18 shows that the waveform and spectrum in the xdirection when the disc and the shaft have a large difference in rotation speed. From Figure 16, when the disc and the shaft is at the same speed, the collision frequency is mainly composed of one frequency multiplication component. From Figure 18, when the disc and the shaft have a large difference in rotation speed, compared with Figure 16, except for one frequency multiplication component in the collision frequency, a few weak high frequency multiplication components and a few weak low frequency multiplication components also occur in the collision frequency. From Figure 17, when the speed of the disc and the speed of the shaft is close, the vibration of the disc occurs a beat vibration, the collision frequency is mainly composed of a few low frequency components. The experimental results are consistent with the simulation results.
4.2 Analysis of Motion Trajectory
Figure 19 shows that the trajectory of the disc under the different \(\theta^{\prime}_{1}\). When the disc and the shaft is at the same speed, the trajectory of the disc is basically a circleshape. With the decreases of \(\theta^{\prime}_{1}\), the trajectory changes from circleshape to inner eightshape, and then return to circleshape. In the inner eightshape, the inner ring first gradually becomes smaller and then gradually becomes larger, and the outer ring is still getting smaller. the experimental results are also consistent with the simulation results. The experiment results show that the simulation model is very reasonable.
5 Conclusions
A dynamic contact model of the rotor system with loose disc is proposed based on the contact model of the discshaft system with the microscopic surface topography. The influences of Î©, C, H, Î¾_{2} and Î¾_{3} on the motion state of the disc are discussed. The timefrequency characteristics of the disc and the motion trajectory of the disc are discussed. The experiment results verified the effectiveness of the proposed model. The obtained conclusions are as follows.

(1)
The rotation speed of the shaft, the contact stiffness, the clearance, the damping of the disc, and the rotational damping all have an influence on the motion state of the disc. Compared with the rotor system with a loose disc caused by the clearance fit, The variation of vibration characteristics of the proposed model is smaller and more stable. The rotor system with slight fault is more stable and conforms to objective laws.

(2)
The spectrum of the disc in the x direction is composed of the rotation frequency of the disc and the rotation frequency of the shaft. When the rotation speed of the disc and the rotation speed of the shaft are same, the collision frequency is mainly composed of one frequency multiplication component. When the rotational speed of the disc and the rotation speed of the shaft is close, the beat vibration occurs in the x direction of the disc. Simultaneously, the periodical similar beat vibration phenomenon also occurs in the waveform of the discshaft displacement difference. The collision frequency is mainly composed of low frequency multiplication components and weak high frequency multiplication components. When the disc and the shaft have a large difference in rotation speed, the collision frequency is mainly composed of one frequency multiplication component, a few weak high frequency multiplication components and a few weak low frequency multiplication components.

(3)
With the decrease of the relative rotation speed of the disc, the trajectory of the disc changes from changes from circleshape to inner eightshape, and then return to circleshape. In the inner eightshape, the inner ring first gradually becomes smaller and then gradually becomes larger, and the outer ring is still getting smaller. The appearance of the inner eightshape is caused by difference in the speed between the disc and the shaft. With the rotating speed of the disc is getting lower and lower, the centrifugal force provided by the disc is getting smaller and smaller, and the outer ring of inner eightshape is getting smaller.
References
H Ma, Z Zhang, X Y Tai, et al. Dynamic characteristic analysis of a rotor system with pedestal looseness under two load case. Proceedings of the CSEE, 2012, 32(26): 132â€“137+158.
C Y Sun, R Yang, Y S Chen, et al. Modeling and experiment verification of a casingdualrotor highdimensional system. Journal of Vibration and Shock, 2018, 37(18): 152â€“157 (In Chinese).
J Zhang, B C Wen. A study of frequency characteristics of rotor system with pedestal looseness at two supports. China Mechanical Engineering, 2008(1): 68â€“71 (In Chinese).
Y J Lu, Z H Ren, H Hou, et al. Study on looseness and impactrub coupling faults of vertical dualdisc overhung rotorbearing system. Journal of Vibration, Measurement & Diagnosis, 2007(2): 102â€“107+169.
J Mian, J G Wu, X S Peng, et al. Nonlinearity measure based assessment method for pedestal looseness of bearingrotor systems. Journal of Sound and Vibration, 2017, 411: 232â€“246.
Z Y Qin, Q K Han, F L Chu. Bolt loosening at rotating joint interface and its influence on rotor dynamics. Engineering Failure Analysis, 2016, 59.
Z Y Qin, F L Chu. Numerical studies on timevarying stiffness of diskdrum type rotor with bolt loosening. Journal of Physics: Conference Series, 2015, 628(1).
Y Liu, Z Y Xue, L Jia, et al. Response characteristics of loosenessrubbing coupling fault in rotorsliding bearing system. Mathematical Problems in Engineering, 2017, 2017(PT.11): 1â€“10.
Z P Li, Y G Luo, H L Yao, et al. Dynamics and fault characteristics of rotorbearing system with pedestal looseness. Journal of Northeastern University (Natural Science), 2002, 23(11): 1048â€“1051 (In Chinese).
Y Liu, X Y Tai, H Ma, et al. Loosenessrubbing coupling fault of dualdisk threesupport rotorbearing system. Journal of Aerospace Power, 2013, 28(5): 977â€“982 (In Chinese).
H F Wang, G Chen. Modeling for whole missile turbofan engine vibration with support looseness fault and characteristics of casing response. Journal of Aerospace Power, 2015, 30(3): 627â€“638 (In Chinese).
H Yu, Y H Ma, S Xiao, et al. Mechanical and dynamic characteristics of bearing with looseness on highspeed flexible rotor. Journal of Beijing University of Aeronautics and Astronautics, 2017, 43(8): 1677â€“1683 (In Chinese).
G Chen. Nonlinear dynamics of unbalancelooseness coupling fults of rotorball bearingstator coupling system. Chinese Journal of Mechanical Engineering, 2008, 44(3): 82â€“88 (In Chinese).
J Zhang, H Li, B C Wen. On the pedestal looseness of dynamic oil filmrotor system. Journal of Northeastern University (Natural Science), 2003, 24(10): 880â€“880 (In Chinese).
Q S Cao, Y M Huang. An improved genetic algorithm for pedestal looseness parameter identification in rotorbearing systems. Journal of Vibration, Measurement & Diagnosis, 2018, 38(3): 446â€“453.
H Z Xu, N F Wang, D X Jiang. Bearing pedestal looseness dynamic model of dual rotor system and fault feature. Journal of Aerospace Power, 2016, 31(11): 2781â€“2794 (In Chinese).
Y Yang, Y R Yang, D Q Cao, et al. Response evaluation of imbalancerubpedestal looseness coupling fault on a geometrically nonlinear rotor system. Mechanical Systems and Signal Processing, 2019, 18: 423â€“442.
Q S Cao, Q Xiang, G L Xiong. Incremental harmonic balance method for the study of support looseness faultâ€™s characteristics. Journal of Mechanical Strength, 2015(6): 999â€“1004 (In Chinese).
H R Cao, F Shi, Y M Li, et al. Vibration and stability analysis of rotorbearingpedestal system due to clearance fit. Mechanical Systems and Signal Processing, 2019, 133: 106275.
M Behzad, M Asayesh. Vibration analysis of rotating shaft with loose disk. IJE Transactions B: Applications, 2002, 15(4): 385â€“393.
M Behzad, M Asayesh. Numerical and experimental investigation of the vibration of rotors with loose discs. Proc. IMechE Part C. Mechanical Engineering Science, 2010, 224: 85â€“94.
G X Yang, J L Xie, S X Zhou, et al. Research on the influence of axle design parameters on contact pressure between axle and hub. Journal of the China Railway Society, 2009, 31(3): 31â€“35 (In Chinese).
J Liu, Z N Li, W X Luo. Effects of unsteady oil film force on rotor system with loose disk and shaft. Journal of Vibration and Shock, 2019, 38(17): 268â€“275 (In Chinese).
Z N Li, J Liu, W X Luo, et al. Research on dynamic modeling and simulation of rotors with loose disc. Journal of Mechanical Engineering, 2020, 56(7): 60â€“71 (In Chinese).
S H Wei, W X Lu. Analysis on vibration characteristics of diskshaft system with loose fit. Journal of Dynamics and Control, 2018, 16(3): 244â€“249 (In Chinese).
S H Wei, W X Lu, F L Chu. Speed characteristics of diskshaft system with rotating part looseness. Journal of Sound and Vibration, 2020, 469.
H Wang, K H Qin, D N Zhou. Nonlinear dynamic modeling of rotor system supported by angular contact ball bearings. Mechanical Systems and Signal Processing, 2017, 85: 16â€“40.
H Zhang, X Zhang, N N He. Rotor dynamic analysis of small vehicle gasoline turbocharger in semifloating bearings. Journal of Beijing Institute of Technology, 2011, 20(4): 502â€“508 (In Chinese).
M F Liao, Y Li, M B Song, et al. Dynamics modeling and numerical analysis of rotor with elastic support/dry friction dampers. Transactions of Nanjing University of Aeronautics and Astronautics, 2018, 35(1): 69â€“83 (In Chinese).
L Yang, S P Yang, J J Wang, et al. Nonlinear vibration analysis of locomotive rotor system. Journal of Mechanical Engineering, 2018, 54(18): 97â€“104 (In Chinese).
Acknowledgements
Not applicable.
Authorsâ€™ Information
ZhiNong Li, born in 1966, is currently a professor at Key Laboratory of Nondestructive Testing, Ministry of Education, Nanchang Hangkong University, China. He received his PhD degree from Zhejiang University, China, in 2003. His research interests include intelligent detection and signal processing, mechanical fault diagnosis.
Fang Qiao, born in 1995, is currently a graduate student at Nanchang Hangkong University, China.
WenXiu Lu, born in 1974, is currently an associate professor at Tsinghua university, China. He received his PhD degree from Tsinghua University, China, in 2002.
Jie liu,born in1991, is a graduate student at Nanchang Hangkong University, China, in 2018.
Dong Wang, born in 1984, is currently an associate professor at State Key Laboratory of Mechanical Systems and Vibration, Shanghai Jiao tong University, China. He received his PhD degree from City University of Hong Kong, China, in 2015.
FuLei Chu, born in 1959, is a professor at Tsinghua University, China. He received his PhD degree from Southampton university, England, in 1993.
Funding
Supported by National Natural Science Foundation of China (Grant Nos. 51675258, 51875301, 51265039), State Key Laboratory of Mechanical System and Vibration of ChinaÂ (Grant No. MSV201914), and Laboratory of Science and Technology on Integrated Logistics Support, National University of Defense Technology of ChinaÂ (Grant No. 6142003190210).
Author information
Authors and Affiliations
Contributions
ZL were in charge of the proposed model; FQ and JL wrote the manuscript; WL, DW and FC were in charge of the experiment. All authors read and approved the final manuscript.
Corresponding author
Ethics declarations
Competing Interests
The authors declare no competing financial interests.
Rights and permissions
Open Access This article is licensed under a Creative Commons Attribution 4.0 International License, which permits use, sharing, adaptation, distribution and reproduction in any medium or format, as long as you give appropriate credit to the original author(s) and the source, provide a link to the Creative Commons licence, and indicate if changes were made. The images or other third party material in this article are included in the article's Creative Commons licence, unless indicated otherwise in a credit line to the material. If material is not included in the article's Creative Commons licence and your intended use is not permitted by statutory regulation or exceeds the permitted use, you will need to obtain permission directly from the copyright holder. To view a copy of this licence, visit http://creativecommons.org/licenses/by/4.0/.
About this article
Cite this article
Li, Z., Qiao, F., Lu, W. et al. Vibration Characteristics of Rotor System with Loose Disc Caused by the Insufficient Interference Force. Chin. J. Mech. Eng. 35, 70 (2022). https://doi.org/10.1186/s10033022007241
Received:
Revised:
Accepted:
Published:
DOI: https://doi.org/10.1186/s10033022007241