# Analysis of the Creeping Phenomenon of Linear Feed System Based on the Rigid-Flexible Coupling Model

Authors: Ruijun Liang; Wenlong Hao; Wenfeng Ran; Wenhua Ye
Journal: Shock and Vibration
 (2021)

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## Abstract

A mathematical model of the creeping phenomenon based on the mechanical model of the linear feed system was established. The dynamic characteristic parameters of each fixed joint were obtained by Yoshimura’s integral. Using the method, only the dynamic characteristic parameters of the joint surface per unit area with simple structure need to be studied, and then, the dynamic characteristic parameters of the whole joint surface can be obtained by integration. Based on the principle of the half-power bandwidth method and the frequency response function identification, the dynamic parameters of each moving joint were solved by the method of experimental modal analysis. Through the parameters of the fixed and moving joints, a rigid body model of the feed system and a flexible body model including the power transmission parts (ball screw pair) and the motion guide parts (guide slide pair and rolling bearing) were, respectively, established. And then, a rigid-flexible coupling dynamic model of the feed system was obtained through the constraint relationships between joints. The influence of both the external load and the feed rate on the fluctuation of motion speed of the system was analyzed from this model. The difference between the experimental results and the simulation results on a feed system platform is not greater than 10%, which verifies the creeping phenomenon. This conclusion can provide a basis for the optimization of the dynamic performance of the ball screw linear-feeding workbench.


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## Body

## 1. Introduction

As a moving part of general equipment, a linear worktable’s accuracy has an important influence on the working accuracy of the equipment. The factors that affect the accuracy of motion of the feeding worktable mainly include static errors [1] such as geometric errors, thermal errors of the mechanical parts, and dynamic errors [2] such as deformation and vibration during operation of the worktable, respectively. Among them, the research on the static error has been relatively complete [3–7], while the impact of the dynamic error factors on the accuracy of equipment becomes an issue.When a feeding worktable runs in a state of low speed with a heavy load, the phenomenon of periodic working and halting, both fast and slow, is called the creeping phenomenon, also called the stick-slip motion, which is a type of dynamic error. The time interval between the occurrences of the creeping phenomenon is uncertain [8], which causes an uneven movement of the worktable and seriously affects the working accuracy and the performance of the equipment [9, 10]. At the same time, it also causes wear of the moving parts in the feed system and reduces the service life of the equipment.Creeping on a machine tool feeding system is a form of self-excited vibration, which is generally improved and suppressed by improving the dynamic characteristics of the feed system. It is pointed out that, in [11, 12], the factors affecting the creeping phenomenon mainly include (1) the friction condition of the feed system, (2) the operating speed of the feed system, (3) the stiffness condition of the feed system, and (4) the magnitude of the load, respectively. The relationship between the critical stick-slip speed of the workbench and the structural parameters of the workbench was derived by using the nonlinear least square method in an earlier study [13]. The relationship between the critical stick-slip speed and the axial stiffness of the ball screws was obtained in an earlier study [14].With the continuous application of virtual simulation technology in the study of the creeping phenomenon, different types of friction models [15–18] have been established and verified by the simulation. A two-degree-of-freedom mathematical model [19] was used in an earlier study to simulate the stick-slip motion of the guide moving and agreed with the experimental results. A dynamic model of the creeping phenomenon of the feed system based on the LuGre model [20] was established, and the state of motion of the creeping was obtained through simulation analysis. A dynamic model was established based on Newton’s law and Kirchhoff’s law, and based on the model results, the airgap and external interference were controlled [21, 22]. A study carried out the dynamic simulation of the creeping process [23], and the influence of various factors on the creeping phenomenon was analyzed in it.This paper studies mechanism of the creeping phenomenon and the factors that affect creeping phenomenon, and based on the mechanical model of the linear feed system, a mathematical model of its creeping phenomenon is established. On this basis, the rigid and flexible dynamic coupling model of the feed system was established, and the effects of the moving speed and load on its creeping dynamic performance are simulated and analyzed. The simulation results are verified on our designed feed system performance test platform [24].


## 2. Mathematical Model of Creeping of the Feed System

A feed system usually performs linear reciprocating motion through the motion conversion. Its rigid body model is shown in Figure1, in which the rolling-bearing pair, the screw-nut pair, and the guide-slider pair are flexible joints having relative motion, and the others are fixed joints with fixed connections.Figure 1
Rigid body model of a feed system.The feed system can be simplified as the mechanical model shown in Figure2, the driving part A drives the driven part B to move at a constant speed v, the displacement of B is x, the friction force of the moving pair is F, and the dynamic characteristics are of an equivalent spring having stiffness K and a damper with damping C, respectively.Figure 2
Simplified mechanical model of the feed system.The motion equation of the mechanical model is(1)Kx0+vt−x−Cx˙−v−F=Mx¨,where M is the mass of the driven part, x0 is the compression amount of the spring when the driven part starts to move, and t is the time of movement in the driving part.When the creeping phenomenon occurs, the motion of the driven part becomes unstable and self-excited oscillation takes place. During this time, the extent of expansion and contractionx0+vt−x of the spring changes with time, and the elastic force and the frictional force become unbalanced. Assume the deviation of x relative to the displacement of the steady-state movement is as Δx, and at the beginning of the movement, the friction force decreases with the increase of the speed, so C1<0; then, (1) is(2)MΔ¨x+C+C1Δ˙x+KΔx=0.Solve equation (2) to obtain(3)Δx=e−ςωnt1Asinωdt1+Bcosωdt1,where ς is the damping coefficient of the system, ς=C+C1/2KM, ωn is the natural frequency of the system and ωd is the natural frequency of the system along with damping, ωd≈ωn=K/M, t1 is the movement time of the driven part, and the constants A and B are determined according to the initial conditions, and they are(4)A=vωn−Dς−1,(5)B=vωn2ς−D,and in equations, (4) and (5)(6)D=Fs−Fdv⋅KM,where Fs is the force during the period from the beginning of the movement of the driving part till the start of the movement of the driven part, which is a static friction force, and Fd is the dynamic friction force converted from Fs when the driven part starts to move. Taking the derivative of equation (3) and ignoring the high-order term ς2≪1, speed of the self-excited vibration is obtained as(7)Δ˙x=ve−ςωnt1D−ςsinωnt1−cosωnt1.As seen from equation (7), if(8)e−ςωnt1D0−ςsinωnt1−cosωnt1<1,then the vibration speed of the driven part will gradually attenuate to zero. When,(9)e−ςωnt1D0−ςsinωnt1−cosωnt1=1,that is,(10)D0≈4πς.The critical speed at which the driven part does not creep is(11)v0=Fs−Fd2πC+C1KM.It can be seen from equation (11) that the factors affecting the creeping phenomenon include the dynamic characteristics of the feed system (stiffness K, damping C+C1), the moving speed of the worktable v, and the load mass M. Among these, the dynamic characteristic is an essential factor that affects the creeping phenomenon.


## 3. Joint Parameters

### 3.1. Dynamic Parameters of the Fixed Joints

The fixed joints include the joints connected by bolts, which are base guide, base-bearing seat, workbench slider, workbench-nut seat, and nut-nut seat, and the joints with an interference fit, which are bearing seat-outer ring and inner ring-screw shaft. Based on the characteristic dynamic parameters of unit area of the fixed joints and then according to their area, the dynamic parameters of each fixed joint were obtained by Yoshimura’s integral, as shown in Tables1 and 2.Table 1
Dynamic parameters of the bolted fixed joints.
JointsNormal stiffness,Kn×1010N/mNormal damping,Cn×108N⋅s/mTangential stiffness,Kτ×1010N/mTangential damping,Cτ×108N⋅s/mBase guide43.477.551×10−37.599×10−329.30×10−3Base-bearing seat12.152.111×10−32.124×10−38.190×10−3Workbench slider10.851.891×10−31.903×10−37.338×10−3Workbench-nut seat11.071.923×10−31.935×10−37.462×10−3Nut-nut seat7.831.360×10−31.369×10−35.278×10−3Table 2
Dynamic parameters of the interference fit joints.
JointsAxial stiffness,Ka×1010N/mAxial damping,Ca×108N⋅s/mRadial stiffness,Kr×1010N/mRadial damping,Cr×108N⋅s/mBearing seat-outer ring0.6431.61×10−34.11×10−30.411×10−3Inner ring-screw shaft0.220.55×10−32.75×10−30.275×10−3


### 3.2. Dynamic Parameters of the Motion Joints

#### 3.2.1. Theoretical Calculation

Dynamic parameters of the screw-nut pair, which was an example of the motion joint, were calculated based on Hertz contact theory. A single ball in the screw-nut pair contacts with the screw raceway at pointS and contacts with the nut raceway at point N. The axial preload force is Fa, the contact pressure angle is β, and the contact normal force is P, as shown in Figure 3.Figure 3
Force in the joint of screw nut.Force balance on the nut in an axial direction is(12)Fa−Pzsinβcosλ=0,where z is the number of balls, λ is the helix angle of the screw and λ=arctanPh/πd0, where Ph is the screw lead, and d0 is the screw diameter.According to Hertz theory, the contact deformationsδS and δN of a ball on the two raceways are(13)δS=2πnS18321−μ12E1+1−μ22E221/3∑ρS1/3P2/3,δN=2πnN18321−μ22E2+1−μ32E321/3∑ρS1/3P2/3,where nS and nN are the coefficient of the short axis of the ellipse formed by the contact deformation at the contact point S and N, μ1, μ2, and μ3 are Poisson’s ratio of the screw, ball, and nut, E1, E2, and E3 are their elastic modulus, and the curvature ∑ρS at S and the curvature ∑ρN at N are(14)∑ρS=4db−2rdb+2cosβcosλd0−dbcosβ,∑ρN=4db−2rdb+2cosβcosλd0−dbcosβ,where db is the ball diameter and r is the curvature ratio of the ball to the raceway. In Figure 3, the contact deformation on the screw nut in an axial direction is(15)δa=δS+δNsinβcosλ,and the contact deformation on the screw nut in the radial direction is(16)δr=δS+δNcosβsinλ.By combining equations (12)–(16), the axial and radial stiffness of the screw-nut joint can be obtained. The results are shown in Table 3. Regarding the screw-nut pair as a single-degree-of-freedom vibration system, we can get its natural frequency in axial fa and in radial fr from its axial and radial stiffness, respectively.Table 3
Calculation results for the screw-nut pair.
zdb (mm)d0 (mm)βBasic parameters473.1752545∘Ph (mm)rfa (N)M (kg)51.045790.335Calculation resultsKa (N·m−1)fa/HzKr (N·m−1)fr (Hz)48.6×10619171.83×105117.63


#### 3.2.2. Experimental Modal Analysis

Dynamic parameters of the ball screw pair, the guide-slider pair, and the rolling-bearing pair in the different directions were also identified with the help of the hammering test. The test scheme is shown in Figure4, and the test photos are shown in Figure 5. In the test, the excitation points were fixed, and the vibration pick-up points were moved.Figure 4
Experimental modal test scheme for the motion joints.Figure 5
Test photos. (a) For screw-nut pair. (b) For guide-slider pair. (c) For bearing pair.
(a)(b)(c)The test points for the mode of the ball screw pair are shown in Figure6. To test the axial parameters, the excitation points 1–4 were set on the end face of the nut, and any point was selected as the vibration pick-up point. To test the radial parameters, excitation points 1–12 were set on the circumference of the nut, and point 1 was selected as the vibration pick-up point to avoid nodes. The measured frequency response is shown in Figure 7. We obtained the stiffness and damping of the ball screw pair from the half-power bandwidth method in Figure 7, as shown in Table 4.Figure 6
Test points for screw-nut pair. (a) Test points on the end face. (b) Test points on the circumference.
(a)(b)Figure 7
Frequency response of the screw-nut pair. (a) Axial amplitude frequency. (b) Axial real and imaginary frequency. (c) Radial amplitude frequency. (d) Radial real and imaginary frequency.
(a)(b)(c)(d)Table 4
Stiffness and damping of the screw-nut pair.
Axial stiffness,KaN/mAxial damping,CaN⋅s/mRadial stiffness,KrN/mRadial damping,CrN⋅s/m51.6×106898.881.62×1051.68Comparing the dynamic stiffness in Tables5 and 6, it can be seen that the difference is not as big as 10%. It indicates that the calculated and the experimental results can be mutually verified. The test points for the guide-slider pair is shown in Figure 8(a). To test the vertical parameters, the excitation points 1–6 were set on the plane of the slider, and point 1 was selected as the vibration pick-up point to avoid nodes. To test the lateral parameters, the excitation points 9–12 were set on the side plane of the slider, and point 7 was selected as the vibration pick-up point. The test points for the bearing are shown in Figures 8(b) and 8(c). To test the axial parameters, eight excitation points were set on each end of the bearing seat, and point 1 was selected as the vibration pick-up point to avoid the middle nodes; to test the radial parameters, six excitation points were set on each side of the bearing seat, and point 1 was selected as the vibration pick-up point.Table 5
Dynamic parameters of the guide-slider pair.
Vertical stiffness,KZN/mVertical damping,CZN⋅s/mLateral stiffness,KYN/mLateral damping,CYN⋅s/m52.3×106258.2855.2×106297.17Table 6
Dynamic parameters of bearing pair.
Axial stiffness,KaN/mAxial damping,CaN⋅s/mRadial stiffness,KrN/mRadial damping,CrN⋅s/m9.38×10660.615.02×10668.14Figure 8
Test points. (a) For guide-slider pair. (b) For bearing axial. (c) For bearing radial.
(a)(b)(c)Stiffness and damping of the bearing pair and of the guide-slider pair were then obtained, as shown in Tables5 and 6, respectively.


## 3.1. Dynamic Parameters of the Fixed Joints

The fixed joints include the joints connected by bolts, which are base guide, base-bearing seat, workbench slider, workbench-nut seat, and nut-nut seat, and the joints with an interference fit, which are bearing seat-outer ring and inner ring-screw shaft. Based on the characteristic dynamic parameters of unit area of the fixed joints and then according to their area, the dynamic parameters of each fixed joint were obtained by Yoshimura’s integral, as shown in Tables1 and 2.Table 1
Dynamic parameters of the bolted fixed joints.
JointsNormal stiffness,Kn×1010N/mNormal damping,Cn×108N⋅s/mTangential stiffness,Kτ×1010N/mTangential damping,Cτ×108N⋅s/mBase guide43.477.551×10−37.599×10−329.30×10−3Base-bearing seat12.152.111×10−32.124×10−38.190×10−3Workbench slider10.851.891×10−31.903×10−37.338×10−3Workbench-nut seat11.071.923×10−31.935×10−37.462×10−3Nut-nut seat7.831.360×10−31.369×10−35.278×10−3Table 2
Dynamic parameters of the interference fit joints.
JointsAxial stiffness,Ka×1010N/mAxial damping,Ca×108N⋅s/mRadial stiffness,Kr×1010N/mRadial damping,Cr×108N⋅s/mBearing seat-outer ring0.6431.61×10−34.11×10−30.411×10−3Inner ring-screw shaft0.220.55×10−32.75×10−30.275×10−3


## 3.2. Dynamic Parameters of the Motion Joints

### 3.2.1. Theoretical Calculation

Dynamic parameters of the screw-nut pair, which was an example of the motion joint, were calculated based on Hertz contact theory. A single ball in the screw-nut pair contacts with the screw raceway at pointS and contacts with the nut raceway at point N. The axial preload force is Fa, the contact pressure angle is β, and the contact normal force is P, as shown in Figure 3.Figure 3
Force in the joint of screw nut.Force balance on the nut in an axial direction is(12)Fa−Pzsinβcosλ=0,where z is the number of balls, λ is the helix angle of the screw and λ=arctanPh/πd0, where Ph is the screw lead, and d0 is the screw diameter.According to Hertz theory, the contact deformationsδS and δN of a ball on the two raceways are(13)δS=2πnS18321−μ12E1+1−μ22E221/3∑ρS1/3P2/3,δN=2πnN18321−μ22E2+1−μ32E321/3∑ρS1/3P2/3,where nS and nN are the coefficient of the short axis of the ellipse formed by the contact deformation at the contact point S and N, μ1, μ2, and μ3 are Poisson’s ratio of the screw, ball, and nut, E1, E2, and E3 are their elastic modulus, and the curvature ∑ρS at S and the curvature ∑ρN at N are(14)∑ρS=4db−2rdb+2cosβcosλd0−dbcosβ,∑ρN=4db−2rdb+2cosβcosλd0−dbcosβ,where db is the ball diameter and r is the curvature ratio of the ball to the raceway. In Figure 3, the contact deformation on the screw nut in an axial direction is(15)δa=δS+δNsinβcosλ,and the contact deformation on the screw nut in the radial direction is(16)δr=δS+δNcosβsinλ.By combining equations (12)–(16), the axial and radial stiffness of the screw-nut joint can be obtained. The results are shown in Table 3. Regarding the screw-nut pair as a single-degree-of-freedom vibration system, we can get its natural frequency in axial fa and in radial fr from its axial and radial stiffness, respectively.Table 3
Calculation results for the screw-nut pair.
zdb (mm)d0 (mm)βBasic parameters473.1752545∘Ph (mm)rfa (N)M (kg)51.045790.335Calculation resultsKa (N·m−1)fa/HzKr (N·m−1)fr (Hz)48.6×10619171.83×105117.63


### 3.2.2. Experimental Modal Analysis

Dynamic parameters of the ball screw pair, the guide-slider pair, and the rolling-bearing pair in the different directions were also identified with the help of the hammering test. The test scheme is shown in Figure4, and the test photos are shown in Figure 5. In the test, the excitation points were fixed, and the vibration pick-up points were moved.Figure 4
Experimental modal test scheme for the motion joints.Figure 5
Test photos. (a) For screw-nut pair. (b) For guide-slider pair. (c) For bearing pair.
(a)(b)(c)The test points for the mode of the ball screw pair are shown in Figure6. To test the axial parameters, the excitation points 1–4 were set on the end face of the nut, and any point was selected as the vibration pick-up point. To test the radial parameters, excitation points 1–12 were set on the circumference of the nut, and point 1 was selected as the vibration pick-up point to avoid nodes. The measured frequency response is shown in Figure 7. We obtained the stiffness and damping of the ball screw pair from the half-power bandwidth method in Figure 7, as shown in Table 4.Figure 6
Test points for screw-nut pair. (a) Test points on the end face. (b) Test points on the circumference.
(a)(b)Figure 7
Frequency response of the screw-nut pair. (a) Axial amplitude frequency. (b) Axial real and imaginary frequency. (c) Radial amplitude frequency. (d) Radial real and imaginary frequency.
(a)(b)(c)(d)Table 4
Stiffness and damping of the screw-nut pair.
Axial stiffness,KaN/mAxial damping,CaN⋅s/mRadial stiffness,KrN/mRadial damping,CrN⋅s/m51.6×106898.881.62×1051.68Comparing the dynamic stiffness in Tables5 and 6, it can be seen that the difference is not as big as 10%. It indicates that the calculated and the experimental results can be mutually verified. The test points for the guide-slider pair is shown in Figure 8(a). To test the vertical parameters, the excitation points 1–6 were set on the plane of the slider, and point 1 was selected as the vibration pick-up point to avoid nodes. To test the lateral parameters, the excitation points 9–12 were set on the side plane of the slider, and point 7 was selected as the vibration pick-up point. The test points for the bearing are shown in Figures 8(b) and 8(c). To test the axial parameters, eight excitation points were set on each end of the bearing seat, and point 1 was selected as the vibration pick-up point to avoid the middle nodes; to test the radial parameters, six excitation points were set on each side of the bearing seat, and point 1 was selected as the vibration pick-up point.Table 5
Dynamic parameters of the guide-slider pair.
Vertical stiffness,KZN/mVertical damping,CZN⋅s/mLateral stiffness,KYN/mLateral damping,CYN⋅s/m52.3×106258.2855.2×106297.17Table 6
Dynamic parameters of bearing pair.
Axial stiffness,KaN/mAxial damping,CaN⋅s/mRadial stiffness,KrN/mRadial damping,CrN⋅s/m9.38×10660.615.02×10668.14Figure 8
Test points. (a) For guide-slider pair. (b) For bearing axial. (c) For bearing radial.
(a)(b)(c)Stiffness and damping of the bearing pair and of the guide-slider pair were then obtained, as shown in Tables5 and 6, respectively.


## 3.2.1. Theoretical Calculation

Dynamic parameters of the screw-nut pair, which was an example of the motion joint, were calculated based on Hertz contact theory. A single ball in the screw-nut pair contacts with the screw raceway at pointS and contacts with the nut raceway at point N. The axial preload force is Fa, the contact pressure angle is β, and the contact normal force is P, as shown in Figure 3.Figure 3
Force in the joint of screw nut.Force balance on the nut in an axial direction is(12)Fa−Pzsinβcosλ=0,where z is the number of balls, λ is the helix angle of the screw and λ=arctanPh/πd0, where Ph is the screw lead, and d0 is the screw diameter.According to Hertz theory, the contact deformationsδS and δN of a ball on the two raceways are(13)δS=2πnS18321−μ12E1+1−μ22E221/3∑ρS1/3P2/3,δN=2πnN18321−μ22E2+1−μ32E321/3∑ρS1/3P2/3,where nS and nN are the coefficient of the short axis of the ellipse formed by the contact deformation at the contact point S and N, μ1, μ2, and μ3 are Poisson’s ratio of the screw, ball, and nut, E1, E2, and E3 are their elastic modulus, and the curvature ∑ρS at S and the curvature ∑ρN at N are(14)∑ρS=4db−2rdb+2cosβcosλd0−dbcosβ,∑ρN=4db−2rdb+2cosβcosλd0−dbcosβ,where db is the ball diameter and r is the curvature ratio of the ball to the raceway. In Figure 3, the contact deformation on the screw nut in an axial direction is(15)δa=δS+δNsinβcosλ,and the contact deformation on the screw nut in the radial direction is(16)δr=δS+δNcosβsinλ.By combining equations (12)–(16), the axial and radial stiffness of the screw-nut joint can be obtained. The results are shown in Table 3. Regarding the screw-nut pair as a single-degree-of-freedom vibration system, we can get its natural frequency in axial fa and in radial fr from its axial and radial stiffness, respectively.Table 3
Calculation results for the screw-nut pair.
zdb (mm)d0 (mm)βBasic parameters473.1752545∘Ph (mm)rfa (N)M (kg)51.045790.335Calculation resultsKa (N·m−1)fa/HzKr (N·m−1)fr (Hz)48.6×10619171.83×105117.63


## 3.2.2. Experimental Modal Analysis

Dynamic parameters of the ball screw pair, the guide-slider pair, and the rolling-bearing pair in the different directions were also identified with the help of the hammering test. The test scheme is shown in Figure4, and the test photos are shown in Figure 5. In the test, the excitation points were fixed, and the vibration pick-up points were moved.Figure 4
Experimental modal test scheme for the motion joints.Figure 5
Test photos. (a) For screw-nut pair. (b) For guide-slider pair. (c) For bearing pair.
(a)(b)(c)The test points for the mode of the ball screw pair are shown in Figure6. To test the axial parameters, the excitation points 1–4 were set on the end face of the nut, and any point was selected as the vibration pick-up point. To test the radial parameters, excitation points 1–12 were set on the circumference of the nut, and point 1 was selected as the vibration pick-up point to avoid nodes. The measured frequency response is shown in Figure 7. We obtained the stiffness and damping of the ball screw pair from the half-power bandwidth method in Figure 7, as shown in Table 4.Figure 6
Test points for screw-nut pair. (a) Test points on the end face. (b) Test points on the circumference.
(a)(b)Figure 7
Frequency response of the screw-nut pair. (a) Axial amplitude frequency. (b) Axial real and imaginary frequency. (c) Radial amplitude frequency. (d) Radial real and imaginary frequency.
(a)(b)(c)(d)Table 4
Stiffness and damping of the screw-nut pair.
Axial stiffness,KaN/mAxial damping,CaN⋅s/mRadial stiffness,KrN/mRadial damping,CrN⋅s/m51.6×106898.881.62×1051.68Comparing the dynamic stiffness in Tables5 and 6, it can be seen that the difference is not as big as 10%. It indicates that the calculated and the experimental results can be mutually verified. The test points for the guide-slider pair is shown in Figure 8(a). To test the vertical parameters, the excitation points 1–6 were set on the plane of the slider, and point 1 was selected as the vibration pick-up point to avoid nodes. To test the lateral parameters, the excitation points 9–12 were set on the side plane of the slider, and point 7 was selected as the vibration pick-up point. The test points for the bearing are shown in Figures 8(b) and 8(c). To test the axial parameters, eight excitation points were set on each end of the bearing seat, and point 1 was selected as the vibration pick-up point to avoid the middle nodes; to test the radial parameters, six excitation points were set on each side of the bearing seat, and point 1 was selected as the vibration pick-up point.Table 5
Dynamic parameters of the guide-slider pair.
Vertical stiffness,KZN/mVertical damping,CZN⋅s/mLateral stiffness,KYN/mLateral damping,CYN⋅s/m52.3×106258.2855.2×106297.17Table 6
Dynamic parameters of bearing pair.
Axial stiffness,KaN/mAxial damping,CaN⋅s/mRadial stiffness,KrN/mRadial damping,CrN⋅s/m9.38×10660.615.02×10668.14Figure 8
Test points. (a) For guide-slider pair. (b) For bearing axial. (c) For bearing radial.
(a)(b)(c)Stiffness and damping of the bearing pair and of the guide-slider pair were then obtained, as shown in Tables5 and 6, respectively.


## 4. Creeping Simulation of the Feed System

### 4.1. Rigid-Flexible Coupling Dynamic Model

The process of establishing the dynamic model of the feed system based on rigid-flexible coupling is shown in Figure9. First, the three-dimensional solid model of all components was established. Then, the models of the screw-nut pair, the guide-slider pair, and the rolling-bearing pair were imported into finite element software to generate the modal neutral files (MNFs) for the flexible body. Then, the other components’ solid model and the MNFs were imported into dynamic analysis of mechanical systems. Finally, the fixed and motion joints were defined based on the joint parameters to establish the dynamic model of the rigid-flexible coupling feed system.Figure 9
Flowchart for establishing the dynamic model.The joints were defined by the constraints created among the components, as shown in Figure10. The fixed joints were defined by creating the contact forces according to the dynamic parameters (see Tables 1 and 2). The motion joints of both moving pairs and rotating pairs were defined by creating the sleeve forces based on the dynamic parameters (see Tables 4–6), and the friction coefficients were added to define the moving pairs and rotating pairs. The torque and speed were also specified on the lead screw.Figure 10
Definition of the joint in the rigid-flexible model.


### 4.2. Influence of Working Load and Feed Speed on Creeping

According to the feed system creeping model in Section1, when the dynamic parameters and the friction coefficient have been determined, the factors that affect the feed system creeping phenomenon are the cutting load and feed speed.The feed speed of the workbench was set to 1 mm/s, and the load was to the moving part of the workbench as per the actual working conditions, as shown in Table7 (the force applying part is the Earth). The fluctuations in the speed of the moving part under different loads are obtained through the simulation, as shown in Figure 11. It can be seen that when the feed speed is 1 mm/s, corresponding to the three loads, the maximum fluctuations of the speed of the moving part are 0.0316 mm/s, 0.033 mm/s, and 0.0351 mm/s, respectively, and the maximum fluctuation rates are 3.16%, 3.3%, and 3.51%, respectively. This shows that, as the load increases, the amount of fluctuation in the feed rate increases. It is also seen that when the load increases, the critical speed of the feed system without creeping oscillation will decrease, which is consistent with the result of equation (11).Table 7
Load setting.
No.X-axis (N)Y-axis (N)Z-axis (N)1174151524225222878734344071001Figure 11
Simulated speed under different loads.The working load was given as No. 3 in Table7, and the feed speeds of the workbench were set to 0.5 mm/s, 1 mm/s, and 2 mm/s, respectively. The fluctuations in the speed of the workbench at different feed speeds are obtained through the simulation, as shown in Figure 12. It can be seen that corresponding to the three feed speeds, the maximum fluctuations in the speed are 0.0203 mm/s, 0.0351 mm/s, and 0.0577 mm/s, and the maximum fluctuation rates are 4.06%, 3.51%, and 2.89%, respectively. This shows that, as the feed speed increases, the fluctuation of the feed speed decreases; when the feed system moves at a low speed, the smaller the movement speed, the more serious the creeping phenomenon becomes.Figure 12
Simulated speed under different feed speeds.


## 4.1. Rigid-Flexible Coupling Dynamic Model

The process of establishing the dynamic model of the feed system based on rigid-flexible coupling is shown in Figure9. First, the three-dimensional solid model of all components was established. Then, the models of the screw-nut pair, the guide-slider pair, and the rolling-bearing pair were imported into finite element software to generate the modal neutral files (MNFs) for the flexible body. Then, the other components’ solid model and the MNFs were imported into dynamic analysis of mechanical systems. Finally, the fixed and motion joints were defined based on the joint parameters to establish the dynamic model of the rigid-flexible coupling feed system.Figure 9
Flowchart for establishing the dynamic model.The joints were defined by the constraints created among the components, as shown in Figure10. The fixed joints were defined by creating the contact forces according to the dynamic parameters (see Tables 1 and 2). The motion joints of both moving pairs and rotating pairs were defined by creating the sleeve forces based on the dynamic parameters (see Tables 4–6), and the friction coefficients were added to define the moving pairs and rotating pairs. The torque and speed were also specified on the lead screw.Figure 10
Definition of the joint in the rigid-flexible model.


## 4.2. Influence of Working Load and Feed Speed on Creeping

According to the feed system creeping model in Section1, when the dynamic parameters and the friction coefficient have been determined, the factors that affect the feed system creeping phenomenon are the cutting load and feed speed.The feed speed of the workbench was set to 1 mm/s, and the load was to the moving part of the workbench as per the actual working conditions, as shown in Table7 (the force applying part is the Earth). The fluctuations in the speed of the moving part under different loads are obtained through the simulation, as shown in Figure 11. It can be seen that when the feed speed is 1 mm/s, corresponding to the three loads, the maximum fluctuations of the speed of the moving part are 0.0316 mm/s, 0.033 mm/s, and 0.0351 mm/s, respectively, and the maximum fluctuation rates are 3.16%, 3.3%, and 3.51%, respectively. This shows that, as the load increases, the amount of fluctuation in the feed rate increases. It is also seen that when the load increases, the critical speed of the feed system without creeping oscillation will decrease, which is consistent with the result of equation (11).Table 7
Load setting.
No.X-axis (N)Y-axis (N)Z-axis (N)1174151524225222878734344071001Figure 11
Simulated speed under different loads.The working load was given as No. 3 in Table7, and the feed speeds of the workbench were set to 0.5 mm/s, 1 mm/s, and 2 mm/s, respectively. The fluctuations in the speed of the workbench at different feed speeds are obtained through the simulation, as shown in Figure 12. It can be seen that corresponding to the three feed speeds, the maximum fluctuations in the speed are 0.0203 mm/s, 0.0351 mm/s, and 0.0577 mm/s, and the maximum fluctuation rates are 4.06%, 3.51%, and 2.89%, respectively. This shows that, as the feed speed increases, the fluctuation of the feed speed decreases; when the feed system moves at a low speed, the smaller the movement speed, the more serious the creeping phenomenon becomes.Figure 12
Simulated speed under different feed speeds.


## 5. Experimental Verification

We apply a test platform for the feed system, as shown in Figure13 to verify the established simulation model. The platform consists of the feeding unit and the follow-up unit. The follow-up unit moves synchronously with the feeding unit so that the feeding unit receives the load during its movement.Figure 13
Test platform for the feed system. (a) Model. (b) Photo.
(a)(b)The experiments used the working condition of simulation, i.e., the feed rate was set at 1 mm/s, and the three sets of loads in Table7 were applied. During the linear motion of the feed unit, the data from the motor encoder and the grating were collected, and the speed fluctuation was obtained, as shown in Figure 14. The maximum fluctuations of the workbench’s feed speed are 0.028 mm/s, 0.030 mm/s, and 0.031 mm/s to these three sets of loads, i. e., the maximum fluctuation rates are 2.8%, 3.0%, and 3.1%, respectively.Figure 14
Measured speeds under different loads.The workbench was applied the No. 3 load in Table7, the feed speed was set to 0.5 mm/s, 1 mm/s, and 2 mm/s, respectively. The speed fluctuation of the feed movement of the workbench was measured and is shown in Figure 15. The maximum fluctuations of the workbench’s feed speed are 0.019 mm/s, 0.031 mm/s, and 0.054 mm/s to these three feed speeds, i.e., the maximum fluctuation rates are 3.8%, 3.1%, and 2.7%, respectively.Figure 15
Measured speeds at different feed speeds.We listed the measured and simulated feed speed fluctuations under the different loads and feed speeds in Tables8 and 9. It can be seen that the difference between the measurement result and the simulation result is about 10%, which verifies the rigid-flexible coupling model of the feed system.Table 8
Maximum fluctuations of feed speed under different loads.
Load (N)Experiment (mm/s)Simulation (mm/s)Error (%)X.174-Y.151-Z.5240.0280.031611.4X.252-Y.228-Z.7870.0300.0339.1X.434-Y.407-Z.10010.0310.035111.7Table 9
Maximum fluctuations of the feed rate at different feed rates.
Feed speed (mm/s)Experiment (mm/s)Simulation (mm/s)Error (%)0.50.0190.02036.410.0310.035111.720.0540.05776.4


## 6. Conclusions

(1)
The factors affecting the creeping phenomenon include the dynamic parameters of the feed system, the moving speed of the workbench, and the load on the moving parts. The dynamic character is an essential factor affecting the creeping phenomenon. The dynamic parameters are solved through the combination of Yoshimura’s integral and experimental modal analysis method, which can reduce the difficulty and workload of the test and also provides a new way of thinking for the solution of the dynamic parameters of the feed system.(2)
As the load on the moving parts increases, the fluctuation of the feed speed increases, and the critical speed decreases when the creeping phenomenon does not occur.(3)
As the feed speed of the moving parts increases, the fluctuation of the feed speed decreases. When the feed system moves at a low speed, the smaller the speed of movement is, the more obvious the creeping phenomenon becomes.(4)
The difference between the experimental and simulation results is less than 10%, and it is verified that the established rigid-flexible coupling dynamic model of the feed system can accurately simulate the creeping phenomenon of the feed system.



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*Source: 9928733-2021-07-05.xml*