Research Article  Open Access
Jie Yang, Yanjiong Yue, Rupeng Zhu, Weifang Chen, Miaomiao Li, "Dynamic Characteristics of Encased Differential Gear Train with Journal Bearing", Mathematical Problems in Engineering, vol. 2020, Article ID 2436191, 15 pages, 2020. https://doi.org/10.1155/2020/2436191
Dynamic Characteristics of Encased Differential Gear Train with Journal Bearing
Abstract
Taking the marine encased differential gear train as an example, the relationship between the journal bearing parameters and the meshing force of the transmission system is analyzed. In this paper, the dynamic model of the encased differential gear train with journal bearing is established considering the factors of timevarying meshing stiffness and comprehensive transmission error. In this dynamic model, four stiffnesses and four damping coefficients are applied to characterize the asymmetry and interaction of the oil film stiffness and damping of planet bearing. The system responses are calculated by the Fourier series numerical algorithm. The results show that the introduction of journal bearing in encased differential gear train can contribute to gearbox vibration reduction. Moreover, the planet bearing parameters (e.g., clearancetoradius ratio and eccentricity ratio) of the differential stage affect the meshing forces of both the differential and encased stages. In addition, the influence of the planet bearing parameters of the encased stage on the meshing force of the encased stage is more obvious than that of the differential stage. This work may develop a theoretical analysis framework for the design and manufacture of marine transmission systems in the future.
1. Introduction
The compound planetary gear train, which performs well in compactness, high power density, and large torquetoweight ratio, is widely used in ship, helicopter, wind turbine, etc. As for heavy machinery transmission systems, journal bearings typically support the carrier and planet in compound planetary gear train, and their stiffness and damping are unequal in all directions [1]. Vibration noise is an important factor affecting the stability and reliability of the gearbox. Studying the dynamic characteristics of a compound planetary gear train may help to reduce the vibration noise.
Some researchers [2–5] focused on the construction of the gearrotor dynamics model. Using the finite element method, Baguet and Jacquenot [2] established a gearrotorbearing coupling model, in which the timevarying properties and nonlinearities of gears and bearings were introduced. Gao and Zhang [3] investigated the gearbearing dynamic interactions using a dynamic model of a rotorbearing system with confluence transmission. Theodossiades and Natsiavas [4] studied the dynamic response of the gearrotor coupled system under the action of oil film force and meshing force and not only investigated the oil film stability but also analyzed the amplitudefrequency response. Based on nonlinear theory, Wan and Chang [5] carried out a systematic analysis of the dynamic behaviors of a gearrotorbearing with strongly nonlinear.
Some scholars have thrown more lights to study the influence of backlash, timevarying meshing stiffness, and manufacturing error on the dynamic characteristics of planetary gear train. Based on elastodynamics, Guo and Parker [6] established a pure torsional dynamic model of the planetary gear train and analyzed the natural frequency and parametric sensitivity. Similarly, Zhang et al. [7] analyzed the characteristics of natural frequencies and mode shapes of an encased differential gear train. Wu et al. [8] studied the influence of backlash on the nonlinear characteristics of compound planetary gear set by analyzing the change of backlash with gear wear. Zhu et al. [9] explored the influence of multibacklash on bilateral shock of encased differential planetary gear train. Lin and Parker [10] used a multiscale method to investigate the influence of meshing stiffness waves on the stability of the planetary gear train, and they derived the analytical formula of the relationship between meshing stiffness variation and system dynamic stability. AlShyyab et al. [11] constructed a torsional dynamics model of a discrete twostage planetary transmission system, and they studied the frequency response of the system under stiffness excitation. Considering timevarying meshing stiffness, Xiang et al. [12] developed a torsional nonlinear dynamic model of multistage planetary gear trains and studied the effects of bifurcation parameters such as excitation frequency, backlash, and damping.
In addition, Iglesias et al. [13] investigated the effects of eccentricity and planet pin positioning errors on the quasistatic behavior of planetary gear train for different configurations of support parameters. In terms of multiple errors (e.g., eccentricity error, base pitch error, and assembly error), Mo et al. [14] proposed an improved mathematical model of the loading sharing coefficient for compound planetary gear train, which was verified by experiments. Using a lumped parameter method, Zhai et al. [15] developed the bendingtorsionalpendular coupled vibration model of the wind turbine gearbox transmission system and studied the frequency response of carrier assembly errors (e.g., carrier horizontal assembly error, carrier axis assembly error, and carrier spatial assembly error) to the dynamic meshing force and dynamic transmission error.
In most of the abovementioned studies, the planet bearing assumes a fixed stiffness spring and the bearing force is assumed to be linear with displacement. However, the asymmetry and interaction of oil film stiffness and damping of planet bearing is neglected. To fill this gap, we use four stiffnesses and four damping coefficients to explore the dynamic characteristics of encased differential gear train with journal bearing.
In this study, the marine encased differential gear train with journal bearing is taken as the research object. First, the lumped parameter method is utilized to establish a dynamic model of the encased differential gear train considering timevarying meshing stiffness and comprehensive transmission errors. Second, the asymmetry and interaction of the oil film stiffness and damping of journal bearing is employed to the dynamic model, and the gears and carrier are treated according to the concentrated mass. Third, the timevarying meshing stiffness of herringbone gears is calculated based on the timevarying total contact line length [16]. Fourth, the dynamic equations of coupling journal bearings are constructed by Newton’s second law. The dynamic response of encased differential gear train with journal bearing are solved by the numerical algorithm of Fourier series. Finally, the effects of parameters of planet bearing on the meshing forces are explored in detail.
2. Dynamic Model of Compound Planetary Gear Train
The encased differential gear train consisting of a singledegreeoffreedom encased planetary gear train (encased stage) and a twodegreeoffreedom differential planetary gear train (differential stage) is shown in Figure 1. In the differential stage, No. 1, No. 2, No. 3, No. 4, No. 5, No. 6, and No. 7, respectively, represent the floating ring gear Z_{f1}, the intermediate floating member , the composting ring gear Z_{r1}, the planet Z_{pi} (i = 1, 2, …, N, where N is the number of planets), the sun gear Z_{s1,} the carrier H, and planet bearing. In the encased stage, No. 8, No. 9, No. 10, No. 11, No. 12, and No. 13, respectively, represent the floating ring gear Z_{f2}, the intermediate floating member, the composting ring gear Z_{r2}, the planet Z_{mj} (j = 1, 2, …, M, where M is the number of planets), and the sun gear Z_{s2}, and planet bearing. No. 14 and No. 15 represent journal bearings at input and output shafts, respectively.
Herringbone gears are employed in the sun gears Z_{s1}/Z_{s2} (i.e., No. 5/No. 12) and planet Z_{pi}/Z_{m} (i.e., No. 4/No. 11). The composting ring gear Z_{r1}/Z_{r2} (i.e., No. 3/No. 10) is composed of internal and external helical gears on the same gear tooth germ. The intermediate floating member (i.e., No. 2/No. 9) is composed of internal helical and external spur gears on the same gear tooth germ. The internal helical gear of Z_{r1}/Z_{r2} meshes with the unilateral helical gear of planet Z_{pi}/Z_{mj}, and the external helical gear of Z_{r1}/Z_{r2} meshes with the internal helical gear of . The external spur gear of meshes with the internal spur gear of Z_{f1}/Z_{f2} (i.e., No. 1/No. 8).
Figure 2 shows the dynamic model of the encased differential gear train. The rotational coordinate system based on the rotation angular velocity of carrier H is applied to the differential stage. The fixed coordinate system is used for the encased stage. The degreeoffreedom of the encased differential gear train is 19 + 3(N + M). Displacement parameter X is defined bywhere x_{s1}, x_{pi}, x_{r1}, x_{s2}, x_{mj}, and x_{r2} are torsional line displacement parameters along base circle radius of Z_{s1}, Z_{pi}, Z_{r1}, Z_{s2}, Z_{mj}, and Z_{r2}, respectively. x_{g1}, x_{f1}, x_{g2}, and x_{f2} are torsional line displacement parameters along the pitch circle radius of Z_{f1}, , Z_{f2}, and , respectively. x_{H}, H_{H}, and are torsional, lateral, and longitudinal displacements of carrier H, respectively. H_{s1}, H_{pi}, and H_{r1} are lateral center displacements of Z_{s1}, Z_{pi}, and Z_{r1}, respectively, in their rotational coordinates. , , and are longitudinal center displacements of Z_{s1}, Z_{pi}, and Z_{r1}, respectively, in their rotational coordinates. H_{s2}, H_{mj}, and H_{r2} are lateral center displacements of Z_{s2}, Z_{mj}, and Z_{r2}, respectively, in fixed coordinate. , , and are longitudinal center displacements of Z_{s2}, Z_{mj}, and Z_{r2}, respectively, in fixed coordinate.
2.1. Model for Calculation of Journal Bearing Stiffness and Damping
The dynamic model of the journal bearing is shown in Figure 3. Assuming that displacement and velocity remain small disturbances in the vicinity of the static position O_{0} (x_{0}, y_{0}) [1], the firstorder expansion of dynamic bearing forces (i.e., F_{x} and F_{y}) is defined bywhere is the partial derivative at static position O_{0} (x_{0}, y_{0}).
Here,where F_{x0} and F_{y0} are static bearing forces in x and y directions, respectively. ΔF_{x} and ΔF_{y} are the increments of the oil film force. ξ and η are the increments of oilfilm thickness in x and y directions, respectively. k_{ij} and d_{ij} (i, j = x, y) are the stiffness and damping coefficients of journal bearing, which are obtained bywhere B and d are the width and diameter of the journal bearing, μ is the oil viscosity, ε is the eccentricity ratio, ψ is the clearancetoradius ratio, and K_{ij} and D_{ij} (i, j = x, y) are the corresponding dimensionless stiffness and damping coefficient of journal bearing.
2.2. Model for Calculation of TimeVarying Meshing Stiffness
The planets are evenly distributed among the sun gear in the planetary gear train. There are varying mesh phasing factors in meshing pairs (i.e., sunplanet and ringplanet) because of the influence of the number of teeth. The mesh phasing factor affects the phase of the timevarying meshing stiffness [17]. Therefore, the mesh phasing factor is incorporated into the encased differential gear train. Following Parker and Lin [18], the mesh phasing factor is defined bywhere γ_{sm} and γ_{rm} are the mesh phasing factor of sunplanet and ringplanet, respectively, and Z_{s} and Z_{r} are the teeth number of the sun and ring, respectively.
Maatar and Velex [16] proposed the timevarying meshing stiffness of helical gear pairs based on the timevarying total contact line length. Following Maatar and Velex [16], the timevarying meshing stiffness model of the helical gear is applied to the herringbone gear by a parallel method in this study. The timevarying meshing stiffness model of the herringbone gear is constructed as follows:where k_{0} is the average value of the timevarying meshing stiffness, is dimensionless time, = t/T_{m}, t and T_{m} are meshing time and period, and L() is the instantaneous total contact length of gear pair, which is obtained bywhere γ is the mesh phasing factor, ε_{α} is the endface contact ratio of the herringbone gear, β_{b} is the helical angle of the base circle of the herringbone gear, B is the unilateral tooth width of the herringbone gear, A_{k} and B_{k} are the kthorder harmonic amplitude of the timevarying meshing stiffness, S represents the order of the Fourier series, and the series is truncated at S = 5 in this article.
The parameters A_{k} and B_{k} are calculated bywhere ε_{β} is axial contact ratio of the herringbone gear.
2.3. The Equivalent Displacement and Meshing Forces
The center displacement of each gear is projected into the meshing line direction. We ignored the change of the meshing angle caused by the movement of each gear center. Figure 4 shows the positional relationship sun gear, planet, and composting ring gear in the differential stage.
The equivalent displacement of the ith planetcomposting ring and sunplanet along meshing line direction can be calculated bywhere e_{rpi} and e_{spi} are the equivalent error of the ith planetcomposting ring and sunplanet along meshing line direction.
According to the equivalent displacement calculation method of the differential stage, the equivalent displacement of the jth sunplanet and composting ringplanet along meshing line direction can be calculated by
The planet is supported by planet bearing in the encased differential gear train, and thus the compression deformation between carrier and planet bearing can be obtained bywhere δ_{Hcpi} is the compression deformation of planet bearing along the lateral direction and is the compression deformation along the longitudinal direction.
Elastic and damping meshing forces in the differential and encased stages can be calculated bywhere is the first derivative of equivalent displacement. C_{i} is the meshing damping factor, which can be calculated by [19]where ζ_{i} denotes the constant damping ratios of the gear pair, ζ_{i} = 0.07; k_{0,i} represents the average value of the meshing stiffness in the gear pair, and m_{pi} and are the masses of the driving and driven gears in the gear pair.
2.4. Differential Equations of the Compound Planetary Gear Trains
According to Figure 2 of the dynamic model of the encased differential gear train, differential equations of 19 + 3 (N + M) degreeoffreedom are constructed for the encased differential gear train by the lumped parameter method. Differential equations of the differential stage can be deduced using Newton’s second law as follows.(1)The sun gear Z_{s1}:(2)The ith planet gear Z_{pi}: where F_{Hcpi}(t) is the oil film force of the ith planet bearing in the lateral direction, is the oil film force of the ith planet bearing in the longitudinal direction, F_{Hcpi}(t) and are calculated by equation (3), and F_{pi} is the centrifugal force of the ith planet Z_{pi.}(3)The composting ring gear Z_{r1}:(4)The intermediate floating member and floating ring Z_{f1}:
Differential equations of the encased stage can be deduced using Newton’s second law as follows:(1)The sun gear Z_{s2}:(2)The jth planet gear Z_{mj}: where F_{Hmj}(t) is the oil film force of the jth planet bearing in the lateral direction and is the oil film force of the jth planet bearing in the longitudinal direction.(3)The composting ring gear Z_{r2}:(4)The intermediate floating member and floating ring Z_{f2}:
The differential equation of carrier H can be deduced by Newton’s law:where is the oil film force of journal bearing in the lateral direction and is the oil film force of journal bearing in the longitudinal direction.
3. Numerical Studies and Analysis
The dynamics simulation is performed with a constant power of 1000 kW. Different rotational speeds are applied to the differential stage sun gear. Table 1 lists the parameters of the encased differential gear train. The system response is solved by the numerical calculation method of the Fourier series [20]. The eccentricity ratio of the journal bearing is taken from 0.01 to 0.91. The clearancetoradius ratio ranges from 0.1‰ to 2.5‰. The oil viscosity is 0.04 N·s/m^{2}.

3.1. Comparative Analysis of Spring Bearing and Journal Bearing
Before performing planet bearing parameter analysis, we calculate the dynamic deviation of the meshing force to investigate whether the encased differential gear train with journal bearing is more stable than the encased differential gear train with spring (i.e., planet bearings are replaced by fixed stiffness springs). The dynamic model of the encased differential gear train with spring is constructed by equations (14)–(22), and the stiffness of spring is 6 × 10^{6} N/m. According to the stiffness and damping coefficient of the journal bearing in Section 2.1, we establish the dynamic model of the encased differential planetary gear train with journal bearing. The dynamic deviation of meshing force (i.e., the degree of dynamic meshing force deviates from the theoretical static meshing force) is defined as follows [21]:where S_{e} and S_{i} represent the dynamic deviation of external and internal meshing forces, respectively, N is the number of planets, Q is the sampling times of the dynamic meshing force, and are external and internal instantaneous meshing force, respectively, and F_{s} is the theoretical static meshing force.
Figure 5(a) shows that, in the differential stage, the dynamic deviation curves of the external meshing force of the bearing model and the spring model are basically the same at a low rotational speed. However, the dynamic deviation curve of the external meshing force of the spring model has a peak when the rotational speed is 4800 r/min. It is worth noting that the dynamic deviation curve of the internal meshing force is approximately the same as that of the external meshing force (see Figures 5(a) and 5(b)). Figures 5(c) and 5(d) detail that, in the encased stage, the dynamic deviation curves of the meshing forces of the bearing model and the spring model are substantially the same. These findings indicate that at high rotational speeds, the oil film force of journal bearing can curb the vibration of the encased differential gear train. Therefore, we investigate the influence of the parameters of planet bearings on dynamic meshing forces in the next step.
(a)
(b)
(c)
(d)
3.2. Influence of the Differential Planet Bearing Parameters on Meshing Forces
After modeling differential equations of the encased differential gear train with journal bearing, the dynamic meshing force of differential equations can be calculated by equation (12). Besides, we use the rootmeansquare to describe dynamic meshing forces.
3.2.1. ClearancetoRadius Ratio
To investigate the effects of the differential planet bearing clearancetoradius ratios on meshing forces, we set the clearancetoradius ratio as 0.1‰, 0.5‰, 1‰, 1.5‰, 2‰, and 2.5‰, respectively (see Figure 6). The amplitudefrequency response curves of the meshing forces within the rotation speed range of 500–6000 r/min are calculated.
(a)
(b)
(c)
(d)
Figures 6(a) and 6(b) report that the amplitudefrequency response curves of meshing forces (i.e., external and internal) in the differential stage overlap at low rotational speed and diverge at a high rotational speed. This finding indicates that the impact of the clearancetoradius ratio of the differential planet bearing on the meshing force is small when the rotational speed is low; the impact of clearancetoradius ratio on the meshing force is obvious when the rotational speed is high. Figures 6(c) and 6(d) show that the amplitudefrequency response curves of meshing forces in the encased stage tend to diverge at low rotational speed and tend to overlap at a high rotational speed. Therefore, when the rotational speed in the encased planetary gear train is low, the clearancetoradius ratio of the differential planet bearing should be reduced to diminish its influence on the meshing force to ensure the stability and reliability of the system.
3.2.2. Eccentricity Ratio
To investigate the influence of the eccentricity ratio of the differential planet bearing on meshing forces, the eccentricity ratios amount on the planet bearing are taken as 0.01, 0.19, 0.37, 0.55, 0.73, and 0.91, respectively. We calculate the amplitudefrequency response curves of meshing forces within the rotation speed range of 500–6000 r/min, as shown in Figure 7.
(a)
(b)
(c)
(d)
It can be seen from Figure 7(a) that, for different eccentricity ratios of the differential planet bearing, amplitudefrequency response curves are basically consistent. This finding indicates that the eccentricity ratio of the differential planet bearing has little effect on the external meshing force in the differential planetary gear train. Figure 7(b) reveals that amplitudefrequency response curves tend to diverge when the rotational speed is between 2000 r/min and 5000 r/min. This result indicates that when the rotational speed is between 2000 r/min and 5000 r/min, the eccentricity ratio of the differential planet bearing has a significant influence on the internal meshing force in the differential stage. Amplitudefrequency response curves tend to diverge at low rotational speed and tend to overlap at high rotational speed (see Figures 7(c) and 7(d)), suggesting that the impact of eccentricity ratio of the differential planet bearing on the meshing forces in the encased stage is obvious when the rotational speed is low.
3.3. Influence of the Encased Planet Bearing Parameters on Meshing Forces
3.3.1. ClearancetoRadius Ratio
In this section, the impact of clearancetoradius ratios of the encased planet bearing on meshing forces is explored. Similar to the differential planet bearing, we set the clearancetoradius ratio of the encased planet bearing as 0.1‰, 0.5‰, 1‰, 1.5‰, 2‰, and 2.5‰, respectively, as shown in Figure 8.
(a)
(b)
(c)
(d)
Figures 8(a) and 8(b) uncover that the amplitudefrequency response curves of meshing forces (i.e., external and internal) in the differential stage are substantially uniform for different clearancetoradius ratios of the encased planet bearing. Figures 8(c) and 8(d) show that the amplitudefrequency response curves of meshing forces in the encased stage tend to overlap at low rotational speed and tend to diverge at a high rotational speed. It is worth noting that the amplitudefrequency response curves of meshing forces in the encased stage have a peak when the rotational speed is between 1500 r/min and 2000 r/min. These findings indicate that when the rotational speed is between 1500 r/min and 2000 r/min, the clearancetoradius ratio of the encased planet bearing has an obvious impact on the meshing forces in the encased stage, especially 1‰ and 0.5‰ clearancetoradius ratios.
3.3.2. Eccentricity Ratio
Using the same eccentricity parameters as in Section 3.2.2, the impact of the eccentricity ratio of the encased planet bearing on dynamic meshing forces is investigated, as shown in Figure 9.
(a)
(b)
(c)
(d)
Figures 9(a) and 9(b) show that the corresponding amplitudefrequency response curves are the same regardless of which eccentricity ratio of the encased planet bearing is incorporated into our model. Similar to the clearancetoradius ratio of the differential planet bearing, the eccentricity ratio of the encased planet bearing has a significant impact on the meshing forces in the encased stage when the rotational speed is between 1500 r/min and 2000 r/min (see Figures 9(c) and 9(d)). Besides, the eccentricity ratio of the encased planet bearing has more influence on the internal meshing force than on the external meshing force. Compared with Section 3.2.2, it can be found that the influence of the eccentricity ratio of the encased planet bearing on the meshing force of the encased stage is more significant than that of the differential stage.
4. Conclusions
Considering the factors of timevarying meshing stiffness and comprehensive transmission error, the dynamic model of encased differential gear train with journal bearing is established. The asymmetry and interaction of the oil film stiffness and damping of the planet bearings are employed in this dynamic model. The effects of planet bearing parameters on the meshing forces are explored, and the main conclusions are drawn as follows.
When the rotational speed is low, the stability of the encased differential gear train with journal bearing is basically the same as that of the encased differential gear train with spring. However, when the rotational speed is high, the encased differential gear train with journal bearing is more stable than the encased differential gear train with spring.
The clearancetoradius and eccentricity ratio of the differential planet bearings affect the meshing forces of both the differential stage and encased stage. When the rotational speed is low, the influence of the clearance ratio and eccentricity ratio of the differential planet bearings on the meshing forces of the differential stage is less than that of the encased stage. However, this phenomenon is the opposite when the rotational speed is high.
The clearancetoradius and eccentricity ratio of the encased planet bearing obviously affect the forces meshing of the encased stage, while they slightly affect the meshing forces of the differential stage. When the rotational speed is between 1400 r/min and 2000 r/min, the clearancetoradius and eccentricity ratio of the encased planet bearing have a significant impact on the meshing forces of the encased stage.
Nomenclature
T_{D}/T_{L}:  Input/output torque 
N/M:  Teeth number of planets of the differential/encased stage 
Z_{s1}/Z_{s2}:  Sun gear of the differential/encased stage 
Z_{pi} /Z_{mj}:  The ith/jth planet 
Z_{r1}/Z_{r2}:  Composting ring gear of the differential/encased stage 
Z_{f1}/Z_{f2}:  Floating ring gear of the differential/encased stage 
:  Intermediate floating member of the differential/encased stage 
Z_{s}/Z_{r}:  Teeth number of the sun/ring 
T_{m}/t:  Meshing period/time 
r_{s1b}/r_{s2b}:  Base circle radius of Z_{s1}/Z_{s2} 
r_{f1b}:  Pitch circle radius of Z_{f1} 
r_{c}:  Rotation radius of planet 
k_{0}:  Average meshing stiffness 
α_{n}/β:  Pressure/helical angle of gear 
ε_{α} /ε_{β}:  Endface/axial contact ratio 
β_{b}:  Helical angle of the base circle 
ψ_{i} /ψ_{j}:  Phase angle of Z_{pi}/Z_{mj}: ψ_{i} = 2π(i − 1)/N, ψ_{i} = 2π(j − 1)/M 
α_{rp}:  Meshing angle of Z_{s1} − Z_{pi} 
α_{sp}:  Meshing angle of Z_{r1} − Z_{pi} 
α_{sm}:  Meshing angle of Z_{s2} − Z_{mj} 
α_{rm}:  Meshing angle of Z_{r2} − Z_{mj} 
e_{spi}:  Equivalent error of Z_{r1} − Z_{pi} 
e_{rpi}:  Equivalent error of Z_{r1} − Z_{pi} 
e_{smj}:  Equivalent error of Z_{s2} − Z_{mj} 
e_{rmj}:  Equivalent error of Z_{r2} − Z_{mj} 
K_{s1}/K_{s2}:  Supporting stiffness of Z_{s1}/Z_{s2} 
K_{r1}/K_{r2}:  Supporting stiffness of Z_{r1}/Z_{r2} 
K^{H}_{ξη}/D^{H}_{ξη}:  Stiffness/damping of carrier bearing (ξ, η = X, Y) 
K^{p}_{ξη}/D^{p}_{ξη}:  Stiffness/damping of the ith planet bearing (ξ, η = H, ) 
K^{m}_{ξη}/D^{m}_{ξη}:  Stiffness/damping of the jth planet bearing (ξ, η = H, ) 
K_{f1s2}:  Connection torsion stiffness of Z_{f1} − Z_{s2} 
K_{f2H}:  Connection torsion stiffness of Z_{f2} − H 
:  Contact tangential stiffness of Z_{f1} − 
K_{f1q}:  Contact tangential stiffness of − Z_{r1} 
:  Contact tangential stiffness of Z_{f2} − 
K_{f2q}:  Contact tangential stiffness of − Z_{r2} 
K_{rpi}:  Timevarying meshing stiffness of Z_{r1} − Z_{pi} 
K_{spi}:  Timevarying meshing stiffness of Z_{s1} − Z_{pi} 
K_{rmj}:  Timevarying meshing stiffness of Z_{r2} − Z_{mj} 
K_{smj}:  Timevarying meshing stiffness of Z_{s2} − Z_{mj} 
D_{rp}:  Meshing damping of Z_{r1} − Z_{pi} 
D_{sp}:  Meshing damping of Z_{s1} − Z_{pi} 
D_{rm}:  Meshing damping of Z_{r2} − Z_{mj} 
D_{sm}:  Meshing damping of Z_{s2} − Z_{mj} 
m_{s1}/M_{s1}:  Equivalent/actual mass of Z_{s1} 
m_{p}/M_{p}:  Equivalent/actual mass of Z_{pi} 
m_{r1}/M_{r1}:  Equivalent/actual mass of Z_{r1} 
m_{g1}/m_{f1}:  Equivalent mass of /Z_{f1} 
m_{s2}/M_{s2}:  Equivalent/actual mass of Z_{s2} 
m_{m}/M_{m}:  Equivalent/actual mass of Z_{mj} 
m_{r2}/M_{r2}:  Equivalent/actual mass of Z_{r2} 
/m_{f2}:  Equivalent mass of /Z_{f2} 
m_{c}/M_{c}:  Equivalent/actual mass of H. 
Data Availability
The data used to support the findings of this study are available from the corresponding author upon request.
Conflicts of Interest
The authors declare that there are no conflicts of interest regarding the publication of this paper.
Acknowledgments
This work was supported by the National Key R&D Program of China (no. 2018YFB2001300) and National Natural Science Foundation of China (no. 51775265).
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Copyright © 2020 Jie Yang 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.