research article analytical modeling of a ball screw feed ......guide module can be preloaded at a...

Research ArticleAnalytical Modeling of a Ball Screw Feed Drive for VibrationPrediction of Feeding Carriage of a Spindle

Lei Zhang,1,2 Taiyong Wang,1,2 Songling Tian,1,2 and Yong Wang1,3

1School of Mechanical Engineering, Tianjin University, Tianjin 300072, China2Key Laboratory of MechanismTheory and Equipment Design of Ministry of Education, Tianjin University, Tianjin 300072, China3School of Mechanical Engineering, Tianjin University of Commerce, Tianjin 300134, China

Correspondence should be addressed to Taiyong Wang; [email protected]

Received 30 July 2016; Revised 21 October 2016; Accepted 14 November 2016

Academic Editor: Jaromir Horacek

Copyright © 2016 Lei Zhang 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.

An analytical modeling approach for ball screw feed drives is proposed to predict the dynamic behavior of the feeding carriage of aspindle. Mainly considering the rigidity of linear guide modules, a ball-screw-feeding spindle is modeled by a mass-spring system.The contact stiffness of rolling interfaces in linear guide modules is accurately calculated according to the Hertzian theory. Next, amathematical model is derived using the Lagrange method.The presented model is verified by conducting modal experiments. It isfound that the simulated results correspond closely with the experimental data. In order to show the applicability of the proposedmathematical model, parameter-dependent dynamics of the feeding carriage of the spindle is investigated.Thework will contributeto the vibration prediction of spindles.

1. Introduction

Ball screw feed drives are frequently used to position thespindle (or stage) to the desired location due to theirhigh stiffness and accuracy. The positioning precision andefficiency directly determine the quality and productivity ofmachine tools [1]. Hence, it is necessary to possess insightinto the effect of ball screw feed drives on the dynamiccharacteristics of spindles or stages.

It has become mainstream to consider the effect of ballscrew feed drives in the study of dynamic characteristics ofstages. For instance, taking the stiffness of the ball bearings,the screw shaft, and the screw-nut interface into account,the vibration characteristics of the stage were investigatedby means of lumped-parameter method [2–5]. Besides, con-sidering the effects of the preload and rigidity of linearguide modules, the vibration characteristics of the stage wereresearched using the FE method [6–9]. In the analyticalmodel of a linear feeding stage [10], the rigidity of linearguide modules was also taken into account. However, therolling interfaces between the guide rail and slider wereoversimplified. Briefly, the ball screw feed drives have a strong

influence on the dynamic behavior of stages. Nevertheless,the modeling accuracy of the linear guide module requiresfurther improvement.

Research on the dynamic behavior of spindles was con-ducted using a variety of approaches [11–13]. The dynamicmodels accounted for the spindle shaft and bearing, theholder and cutter, and the frame structure of machine tools.However, the mechanical characteristics of ball screw feeddrives were rarely taken into consideration. Preloading ofthe ball screw feed drive was considered and the dynamicbehavior of a vertical column-spindle system was analyzedby Hung et al. [14, 15]. But, it is time-consuming to obtainthe desired results with the FE model. Therefore, furtherresearch on the vibration characteristics of spindles under theinfluences of ball screw feed drives should be conducted.

In order to make a contribution to the vibration pre-diction of spindles, an analytical modeling approach isproposed to establish a mathematical model for a ball-screw-feeding spindle system (BSFSS). The BSFSS is simplifiedas a mass-spring system with six degrees of freedom. Therolling interfaces in linear guide modules are emphaticallyconsidered and accurately described. The motion equations

Hindawi Publishing CorporationMathematical Problems in EngineeringVolume 2016, Article ID 2739208, 8 pageshttp://dx.doi.org/10.1155/2016/2739208

2 Mathematical Problems in Engineering

y

z

x

ba

Ball bearing NutFeeding carriage

Ball screw Ball bearing

Linear guidemodule

Spindle

l

lef

Figure 1: Schematic drawing of the BSFSS.

y

z

x

kna

knrkbskb

m

(a)

v

u

w

d d

cc

y

z

x

𝜃

d0

c0

𝜙

𝜑ki1

ki2 ki3

ki4

𝛽

(b)

Figure 2: Simplified mechanical model of the BSFSS.

and numerical solutions are derived.The modal experimentsare conducted for validation of the presented model. Withthe proposed mathematical model, parameter-dependentdynamics of the feeding carriage of the spindle (FCoS) is alsodiscussed. Finally, a conclusion is drawn.

2. Analytical Modeling Approach

2.1. Dynamic Model. A BSFSS considered here is shown inFigure 1. The spindle is fixed on the feeding carriage. Thefeeding carriage driven by the ball screw is mounted onthe machine base through a pair of linear guides. The ballscrew is supported by ball bearings. At the left end, twoangular contact ball bearings are mounted back to back tosustain the axial and radial forces. At the right end, one deepgroove ball bearing is installed to provide the radial force.Thecommercial linear guide module (composed of a linear guideand a slider) has four ball grooves with a circular arc profileforming a point contact at the angle of 45 degrees. The linearguide module can be preloaded at a different amount. Thepreloads are quantified as low preload (Z0, 0.01𝐶), mediumpreload (ZA, 0.06𝐶), and high preload (ZB, 0.11𝐶), where 𝐶denotes the dynamic load rating (7.83 kN) [16].

The joints between the spindle and feeding carriage alsoinfluence the dynamic behavior of the spindle. For simplifi-cation, the joint stiffness is assumed to be infinite. And the

spindle and feeding carriage are represented together as arigid body with a total mass of 𝑚, as depicted in Figure 2(a).The ball bearings at the left end aremodeled by a linear elasticconnection with axial stiffness 𝑘𝑏. The radial stiffness of ballbearings mainly determines the bending vibration of the ballscrew. In the study of the axial and torsional vibration of theball screw, the radial stiffness of ball bearings is normallyneglected [5, 17, 18]. Inspired by those studies, the radialstiffness of ball bearings is not considered in the model here.The screw shaft is simplified as a linear elastic connectionwithaxial stiffness 𝑘bs.The screw-nut interface is endowed with anaxial stiffness 𝑘na and a torsional stiffness 𝑘nr.

The rigidity of the linear guide module is governed by thecontact stiffness of rolling interfaces between rolling balls andgrooves because the guide rail and slider are more rigid instructure. Hence, by neglecting the mass of rolling balls, eachslider is assumed to be supported by four spring elementsin the direction of the contact angle. The spring elementis located at the middle of each ball groove longitudinally.As shown in Figure 2(b), the spring elements are quantifiedwith stiffness values 𝑘𝑖1, 𝑘𝑖2, 𝑘𝑖3, and 𝑘𝑖4 for each linear guidemodule (𝑖 = 1, 2, 3, 4).

A coordinate system for vibration motions of the FCoSis illustrated in Figure 2(b). The origin of the coordinatesystem is the gravity center of the simplified rigid body 𝑚.The feeding direction of the spindle is along 𝑥-axis. 𝑐 ± 𝑐0 and

Mathematical Problems in Engineering 3

𝛽

Q

Q

R2

R1Oc Or

Figure 3: Contact model of the ball and grooves.

𝑑 ± 𝑑0 are the distances from the contact points betweenthe rolling balls and sliders to x-y plane and x-z plane,respectively. As depicted in Figure 1, 𝑎 and 𝑏 are defined asthe distances from the contact points to y-z plane.

According to the aforementioned simplification, theBSFSS is modeled as a mass-spring system. The vibrationmodes associated with the motion degrees of freedom can bedefined. The translational motions 𝑢, V, and 𝑤 are defined aslinear displacements in axial, lateral, and vertical directions,respectively. The displacement in the axial direction is calledaxial vibration mode. The rocking motion 𝜑 about 𝑥-axis isknown as rolling vibrationmode.The rockingmotion𝜙 abouty-axis is pitching vibration mode and the rocking motion 𝜃about z-axis is yawing vibration mode.

2.2. Contact Stiffness of Rolling Interfaces. In the linear guidemodule, the stiffness of the spring elements is determined bythe contact characteristics of rolling interfaces. The contactstiffness of the rolling interface can be calculated accordingto the Hertzian theory [19]. As illustrated in Figure 3, 𝑅1is the radius of the ball, 𝑅2 is the radius of grooves in theguide rail and the slider, 𝛽 is the contact angle, and𝑄 denotesthe contact force. To enhance the accuracy of the model, thecontact is regarded as elastic.

The contact force can be related to the local deformationat the contact point by the Hertzian expression [20]:

𝑄 = 𝑘ℎ𝑑3/2, (1)where 𝑑 is the elastic deformation at the contact point.𝑘ℎ [N/m3/2] representing the Hertz constant is determined as

𝑘ℎ = 4√23𝐵3/2𝐸𝑅1/2 , (2)where

𝐸 = 1 − 𝑢21𝐸1 + 1 − 𝑢22𝐸2 ,

𝑅 = 𝜌11 + 𝜌12 + 𝜌21 + 𝜌22 = 1𝑅1 + 1𝑅1 + 0 − 1𝑅2 ,𝑓 (𝜌) = (𝜌11 − 𝜌12) + (𝜌21 − 𝜌22)𝑅 .

(3)

In the above equations, 𝐸 [m2/N] describes the compre-hensive material property of the ball and groove. 𝐸1 and 𝐸2are Young’s modulus for materials of the ball and groove,respectively. 𝑢1 and 𝑢2 are Poisson’s ratios. 𝑅 [m−1] is thesynthetic curvature of the ball and groove. 𝐵, named as theHertz coefficient, can be obtained based on the value of 𝑓(𝜌)[21]. 𝑓(𝜌) and 𝐵 are both dimensionless.The contact stiffness𝐾𝑛 between the ball and groove can then be obtained using

𝐾𝑛 = d𝑄d𝑑 = 32𝑘ℎ𝑑1/2 = 32𝑘2/3ℎ 𝑄1/3. (4)As revealed in (4), the contact stiffness depends nonlin-

early on the contact force, which is essentially determined bythe preload set on the rolling ball.

2.3. Mathematical Model and Analytical Solution. Thekineticenergy 𝑇 regarding the mass and inertia of the system is

𝑇 = 12𝑚�̇�2 + 12𝑚V̇2 + 12𝑚�̇�2 + 12𝐽𝑥�̇�2 + 12𝐽𝑦�̇�2+ 12𝐽𝑧�̇�2.

(5)

The potential energy𝑈, due to the ball bearing, the ball screw,and the linear guide module, can be written as

𝑈 = 4∑𝑖=1

4∑𝑗=1

12𝑘𝑖𝑗𝛿2𝑖𝑗 + 12𝑘eq𝑢2 + 12𝑘nr𝜑2

= 4∑𝑖=1

12 [𝑘𝑖1 (𝐴 𝑖1 cos𝛽 + 𝐵𝑖1 sin𝛽)2+ 𝑘𝑖2 (𝐴 𝑖2 cos𝛽 − 𝐵𝑖2 sin𝛽)2+ 𝑘𝑖3 (𝐴 𝑖3 cos𝛽 + 𝐵𝑖3 sin𝛽)2+ 𝑘𝑖4 (𝐴 𝑖4 cos𝛽 − 𝐵𝑖4 sin𝛽)2] + 12𝑘eq𝑢2 + 12𝑘nr𝜑2,

(6)

where

𝐴11 = V + (𝑐 + 𝑐0) 𝜑 + 𝑎𝜃,𝐵11 = 𝑤 − (𝑑 + 𝑑0) 𝜑 − 𝑎𝜙,𝐴12 = V + (𝑐 − 𝑐0) 𝜑 + 𝑎𝜃,𝐵12 = 𝑤 − (𝑑 + 𝑑0) 𝜑 − 𝑎𝜙,𝐴13 = V + (𝑐 − 𝑐0) 𝜑 + 𝑎𝜃,𝐵13 = 𝑤 − (𝑑 − 𝑑0) 𝜑 − 𝑎𝜙,𝐴14 = V + (𝑐 + 𝑐0) 𝜑 + 𝑎𝜃,𝐵14 = 𝑤 − (𝑑 − 𝑑0) 𝜑 − 𝑎𝜙,𝐴21 = V − (𝑐 + 𝑐0) 𝜑 + 𝑎𝜃,𝐵21 = 𝑤 + (𝑑 − 𝑑0) 𝜑 − 𝑎𝜙,


𝐴22 = V − (𝑐 − 𝑐0) 𝜑 + 𝑎𝜃,𝐵22 = 𝑤 + (𝑑 − 𝑑0) 𝜑 − 𝑎𝜙,𝐴23 = V − (𝑐 − 𝑐0) 𝜑 + 𝑎𝜃,𝐵23 = 𝑤 + (𝑑 + 𝑑0) 𝜑 − 𝑎𝜙,𝐴24 = V − (𝑐 + 𝑐0) 𝜑 + 𝑎𝜃,𝐵24 = 𝑤 + (𝑑 + 𝑑0) 𝜑 − 𝑎𝜙,𝐴31 = V − (𝑐 + 𝑐0) 𝜑 − 𝑏𝜃,𝐵31 = 𝑤 + (𝑑 − 𝑑0) 𝜑 + 𝑏𝜙,𝐴32 = V − (𝑐 − 𝑐0) 𝜑 − 𝑏𝜃,𝐵32 = 𝑤 + (𝑑 − 𝑑0) 𝜑 + 𝑏𝜙,𝐴33 = V − (𝑐 − 𝑐0) 𝜑 − 𝑏𝜃,𝐵33 = 𝑤 + (𝑑 + 𝑑0) 𝜑 + 𝑏𝜙,𝐴34 = V − (𝑐 + 𝑐0) 𝜑 − 𝑏𝜃,𝐵34 = 𝑤 + (𝑑 + 𝑑0) 𝜑 + 𝑏𝜙,𝐴41 = V + (𝑐 + 𝑐0) 𝜑 − 𝑏𝜃,𝐵41 = 𝑤 − (𝑑 + 𝑑0) 𝜑 + 𝑏𝜙,𝐴42 = V + (𝑐 − 𝑐0) 𝜑 − 𝑏𝜃,𝐵42 = 𝑤 − (𝑑 + 𝑑0) 𝜑 + 𝑏𝜙,𝐴43 = V + (𝑐 − 𝑐0) 𝜑 − 𝑏𝜃,𝐵43 = 𝑤 − (𝑑 − 𝑑0) 𝜑 + 𝑏𝜙,𝐴44 = V + (𝑐 + 𝑐0) 𝜑 − 𝑏𝜃,𝐵44 = 𝑤 − (𝑑 − 𝑑0) 𝜑 + 𝑏𝜙.

(7)

In the above equations, 𝐽𝑥, 𝐽𝑦, and 𝐽𝑧 are moments ofinertia of the simplified rigid body 𝑚 about x-axis, y-axis,and z-axis, respectively. 𝛿𝑖𝑗 is the displacement of the slider.The subscripts 𝑖 = 1, 2, 3, 4 represent the four sliders. Thesubscripts 𝑗 = 1, 2, 3, 4 represent the four rows of grooves ineach slider.

Assume that all the linear guide modules have thesame specifications and preloaded amounts and that all thevibration motions meet the small displacement assumption.The stiffness of the spring elements can then be presented as

𝑘𝑖1 = 𝑘𝑖2 = 𝑘𝑖3 = 𝑘𝑖4 = 𝑘 = 𝐾𝑛2 , 𝑖 = 1, 2, 3, 4. (8)As shown in Figure 2(a), the linear elastic connections 𝑘𝑏,𝑘bs, and 𝑘na are in series. Hence, the equivalent axial stiffness

can be derived as 1𝑘eq = 1𝑘𝑏 + 1𝑘bs + 1𝑘na , (9)

where

𝑘bs = 𝐸𝜋𝐷24𝑙ef , (10)with Young’s modulus 𝐸, the diameter 𝐷, and the equivalentlength 𝑙ef of the screw shaft. The torsional stiffness of thescrew-nut interface can be obtained using

𝑘nr = ( 𝑝2𝜋)2 𝑘na (11)with lead p of the screw shaft.

Applying the Lagrange approach to (5)–(11), the motionequations of the mass-spring system can be derived:

𝑚�̈� + 𝑘eq𝑢 = 0,𝑚V̈ + 16𝑘 cos2𝛽 ⋅ V + 8𝑘 (𝑎 − 𝑏) cos2𝛽 ⋅ 𝜃 = 0,𝑚�̈� + 16𝑘 sin2𝛽 ⋅ 𝑤 + 8𝑘 (𝑏 − 𝑎) sin2𝛽 ⋅ 𝜙 = 0,𝐽𝑥�̈� + 16𝑘 [(𝑐2 + 𝑐20 ) cos2𝛽 + (𝑑2 + 𝑑20) sin2𝛽] 𝜑+ 𝑘nr𝜑 = 0,

𝐽𝑦�̈� + 8𝑘 (𝑎2 + 𝑏2) sin2𝛽 ⋅ 𝜙+ 8𝑘 (𝑏 − 𝑎) sin2𝛽 ⋅ 𝑤 = 0,

𝐽𝑧�̈� + 8𝑘 (𝑎2 + 𝑏2) cos2𝛽 ⋅ 𝜃 + 8𝑘 (𝑏 − 𝑎) cos2𝛽 ⋅ V = 0.

(12)

It is noticed that the first motion equation governs thedisplacement u along x-axis, and the fourth motion equationdescribes the angular displacement 𝜑 about x-axis. Thenatural frequencies including the axial (𝑓𝐴) and rolling (𝑓𝑅)modes can be calculated using

𝑓𝐴 = 12𝜋√𝑘eq𝑚 ,𝑓𝑅 = 12𝜋⋅ √ 16𝑘 [(𝑐2 + 𝑐20 ) cos2𝛽 + (𝑑2 + 𝑑20) sin2𝛽] + 𝑘nr𝐽𝑥 .

(13)

As revealed in (12), the linear displacement v and theangular displacement 𝜃 are highly coupled. The solution isassumed to be the form

V = V ⋅ 𝑒𝑗𝜔𝑡,𝜃 = 𝜃 ⋅ 𝑒𝑗𝜔𝑡. (14)

Substituting (14) into the second and the sixth motionequations of (12), the characteristic equation can be derivedas

𝑚𝐽𝑧𝜔4 − (𝐵1𝐽𝑧 + 𝐵2𝑚)𝜔2 + 𝐵1𝐵2 − 𝐵23 = 0, (15)


where

𝐵1 = 16𝑘 cos2𝛽,𝐵2 = 8𝑘 (𝑎2 + 𝑏2) cos2𝛽,𝐵3 = 8𝑘 (𝑎 − 𝑏) cos2𝛽.

(16)

Similarly, the characteristic equation for the linear displace-ment 𝑤 and the angular displacement 𝜙 can also be derivedas

𝑚𝐽𝑦𝜔4 − (𝐶1𝐽𝑦 + 𝐶2𝑚)𝜔2 + 𝐶1𝐶2 − 𝐶23 = 0, (17)where

𝐶1 = 16𝑘 sin2𝛽,𝐶2 = 8𝑘 (𝑎2 + 𝑏2) sin2𝛽,𝐶3 = 8𝑘 (𝑏 − 𝑎) sin2𝛽.

(18)

Finally, the natural frequencies of the yawing (𝑓𝑌) andpitching (𝑓𝑃)modes are computed as

𝑓𝑌 = 𝜔12𝜋 ,𝑓𝑃 = 𝜔22𝜋 ,

(19)

where

𝜔21= 𝐵1𝐽𝑧 + 𝐵2𝑚 − √(𝐵1𝐽𝑧 + 𝐵2𝑚)2 − 4𝑚𝐽𝑧 (𝐵1𝐵2 − 𝐵23)2𝑚𝐽𝑧 ,𝜔22= 𝐶1𝐽𝑦 + 𝐶2𝑚 − √(𝐶1𝐽𝑦 + 𝐶2𝑚)

2 − 4𝑚𝐽𝑦 (𝐶1𝐶2 − 𝐶23)2𝑚𝐽𝑦 .

(20)

The parameters of the BSFSS are either obtained frommanufacturers’ catalogs or computed from the CAD modelof the components, as listed in Table 1. With the low preloadset on the linear guide modules (𝑘 = 1.9888N/𝜇m) andthe spindle positioned at the middle of its travel range(𝑙ef = 242.5mm), the natural frequencies associated withthe fundamental vibration modes of the FCoS are calculatedand listed in Table 2. Substituting the eigenvalues into thecharacteristic equation, the corresponding eigenvectors canalso be derived, which describe the vibration modes. Theyawingmode (𝑓𝑌) stands for the coupling of the translationalmotion v and the rocking motion 𝜃, in which the rockingmotion 𝜃 is primary. The pitching mode (𝑓𝑃) expresses thecoupling of the translational motion 𝑤 and the rockingmotion 𝜙, where the rocking motion 𝜙 is dominant. Therolling mode (𝑓𝑅)means the rocking motion 𝜑, and the axialmode (𝑓𝐴) represents the translational motion 𝑢.

Table 1: Parameters of the BSFSS.

Parameter Value Unit𝑅1 and 𝑅2 1.389𝑒 − 3 and 1.45𝑒 − 3 m𝐸1 = 𝐸2 2.06e11 Pa𝑢1 = 𝑢2 0.3 —𝛽 𝜋/4 rad𝐸 2.11e11 Pa𝑙,𝐷, and 𝑝 0.78, 0.016, and 0.005 m𝑚 13.7 Kg𝐽𝑥, 𝐽𝑦, and 𝐽𝑧 0.06, 0.016, and 0.12 Kg⋅m2𝑘𝑏 and 𝑘na 2.45e7 and 1.078e8 N/m𝑘 (low, middle, and highpreload) 1.9888e6, 3.6139e6, and 4.4231e6 N/m𝑎 and 𝑏 0.0195 and 0.0445 m𝑐 and 𝑐0 0.093 and 0.0025 m𝑑 and 𝑑0 0.0325 and 0.0075 mTable 2: Predicted and experimental natural frequencies of theBSFSS.

Vibrationmode Prediction/Hz Experiment/Hz Relative error/%

Yawing mode 36.7 35.1 4.56%Pitching mode 86.7 81.8 5.99%Rolling mode 162.1 149.1 8.72%Axial mode 182.0 166.3 9.44%

P1

P2

P3

P4P5 P5

P4y

z

x

Figure 4: Experimental setup.

3. Experimental Validation and Discussion

3.1.Modal Test Validation. As depicted in Figure 4, themodaltest is conducted on a ball-screw-feeding spindle setup. Thelinear guide modules are preloaded at low amounts (Z0)and the spindle is positioned at the middle of its travelrange. Using the impact testing method, the hammeringpoint is placed on the side (small dot) of the feeding carriagein the opposite direction of x-axis. Four accelerometersare arranged at the corners on the upper surface of thefeeding carriage. Another accelerometer is located near thehammering point. The accelerometer at point P1 can obtainthree-direction acceleration signals.The three accelerometersat points P2, P3, and P4 can obtain the acceleration signals inthe direction of z-axis. The last accelerometer at point P5 canobtain the acceleration signals in the direction of x-axis. Thedirections of the signals are consistentwith those of the BSFSScoordinates, and three tests are performed.

With the impulse excitations at point P5, the FRFs ofpoint P1 (in three directions), point P2 (in the direction


P1zP1yP1x

P2zP5x

3582

149

166

231

200 250 30010050 1500Frequency (Hz)

0

0.005

0.01

0.015

0.02

0.025

0.03

Am

plitu

de (d

B)

Figure 5: FRFs of points P1, P2, and P5.

of z-axis), and point P5 (in the direction of x-axis) areobtained using the LMS impact testing system. As illustratedin Figure 5, the natural frequency at 35Hz stands for theyawing dominant vibration, which excites the accelerationin the direction of y-axis. The pitching dominant vibrationat 82Hz and the rolling dominant vibration at 149Hz bothinduce the acceleration in the direction of z-axis. The axialdominant vibration at 166Hz arouses the acceleration inthe direction of x-axis. The coupling vibration of the rigidpitching mode and the flexible bending mode at 231Hz,which is not contained in the mathematical model, alsocauses the acceleration in the direction of z-axis.

With the FRFs, the dynamic parameters at low frequen-cies can be estimated. The experimental natural frequenciesof different vibration modes are listed in Table 2 for com-parison. The maximum relative error between the predictednatural frequency and the experimental natural frequencyis 9.44%. According to the LMS modal analysis software,the modal damping ratio of the vibration modes, rangingfrom 0.83% to 9.05%, can be estimated. The low dampingratio has little effect on the natural frequency estimation.It can be concluded that the predicted results have a goodagreement with the experiment. The deviation may be dueto the inaccuracy or ignorance of various stiffness, damping,and inertia.

3.2. Effects of Structural Parameters. According to the pro-posed mathematical model of the BSFSS, some numer-ical simulations have been conducted to investigate theparameter-dependent dynamics of the FCoS. Figure 6 depictsthe dependence of natural frequencies on the preload set onthe linear guidemodule.This reveals apparent dependence ofnatural frequencies associated with the yawing and pitchingmodes on the preload. It is also shown that the preload hasa strong influence on the natural frequency correspondingto the rolling mode. For the natural frequency of the axialmode, the preload has no effect at all. The reason is that theaxial mode is mainly determined by the stiffness of the ballbearings, the ball screw, and the screw-nut interface ratherthan the rigidity of the linear guide module.

36.749.5

54.8

86.7

116.9129.4

162.1

218.4241.7

182.0

YawingPitching

RollingAxial

ZA ZBZ0Preload (N)

50

100

150

200

250

Freq

uenc

y (H

z)

Figure 6: Dependence of the natural frequencies on the preload.

36.748.1

59.570.8

86.799.1 103.6

105.5

162.1

182

YawingPitching

RollingAxial

10080 907060 120 130110Span of sliders (mm)

40

60

80

100

120

140

160

180

Freq

uenc

y (H

z)

Figure 7: Influence of the span of sliders on the natural frequencies.

On the other hand, Figure 7 presents the influence of thespan of sliders on the natural frequencies. With the span ofsliders increasing from 64mm to 124mm, the yawing naturalfrequency changes from 36.7Hz to 70.8Hz, and the pitchingnatural frequency increases to 105.5Hz from the original86.7Hz. Obviously, the span of sliders has no effect on thenatural frequencies of the rolling mode and axial mode.

As observed from the analysis, the dependence of naturalfrequencies for the FCoS on various structural parametersis distinct. When the preload and span of sliders change,the maximum changing rates of natural frequencies reach49.1% and 21.7%, respectively. Meanwhile, the preload andspan of sliders both have no effect on certain natural


frequencies. Hence, the analysis and obtaining parameter-dependent dynamics of the FCoS are useful and meaningful.The selection of ball screw feed drives can be rapidly finalizedso that the dynamic characteristics of the FCoS can beoptimized in prototyping design.

In the engineering practice, the lateral and vertical defor-mation of the ball screw, which is neglected in the proposedmodel, may have some influence on the yawing and pitchingvibrations of the FCoS, especiallywith the FCoS feeding alongthe ball screw. Hence, the bending deformation of the ballscrew will be additionally considered in the future work.

4. Conclusions

Mainly considering the rolling interfaces in the linear guidemodules, a mathematical model of the BSFSS is derived. Thepredicted results correspond closely with the experimentaldata. The ball screw feed drive, especially the linear guidemodule, is shown to determine the vibration behavior of theFCoS. The bending deformation of the ball screw and thebearing stiffness of the spindle will be additionally consideredto predict the dynamic behavior of spindles.

With the mathematical model, the vibration behavior ofthe FCoS is rapidly analyzed for various structural param-eters. This will be helpful for optimizing the structuralparameters of ball screw feed drives in prototyping design.

Competing Interests

The authors declare that there are no competing interestsregarding the publication of this paper.

Acknowledgments

The authors would like to thank Mr. Guofeng Wang for hishelp in design and manufacturing of the experimental setup.And this work was supported in part by National Natural Sci-ence Foundation of China (Grant no. 51475324) and NaturalScience Foundation of Tianjin (Grant no. 13JCZDJC34000).

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