1-s2.0-S0007850610001897-main

CIRP Annals - Manufacturing Technology 59 (2010) 781–802

Machine tool spindle units

E. Abele (2)a,*, Y. Altintas (1)b, C. Brecher (2)c

a Institute of Production Management, Technology and Machine Tools (PTW), Technische Universitat Darmstadt, Germanyb Manufacturing Automation Laboratory, University of British Columbia, Vancouver, Canadac Laboratory for Machine Tools and Production Engineering (WZL), RWTH Aachen, Germany

A R T I C L E I N F O

Keywords:

Spindle

Mechatronic

Machine tools

A B S T R A C T

This paper presents the state-of-the-art in machine tool main spindle units with focus on motorized

spindle units for high speed and high performance cutting. Detailed information is given about the main

components of spindle units regarding historical development, recent challenges and future trends. An

overview of recent research projects in spindle development is given. Advanced methods of modeling the

thermal and dynamical behavior of spindle units are shown in overview with specific results.

Furthermore concepts for sensor and actuator integration are presented which all focus on increasing

productivity and reliability.

� 2010 CIRP.

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CIRP Annals - Manufacturing Technology

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1. Introduction

Machine tool spindles basically fulfill two tasks:

� rotate the tools (drilling, milling and grinding) or work piece(turning) precisely in space� transmit the required energy to the cutting zone for metal

removal

Obviously spindles have a strong influence on metal removalrates and quality of the machined parts. This paper reviews thecurrent state and presents research challenges of spindletechnology.

1.1. Historical review

Classically, main spindles were driven by belts or gears and therotational speeds could only be varied by changing either thetransmission ratio or the number of driven poles by electricalswitches.

Later simple electrical or hydraulic controllers were developedand the rotational speed of the spindle could be changed by meansof infinitely adjustable rotating transformers (Ward Leonardsystem of motor control).

The need for increased productivity led to higher speedmachining requirements which led to the development of newbearings, power electronics and inverter systems. The progress inthe field of the power electronics (static frequency converter) led tothe development of compact drives with low-cost maintenanceusing high frequency three-phase asynchronous motors.

Through the early 1980’s high spindle speeds were achievableonly by using active magnetic bearings. Continuous developmentsin bearings, lubrication, the rolling element materials and drive

* Corresponding author.

0007-8506/$ – see front matter � 2010 CIRP.

doi:10.1016/j.cirp.2010.05.002

systems (motors and converters) have allowed the construction ofdirect drive motor spindles which currently fulfill a wide range ofrequirements. A historical review of spindle technology is given inFig. 1.

1.2. Principal setup

Today, the overwhelming majority of machine tools areequipped with motorized spindles. Unlike externally drivenspindles, the motorized spindles do not require mechanicaltransmission elements like gears and couplings. A motor spindlemainly consists of the elements shown in Fig. 2.

The spindles have at least two sets of mainly ball bearingsystems. The bearing system is the component with the greatestinfluence on the lifetime of a spindle. Most commonly the motor isarranged between the two bearing systems.

Due to high ratio of ‘power to volume’ active cooling is oftenrequired, which is generally implemented through water basedcooling. The coolant flows through a cooling sleeve around thestator of the motor and often the outer bearing rings.

Seals at the tool end of the spindle prevent the intrusion of chipsand cutting fluid. Often this is done with purge air and a labyrinthseal.

A standardized tool interface such as HSK and SK is placed at thespindles front end. A clamping system is used for fast automatictool changes. Ideally, an unclamping unit (drawbar) which can alsomonitor the clamping force is needed for reliable machining. Ifcutting fluid has to be transmitted through the tool to the cutter,adequate channels and a rotary union become required features ofthe clamping system.

Today, nearly every spindle is equipped with sensorsfor monitoring the motor temperature (thermistors or thermo-couples) and the position of the clamping system. Additionalsensors for monitoring the bearings, the drive and the processstability can be attached, but are not common in many industrialapplications.

http://dx.doi.org/10.1016/j.cirp.2010.05.002

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http://dx.doi.org/10.1016/j.cirp.2010.05.002

[(Fig._1)TD$FIG]

Fig. 1. Historical review.

[(Fig._2)TD$FIG]

Fig. 2. Sectional view of a motor spindle [courtesy: GMN].

[(Fig._4)TD$FIG]

Fig. 4. Spindles available on the market [PTW].[(Fig._5)TD$FIG]

Fig. 5. Main trends in industry on spindle development.

E. Abele et al. / CIRP Annals - Manufacturing Technology 59 (2010) 781–802782

The focus of this paper is the spindle technology. To clearly limitthe topics of this paper the chosen system boundaries areillustrated in Fig. 3.

1.3. State of the art

Current spindle manufacturers offer wide variety of spindlesolutions for the application areas shown in Fig. 4.

Spindles with high power and high speeds are mainlydeveloped for the machining of large aluminum frames in theaerospace industry. Spindles with extremely high speeds and lowpower are used in electronics industry for drilling printed circuitboards (PCB).

1.4. Actual development areas in industry

Current developments in motor spindle industrial applicationfocus on motor technology, improving total cost of ownership(TCO) and condition monitoring for predictive maintenance (seeFig. 5). Another central issue is the development of drive systemswhich neutralize the existing constraints of power and outputfrequency while reducing the heating of the spindle shaft.
[(Fig._3)TD$FIG]
Fig. 3. Overview of the spindle system and its integration into the machine tool.

Particular attention was paid to the increase of the reliablereachable rotational speeds in the past. However, the focus haschanged towards higher torque at speeds up to 15,000 rpm.Because of Increased requirements in reliability, life-cycle andpredictable maintenance the ‘condition monitoring’ systems inmotor spindles have become more important. Periodic and/orcontinuous observation of the spindle status parameters isallowing detection of wear, overheating and imminent failures.

Understanding the life cycle cost (LCC) of the spindles hassteadily gained importance in predicting their service period withmaintenance, failure and operational costs.

2. Fields of application and specific demands

Spindles are developed and manufactured for a wide range ofmachine tool applications with a common goal of maximizing themetal removal rates and part machining accuracy. Fig. 6 gives anoverview regarding the application areas, material grades and theresultant requirements concerning rotational speed, power, torqueand accuracy.

The work materials range from easy to machine materials likealuminum at high speeds with high power spindles, to nickel and

[(Fig._6)TD$FIG]

Fig. 6. Application areas and branches.

[(Fig._7)TD$FIG]

Fig. 7. Flowchart of spindle analysis.

E. Abele et al. / CIRP Annals - Manufacturing Technology 59 (2010) 781–802 783

titanium alloys which require spindles having high torque andstiffness at low speeds. Cutting work materials with abrasivecarbon or fiber-reinforced plastics (FRP) content need good seals atthe spindle front end.

Spindles for drilling printed circuit boards operate in theangular speed range of 100,000 to 300,000 rpm. The increase inproductivity and speed in this application field over the last fewyears was possible with the development of precision air bearings.

Spindles used in die and mould machining have to fulfill theroughing operations (high performance cutting, HPC) at high feedrates as well as the finishing processes (high-speed cutting, HSC) athigh cutting speeds. Depending on the strategy and the machineryof the mould and die shop either two different machine toolsequipped with two different spindles are used or one machine isequipped with a spindle changing unit. Another possibility is to usea spindle which can fulfill both, HSC and HPC conditions, but thisstill remains a compromise regarding overall productivity.

Aerospace spindles are defined by high power as well as highrotational speeds. Today’s spindles allow a material removal rate(MRR) of more than 10 l of aluminum per minute.

Grinding is a finishing operation where high accuracy isnecessary, which requires stiff spindles with bearings havingminimum runout. The present internal cylindrical grindingspindles have a runout requirement of less than 1 mm.

Spindle units which are used mainly for boring and drillingoperations require high axial stiffness, which is achieved by usingangular contact bearings with high contact angles. On the contrary,high-speed milling operations use spindles with bearings havingsmall contact angles in order to reduce the dependency of radialstiffness on the centrifugal forces.

Contemporary machining centers tend to have multi functionswhere milling, drilling, grinding and sometimes honing operationscan be realized on the same work piece. The bottleneck for theenhancement of the multi-technology machines is still the spindle,which cannot satisfy all the machining operations with the samedegree of performance. Reconfigurable and modular machine toolsrequire interchangeable spindles with standardized mechanical,hydraulic, pneumatic and electrical interfaces.

3. Spindle analysis

The aim of modeling and analysis of spindle units is to simulatethe performance of the spindle and optimize its dimensions duringthe design stage in order to achieve maximum dynamic stiffnessand increased material removal rate with minimal dimensions andpower consumption. Modeling of the cutting process and theprediction of chatter stability can be found in previous key notearticles [10] and are not covered in this article. The mechanical partof the spindle assembly consists of hollow spindle shaft mountedto a housing with bearings. Angular contact ball bearings are mostcommonly used in high-speed spindles due to their low-frictionproperties and ability to withstand external loads in both axial andradial directions. The spindle shaft is modeled by beam, brick orpipe elements in finite element environment. The bearing stiffnessis modeled as a function of ball bearing contact angle, preloadcaused by the external load or thermal expansion of the spindleduring operation. The equation of motion is derived in matrix formby including gyroscopic and centrifugal effects, and solved toobtain natural frequencies, vibration mode shapes and frequencyresponse function at the tool attached to the spindle. If the bearingstiffness is dependent on the speed, or if the spindle needs to besimulated under cutting loads, the numerical methods are used topredict the vibrations along the spindle axis as well as contactloads on the bearings. The model allows the simulation ofinteraction between the cutting process and spindle structure(Fig. 7).

Spindle simulation models allow for the optimization of spindledesign parameters either to achieve maximum dynamic stiffness atall speeds for general operation, or to reach maximum axial depthof cut at the specified speed with a designated cutter for a specific

machining application. The objective of cutting maximum materialat the desired speed without damaging the bearings and spindle isthe main goal of spindle design while maintaining all other qualityand performance metrics, e.g. accuracy and reliability.

3.1. Experimental modeling

The dynamic behavior of an existing spindle is most quicklyobtained by measuring its frequency response function (FRF)between force and displacement at the tool tip. The measured FRFcan be curve fitted to estimate the natural frequencies, dampingratios and stiffness values at a range of frequency where thespindle structure may cause vibrations during machining. Theexperimental measurement of FRFs is practical to assess thedynamic stiffness and identify chatter free cutting conditions inprocess planning of part machining operations. However, thefollowing difficulties need to be kept in mind:

� o
nly a small part of the rotating shaft is accessible for testing,hence modeling of entire spindle is not possible � o perational speed and temperature mainly influence the
eigenvalues, but the measurement of FRFs when the spindlerotates is quit difficult
� c urve fitting or other methods to extract parameters out of the
measured input and output data does not always lead to accurateidentification of the spindle’s dynamic parameters.

Traditionally, the FRF is measured by exciting the spindle at thetool tip by impact hammers or shakers manually. There have beenseveral attempts to measure the FRFs automatically. The authors in[20] use a piezo actuator connected in series with a forcemeasurement sensor and a displacement sensor to estimate theFRF at tool tip at standstill. The authors in [2,121] use anelectromagnetic actuator to excite the shaft and measure thedisplacement with a non-contact probe at various rotating speeds.Various authors use an impact hammer for exciting the shaft. Thistype of excitation is also used for hitting the rotating shaft underhigh speeds as presented in [80]. In this work a special mechanismwas designed to hit the shaft with the impact hammer underrepeatable conditions.

The measured data can also be used to verify or update theresults from theoretical models especially to adjust dampingcoefficients as shown in [87]. Direct methods to achieve parametricmodels for further simulations are an issue of system identificationand cause problems like selecting the order of the model, modelswithout minimum phase and stability issues [123].

3.2. Theoretical modeling

Theoretical models are based on physical laws, and used topredict and improve the performance of spindles during the designstage. The models provide mathematical relation between theinputs F (force, speed) and the outputs q (deflections, bearingloads, and temperature). The mathematical models can be

[(Fig._9)TD$FIG]

Fig. 9. Comparisons of experimentally determined tool-tip FRF and simulated FRF

[7].


expressed in state space forms or by a set of ordinary differentialequations. In both cases linear or nonlinear behavior of thespindles can be modeled.

3.2.1. Mechanical modeling of shaft and housing

Finite element (FE) methods are most commonly used to modelstructural mechanics and dynamics of the spindles. The method isbased on discretization of the structure at finite element locationsby partial derivative differential equations. The analysis belongs tothe class of rotor-dynamic studies where the axis-symmetric shaftis usually modeled by beam elements, which lead to constructionof mass (Me) and stiffness (Ke) matrices.

Timoshenko beam element is most commonly used because itconsiders the bending, rotary inertia and shear effects, hence leadsto improved prediction of natural frequencies and mode shapes[165] of the spindle and was applied by the authors [48,128,156,161,177]. The element PIPE16 of the commonly known FEAsoftware ANSYS is also an implementation of the Timoshenkotheory and use the mass matrix from [178] and stiffness matrixfrom [120].

As an example in the finite element model in Fig. 8, the blackdots represent nodes, and each node has three Cartesiantranslational displacements and two rotations [7]. The pulley ismodeled as a rigid disk, the bearing spacer as a bar element, andthe nut and sleeve as a lumped mass. The spindle in this case hastwo front bearings in tandem and three bearings in tandem at therear. The five bearings are in overall back-to-back configuration.The tool is assumed to be rigidly connected to the tool holderwhich is fixed to the spindle shaft rigidly or through springs withstiffness in both directions translation and rotation. The flexibilityof the spindle mounting has to be reflected in the model of thespindle-machine system. Springs are also used between thespindle housing and spindle head, whose stiffness is obtainedfrom experience.

Mqþ ðCVGÞqþ ðK �V2MV þVCVÞq ¼ F

The vector q is the generalized displacement vector whichincludes the movement of all nodes in the chosen degrees offreedom. The reference frame of the vector q can be fixed to therotating shaft of the spindle or be an inertial frame fixed to space.The symmetric matrices M and K are assembled out of the singleelement matrices Me and Ke as described in detail in [49]. Thesematrices are independent of the chosen reference frame. Thestiffness matrix KB of the bearing support is also added in K. Thesymmetric matrix C is also independent of the chosen referenceframe and contains structural damping of the shaft (rotating) andthe damping of the housing (non-rotating). The skew symmetricmatrix G is often named as gyroscopic matrix but in fact onlyincludes the case of inertial reference frame for pure gyroscopiceffects. In the case of rotating reference frame, the matrix G alsoincludes the coriolis acceleration coupling terms which areproportional to the mass as shown in [99]. The term V2MV onlyexists in the rotating reference frame and adds centrifugal forces tothe system. The matrix MV is symmetric and positive definite andso reduces the stiffness of the system. Therefore the authors in[(Fig._8)TD$FIG]

Fig. 8. The finite element model of the spindle-bearing-machine-tool system from

Fig. 10 [7].

[47,99] describe this term as spin-softening effect. The skewsymmetric matrix CV in case of rotating reference frame carries thedamping of the rotating parts and in the other case carries thedamping of the non-rotating parts.

One way to match the model to the physical system is to solvethe eigenvalue problem of the undamped system and then fit theanalytical solution to the experimental data by adding empiricalmodal damping ratios [87] and/or by changing geometricalparameters to match the natural frequencies which result fromexperimental measurements described in the previous section[7,99]. Therefore the system can be rewritten in state spacenotation and transformed to a Jordan canonical form.

Fig. 9 shows the result of the modeled FRF at tool tip ascompared with the experimentally measured FRFs for the spindlegiven in Fig. 10. The closeness of the experimentally measured andsimulated FRFs determine the accuracy of the theoretical models inanalyzing the spindle behavior during the design stage. Severalresearchers have developed FE packages dedicated to the virtualdesign and performance analysis of spindles [26,7,27].

3.2.2. Mechanical modeling of tool–spindle interfaces

The FRF at tool tip is important for stability predictions of themachine tool. The flexibility of the assembly/interface tool-holderand spindle usually dominates the dynamics of the spindle. Due tothe large number of spindle, holder, and tool combinations thatmay be available in a particular production facility, the requiredtesting time for experimental modeling can be significant. Thereare approaches to model the shaft and housing with finite elementtheory as mentioned above and to couple different tool holder andtools. Erturk et al. [41,42,108] use a receptance coupling andstructural modification method to connect the tool-holder to thespindle shaft. Schmitz et al. [135] presents a receptance couplingsubstructure analysis method for modeling a shrink fit tool holder.This work considers distributed springs and dampers between thetool and holder along the interference contact surface. In differenceto that the conventional modeling uses one concentrated springdamper element between the portions of the tool inside andoutside the holder.

3.3. Modeling of angular contact ball bearings

Angular contact ball bearings (Fig. 11) are commonly used inhigh-speed spindles. The bearings require preloading to preventskidding in order to maintain rotational accuracy and sufficient[(Fig._10)TD$FIG]

Fig. 10. Example of a spindle sketch for modeling [7].

[(Fig._11)TD$FIG]

Fig. 11. Geometry of an angular contact ball bearing [27].

[(Fig._12)TD$FIG]

Fig. 12. Radial stiffness changing with cutting forces for spindle-bearing-system

Fig. 10 [25].


stiffness in both the radial and axial directions to support the basicoperational requirements. Basically, there are two types of bearingpreloads: rigid preload and constant preload (see Fig. 22).

Modeling of the bearing support of the spindle shaft isimportant in predicting the spindle’s structural deformationsduring machining. The bearing parameters can vary during theoperation, which in turn affect the stiffness. For example relativedisplacement of raceways curvature centers, relative speedbetween inner and outer ring, axial and radial load, temperaturedifference between inner and outer rings and contact angles mayvary during cutting.

The basic equations to evaluate the tangential stiffness matrixKB for each bearing under working conditions are based on themodels proposed by Jones [76] and De Mul et al. [33]. The theory ofcontact between balls and rings was based on Hertz [61]. Theeffects of cutting loads [25,78], frictional moments [112,157] andthermal deformations [63,95] have been incorporated to theclassical bearing models by spindle researchers [64,81,25,100,60,11].

The bearing models considers the following conditions:

� r
elative speed between inner and outer rings � s peed-related centrifugal forces and gyroscopic moments acting
on rolling elements
� r
Fig. 13. Experimentally estimated 1st (–�–) and 2nd (–*–) radial eigenfrequencies

of a motorized spindle vs. spindle speed [2].

elative displacement of centers of ring groove curvatures due tothermal expansion

The resulting bearing stiffness KB depends on the size of thebearing balls or rollers, curvature of the bearing rings, and contactangle. However, the bearing dynamics changes with preload,cutting forces, spindle speeds and thermal expansion. These factorslead to a nonlinear modeling of the spindle system. The tangentialstiffness matrix KB of the bearings is only valid at one operatingcondition; hence the system dynamics become nonlinear andappear non-repeatable.

Because of the nonlinear dependency of the radial bearingstiffness on the axial or radial load, the bearing stiffness changeswith cutting force dynamically. In [25] it is shown that the axialforce has a larger effect in the bearing stiffness than radial forcesand therefore the variation of bearing stiffness is matched with thefrequency of cutting forces in axial direction. It can be shown thatin case of periodic cutting loads the matrix KB(t) is periodic at theharmonics of the shaft speed. In Fig. 12 this effect is shown fordifferent preload mechanism.

Besides the aspects of preload and load of bearings in highspindle speeds, the bearing ball centrifugal forces and gyroscopicmoments can be of significant magnitude such that inner ringcontact angles tend to increase and outer ring contact angles tendto decrease. The natural frequency of the system is related to thebearing stiffness, hence it increases with preload due to increasedbearing stiffness, but decrease with spindle speed due tocentrifugal forces as experimentally demonstrated in

[25,2,99,121]. Fig. 13 shows one result of experimental measure-ments of the dynamic behavior of a motorized spindle over a rangeof speed.

3.3.1. Thermal modeling of motorized spindle units

Limits of a spindle’s speed, reliability and performance areusually constrained by properties of its bearings, which areaffected by the uneven thermal expansion of spindle parts anddegraded condition of lubricants due to high temperature [22].

The product of mean bearing diameter Dm [mm] and the speed n

[rpm] called the specific speed coefficient is commonly used for theestimation of limiting speed of rolling bearings. When the productDm times n exceeds 0.5 � 106 mm/min, the operation is consideredas high-speed and the spindle design must be able to deal with theheat produced and subsequent thermal expansion. In extremecases, the Dmn value can be as high as 4 � 106 mm/min (jetlubrication). The limiting speed of each application depends on theamount of heat produced by a particular design and on themechanical ‘‘sensitivity’’ of the design to already developedtemperatures (thermal expansion affecting bearings). The impor-tant factors are: bearing type/size, lubrication, bearing configura-tion, type of preload and overall heat management within thespindle.

There is a link between thermal and mechanical behavior ofspindles. Uneven thermal expansion changes the mechanicalcondition of bearings which in turn affects the amount of heatproduced, which is fed back to the system and further increases the[(Fig._13)TD$FIG]


thermal load and heat. The system can be unstable, and increasingthermal expansion and resulting bearing preload can damage thebearings (see also Fig. 20). Spindle builders use three basicprinciples to avoid excessive preloading of bearings:

� C

Fipr

onstant preload of bearings by springs, hydraulic or piezoactuators ([158,36,105])
� R igid preload, theoretically thermally symmetric design � R adial flexible housing [104]
3.3.2. Heat sources in spindles

Principally, there are three main sources of heat in spindles:

� fr
iction within the bearings � p ower losses by the motor – depends on type (asynchronous or
synchronous) motors and frequency converter [127]
� c utting power – since most heat is transferred to the chip or
localized at the tool’s cutting edge, it usually does not play animportant role. This is a more significant effect in materials withlow thermal conductivity like titanium
� fr iction in tool clamping device
The heat generation in bearings is challenging to predict.Theories presented in [115,55] and [74] distinguish three mainsources of friction which occur in ball-groove contact:

� lo
ad friction caused by rolling and proportional to contact forces � v iscous friction caused by viscosity of lubricant – nonlinearly
proportional to speed and mean diameter of bearing [115]
� s pinning friction (spin/roll ratio) caused by kinematics of rolling
elements. If contact angles differ from ‘‘zero’’ value, the rollingelement necessarily spins in one of the bearing grooves.

Currently none of these theoretical analyses have predicted thetemperature distribution and resulting thermal deformations ofthe spindle bearings with sufficient accuracy. Some authors triedto derive dedicated formulations by using ‘coast tests’ [19,62] (seeFig. 14), or by other experiments [77] measuring passive momentsof bearings. However, generalized models applicable to a widerange of bearings, lubrications and operating conditions have notyet been developed.

3.3.3. Heat sinks and heat transfer

Heat transfer by conduction within spindle parts can bemodeled using 2D or 3D FEM elements, once the boundaryconditions are known. It is difficult to model heat transfer correctlyat thermal joints – between surfaces mounted with clearancewhich changes during thermal expansion, for example theinterface between the bearing outer race and the inner diameterof the housing. The authors in [104] show an example of therelationship between bearing temperature and working clearance[(Fig._14)TD$FIG]

g. 14. Frictional moment measured on a single SKF 7010CD bearing, constant

eload Fa0 introduced by springs [62].

as a function of spindle speed. In such cases dedicated nonlinearthermal resistance elements with properties based on heat transfertheories must be created [69].

Heat sinks are usually caused by convection, conduction andradiation on spindle surfaces with the following sources:

� c
onvection to coolant fluids (motor and bearing cooling) � c onvection and radiation to surrounding air or oil–air lubrication � c onduction through the housing flange to the spindle head
Specifically the coefficients of heat convection betweensurfaces and fluids (air, coolant) are difficult to predict anddepend on many parameters, hence care must be taken to specifythese boundary conditions according to established thermaltheories [69].

3.3.4. History of ‘thermal preload’ prediction

There has been a need to predict temperatures and thermalstability in the past, e.g. to avoid thermally induced seizure ofbearings. In 1967 Burton and Staph [22] derived a general theory ofthermal stability of angular contact ball bearings where theauthors distinguished temperatures of various parts of a simplebearing assembly as stabilizing or destabilizing factors. In 1970and 1972 Carmichael and Davies [29,28] investigated experimen-tally the effect of cooling of stationary outer parts of a two-bearingrigid assembly on preload experimentally. In 1974 Sud and Davies[153] investigated the effect of speed and the similarity of thermalpreload with displacement of mechanical systems in time. Theysuggested a first order differential equation and a method tocalculate its constant terms. These terms are based on thermalcapacity, conductivity and convection as well as geometry andthermal expansion coefficients of a particular assembly. In 1983Lacey et al. [92] conducted extensive experimental research onoperational preload of rigid spindles. They investigated the effectof speed, lubrication and initial preload and indicated various typesof behavior based on the most important parameter, speed.

3.3.5. Current models to predict thermal effects

There have been several groups of authors investigating andpublishing results on the thermo-mechanical behavior of high-speed spindle units:

1. S
tein, Bossmanns, Lin, Tu, Harder: In 1994 they presented apredictive model of a simple two-bearing assembly [149]. Themechanical part of the model was based on simplifiedcalculation of relative distance of bearing rings. The paperwas followed by more detailed investigation of heat transfer inspindles [19] and its interface with bearing ring model [99].
2. J
orgensen, Li, Shin determine the steady state heat transfer bygradually introduced boundary conditions [77,96,97]. They usedeMul’s bearing model [33] for calculating condition andproperties of bearings.
3. K
im-Lee focused on the effect of radial bearing fit [82,83]. 4. K owal, Jedrzejewski, Kwasny, Winiarski: program based on FEM
and FDM [71,88,89,73]. The papers showed the results withcomments, but the adopted bearing models were not presented.

5. Z
verev, Eun, Chun, Lee: their model published in [180] and [181]has the same approach to the mechanical part of the problem asin [27], but they added a simplified heat transfer and thermalexpansion model based on beam elements.
6. H
olkup, Holy: model [63] and [62] used Jones’ bearing model,axis-symmetric heat transfer in 2D finite elements, effect ofbearing and spacer radial fit, transient heat transfer (see Fig. 15).
The common structure of the past models can be listed as:

� R
olling bearings are modeled with Jones’ or deMul’s bearingformulations (Groups 2, 5, 6). � S teady state and transient analysis which also captured preload
peaks are considered by all groups except (2).

[(Fig._15)TD$FIG]

Fig. 15. Example of heat transfer model in mounted bearing [62].


� G

[(Fig._16)TD$FIG]

[(Fig._17)TD$FIG]

Fihi

ood model of heat transfer and thermal expansion based onFEM with 2D or 3D elements – all groups except (5).
� T he nonlinear effect of bearing and spacer radial fit, radial
flexibility of supporting structure (rings, housing, shaft) influen-cing heat transfer between parts as well as mechanical stiffness(3, 6).

In principal, the structure of a predictive thermo-mechanicalmodel must be integrated to simulation which updates mechanicalas well as thermal nonlinearities during transient simulation [63]as shown in Fig. 16.

Boundary conditions of thermal models are crucial for thecorrect prediction of thermal loads on the spindle. Since no generaltheory has been proven to be applicable for the estimation ofbearing heat sources, they still need to be identified experimen-tally as performed by groups 1, 2 and 6.

The displacement distribution in the spindle assembly causedby the temperature fields in the system is also studied in high-speed machining centers as shown in Fig. 17 [75,72].

3.3.6. Summary on thermal aspects

� T
hermal issues significantly affect properties (stiffness, life,accuracy) of spindles.
Fig. 16. Closed loop of thermo-mechanical transient analysis.

g. 17. Measurements of displacements in Z-axis over time at different speeds on a

gh speed machining center [72].

� M
odeling, prediction and optimization of spindles with respectto thermal issues has not been studied sufficiently in theliterature. � It is possible to [62] to build a model which performs a closed
loop thermo-mechanical simulation using best available theoriesin each field.
� S pecial care must be taken when introducing thermal boundary
conditions of such model (bearing friction and surface-fluid heatconvection).
� T hermal displacements resulting from thermo-mechanical
simulation can be passed to nonlinear ‘bearing models’ of purelymechanical models [27] and provide improved prediction ofstructural dynamic behavior of spindles.

4. Mechanical design

Depending on the machine tool application area, the spindle-bearing systems are subject to a complex array of requirements.For example, in high-speed cutting (HSC) applications onaluminum components, there is a need to combine high speedswith low stiffness [67] whereas heavy-duty machining oftitanium- or nickel-based alloys require that the bearings mustbe able to absorb high cutting forces at low speeds of rotation[167]. Apart from the choice of a suitable type of bearing, optimumdesign of the bearing configuration makes a decisive contributionto the performance and service life of the main spindle [65].

4.1. Bearing solution

The following bearing types are used depending on theapplication requirements as shown in Fig. 18 [65]:

� r
olling bearings, � e lectromagnetic bearings, � a erostatic bearings, � h ydrostatic bearings, � h ydrodynamic bearings.
4.1.1. Ball bearings

At specific speed coefficient (Dmn value) up to a maximum of3.0 � 106 mm/min (e.g. milling operations), main spindles withhigh axial and radial stiffness are generally mounted on rollingbearings. High-precision spindle bearings combine good radialrun-out and stiffness properties with low assembly and main-tenance effort, and a good cost-effectiveness ratio [133,142,50].Modern ball bearings produce less friction losses with easierlubrication flow since they have smaller contact surfaces[157,150]. Increased speed requirements have led to the devel-opment of special types of high speed (HS) and hybrid bearings.The HS bearings have a larger number of smaller balls. Frictionalbehavior improves due to the consequent improvement in contactparameters.[(Fig._18)TD$FIG]

Fig. 18. Comparison of bearing system properties [167].

[(Fig._20)TD$FIG]

Fig. 20. Events leading to failure of a cylindrical roller bearing [23].


Hybrid bearings possess silicon nitride balls (Si3N4) with alower density of 3.16 g/cm3 and a higher modulus of elasticity of320,000 N/mm2 (as compared to the usual roller bearing steel100Cr6 with a density of 7.85 g/cm3 and a modulus of elasticity of210,000 N/mm2). The modulus of elasticity crucially affectsresilience. The higher modulus of elasticity results in increasedstiffness at the same preload for hybrid bearings, but also in higherHertzian stresses and lower acceptable loads. In order to keep theHertzian stresses in hybrid bearings at a level comparable to thatfor steel bearings, preloading of the hybrid bearings is reduced. Theceramic material has excellent tribological properties in combina-tion with steel, resulting in lower friction and reduced wear[133,142,50,16,173,151]. It is also possible to improve perfor-mance by using special high-nitrided bearing steel (HNS). Themuch finer microstructure of this stainless steel, combined with itsgreater toughness as compared to 100Cr6 roller bearing steel,enables a higher load level to be attained [162].

One of the most recent developments tends towards coating therolling surfaces with hard thin films. The coatings are intended toimprove the wear resistance further while reducing the frictioncoefficient of contact surfaces [40,132].

Despite their many advantages, the rolling bearings also haveperformance limitations, partly due to the geometry of the angularball bearing. Radial displacement of the rolling bodies, or radialwidening of the inner ring due to centrifugal forces or thermalexpansion can lead to a relative axial displacement of the rings inthe case of a bearing adjustment with constant preload. In the caseof a rigid bearing configuration it will lead to rising internal bearingloads [157,55]. Any reduction in the bearing stiffness due tochanges in the contact angle of the bearing will reduce the dynamicstiffness of the spindle, which will in turn reduce the chatter freematerial removal rates [1,3]. Development of new bearingconcepts is aimed at countering the disadvantages of conventionalspindle bearings described above. These concepts are based on thenotion that axial and radial displacement of the balls (and also theinner ring) can be prevented by an additional rolling contact in theouter bearing raceway [147,172]. Bearing types with this kind ofinner geometry are being used experimentally, for example in theaerospace sector (aircraft engines). These applications, however,require different lubrication and rolling contact parameters.

Fig. 19 provides an overview of concepts. Apart from the 3-pointbearing with two rolling contacts on the outer ring, two variants ofthe 4-point bearing are shown. In the 3-point bearing, theproblems associated with migration of the balls to the apex asthe speed of rotation rises are prevented by the double contact onthe outer ring. Preloading of the inner ring by appropriate loadingof the bearing continues to be necessary. The rigid 4-point contactcan be used as a solid bearing. At high speeds, however, the loads atthe contact points increase strongly due to elastic and thermalexpansion of the spindle, and limit the allowable spindle speed.The elastically loaded bearings with 4-point contacts do not sufferfrom this drawback. If a change in contact kinematics due tothermal expansion or centrifugal forces occurs the divided halves
[(Fig._19)TD$FIG]
Fig. 19. New kinematic concepts for spindle bearings.

of the inner ring will change their axial distance until the newequilibrium is achieved. This bearing is safeguarded againstoverload by springs. It is also possible to cool the outer ringwithout endangering the bearing through a build-up of bracingforces and bearing heat by mutual feedback (known as the ‘suicideloop’, Fig. 20). The spring force, in all cases, must be larger than themaximum axial spindle load in the direction in which the rings arepulled apart [147].

4.1.2. Roller bearing

Cylindrical and tapered roller bearings with single or multiplearrangements are most commonly used on machine tool spindles.High-precision cylindrical roller bearings have been used particu-larly as movable bearings on spindles, but can likewise be mountedto increase radial stiffness in the region of the spindle nose [102].Cylindrical roller bearings are radially preloaded via adjustment ofbearing clearance in the installed state [133,23]. Due to much greatercontact area between the rolling elements and the races, cylindricalroller bearings are much stiffer than ball bearings, and are able tocarry heavier loads [61,174]. However, the increased contact areaproduces higher friction and is more difficult to lubricate. Ascompared to that of balls in angular ball bearings, they rotate onlyabout one direction, which is kinematically favorable [139]. Criticaloperating conditions occur especially when there are temperaturegradients between the inner and outer rings. The outer ring canusually dissipate heat significantly better via housing componentsthan the inner ring via the spindle body and by convection.Thermally induced radial expansion of the inner ring directlychanges the preload set during mounting. An increased preload inturn increases heat generation in the bearing. When a thresholdpreload value is reached, the bearing is no longer able to dissipate thegenerated heat quickly, especially from the inner ring. The bearing isin a ‘suicide loop’ (Fig. 20) which can result in destruction of thebearing within a few seconds [23,22,163].

A number of different approaches have been adopted toincrease the reliability of cylindrical roller bearings [23,122,118,119,56]. Systematic weakening of rollers, inner and outer ringsmakes the bearing less sensitive to the changes in radial preloads.Smaller roller-ring contact zones, profiled rolling bodies, and theuse of ceramic rollers are used to reduce the friction and hence theexcessive generation of heat in the bearings (Fig. 21). By optimizingsuch design variables, it is possible to increase the operatingspeeds of the bearings prior to reaching the suicide loop. Thereduction in radial stiffness for roller bearings which accompaniesmodification of the bearing components may be regarded asunproblematic from the viewpoint of statics and dynamics in aspindle-bearing system [23] because its initial stiffness is alwayshigher than that for ball bearings.

4.1.3. Magnetic bearings

Spindles operating with electromagnetic bearings cover a widerange of applications at high speeds. Because of the relatively large

[(Fig._21)TD$FIG]

Fig. 21. Constructive increase of bearing compliance and reduction of friction [167].

[(Fig._22)TD$FIG]

Fig. 22. Preload mechanisms and influence of a temperature difference DT [167].

Fig. 23. Arrangements of movable bearing units [167].


air gaps between the rotor and stator of the axial and radialbearings, the friction is negligible, and the bearings can be operatedwith minimal wear [65]. Because magnetic bearings are activelycontrolled, stiffness and damping properties can be adjusted as afunction of the feasible control loop dynamics [155]. Due to theintegral behavior of the controller, the maximum possible staticstiffness of a magnetic bearing is higher than that of a comparablerolling contact bearing. However, the maximum load rating issignificantly lower than that of rolling bearings. Even at highspeeds of rotation, the rotor in the magnetic bearing can be movedto eccentric paths, permitting a certain degree of self-balancing[52]. Due to the high costs of a complex control system andperipherals these spindles are currently used only in specialapplications. In the research field, for example, magnetic bearingsare used as actuator for contactless application of static or dynamicloads to the spindle-bearing system or as additional bearingsmounted with conventional spindle bearings to provide an activeinfluence on the bearing of a motor spindle [81,94].

4.1.4. Fluid bearing

Depending on the desired speed of rotation, spindles withhydrostatic or hydrodynamic bearings are frequently used for highprecision production tasks in the metalworking sector. Bycomparison with rolling contact bearings, and given good dampingand high stiffness, these have the advantage of achieving minimumradial and axial runout. As a result of increased heating of the fluid,due to internal shearing effects, there is, a limit on the speeds withspindles having larger tool interfaces (e.g. HSK 63; approximately10.000 min�1) [65,124]. Recent developments show the use ofwater as the hydrostatic bearing fluid to minimize frictional lossesand optimize rotational speed performance [43]. Compared tohydrostatic bearings, hydrodynamic applications in machine toolconstruction have lost much interest. Hydrodynamic bearings areused only where operation conditions are in the purely fluidfriction range without speed variations. On slow running spindlesor spindles with frequent starts and stops (e.g. tool changes),operation in the mixed friction range causes increased wear andlarge frictional losses [65].

Aerostatic bearings are employed when much higher rotationalspeeds need to be achieved. These work on the same principle asliquid-lubricated bearings, but the active medium is gaseous air,with a lower viscosity than that of the liquids by two to threeorders of magnitude. In order to realize a high load capacity andstiffness, very small clearances must be used within the bearings.The air fed into the system is blown off to the surroundingenvironment by balancing the design of the bearing land forthe correct flow resistance for stiffness and length for loadcarrying capacity. Due to the low mass flows and small specificthermal capacity of air, the frictional heat produced by shear forcescannot be dissipated completely at high relative speeds of thebearing components. High-speed spindles with aerostatic bearings

therefore require additional cooling. The viscosity of the air isvirtually independent of its temperature. At certain pressures, thecompressibility of air causes pneumatic instabilities, meaning thataerostatic bearings with feed pressures between 4 and 10 � 105 Pashould preferably be operated in the laminar flow range. The lowpressures entail relatively low rated loads and stiffness, requiremuch larger dimensions than hydrostatic bearings [15,175].

4.1.5. Floating bearing and preload mechanisms

The properties of a rolling contact bearing and of the completespindle-bearing system depend greatly on the chosen preload forceand preload mechanism [157,65].

In terms of design, the preload can be realized in different ways,a fundamental distinction being drawn between rigid and elasticbehavior. A rigid configuration is the simplest engineering designsolution, since the sole need is to fix the outer and inner ringsaxially, either by direct opposition of the two bearings or by meansof matched spacer rings. Depending on the orientation of thebearing, one differentiates between X and O arrangements. In an X

arrangement, an axial thermal expansion of the shaft is trans-formed into an increase in internal loads, and may lead to failure ofthe bearing. Analogously, in an O arrangement the preload isrelieved (Fig. 23).

Elastic preloads keep the bearing preload constant, even whenthermally induced relative movements take place between the[(Fig._23)TD$FIG]

[(Fig._24)TD$FIG]

Fig. 24. Layout of an oil–air lubricating system for machine tool spindles.

[(Fig._25)TD$FIG]

Fig. 25. Lubrication variants.


spindle and the housing. This can be achieved by using disk springs,by distributing coil springs around the circumference, viahydraulic or pneumatic pistons or via high dynamic piezo-actuators.

When designing an elastic bearing mounting, space must be leftfor the bearing rings to move axially. The simplest solution is theuse of slide bushings, which are inserted in the housing, wherethey contain the outer rings of the spindle bearings. To perform thefunction of a movable bearing, it is necessary to provide aminimum radial clearance, whose size depends on the collectiveload and on the construction of the spindle. If the clearance is lowerthan this minimum size due to thermal expansion of the sleeve, orif fretting corrosion occurs, sliding may become impossible ormoves in a slip-stick fashion, the bearing may be destroyed.Possible solutions are to use non-ferrous metal or to coat the sleeveor the housing bore. The advantages of the bushings are simplicityof design and good heat dissipation from the bearing to thehousing. To make the movable bearing with sliding bushinginsensitive to radial seizing, a hydraulic counter-pressure bushingcan be inserted. Here the housing and bushing are separated fromone another by an oil film. Additional pressure is applied tocompensate thermally or kinematically induced ball counter-forces which could lead to radial seizing. Design effort is increasedas opposed to a simple sliding bushing.

In order to reduce friction between the elements of the movablebearing unit which move in relation to one another, it is alsopossible to employ an axially mobile ball-bearing linear bushing. Itshould be noted, that if the sleeve expands radially due to heatingor centrifugal force, high Hertzian stresses may occur in contactwith the housing, due to the small diameter of the balls. This canmake the balls exceed the allowable stresses and penetrate thesurface of the housing, leading to failure of the linear bearingfunction. Another disadvantage is poor heat dissipation due to thereduced surface contact.

A completely different design solution is to integrate themovable bearing in a membrane spring bushing, consisting of twonested but radially independent sleeves supported in the housingby radially stiff but axially compliant membrane-spring-typeelements. The axial and radial stiffness of the movable bearing canbe influenced via the spring characteristic. A disadvantage of thissolution is the greater difficulty in dissipating heat loss induced bythe bearing [23]. If it is possible to dispense with spindle bearings,the use of cylindrical roller or floating displacement (FD) bearingsprovides a very simple option to realize movable bearings in termsof engineering design. Thanks to the flat outer or inner ring, anaxial compensating movement for thermal expansion of the shaftcan take place in the form of spiral rolling directly in the bearingitself. Disadvantages are the complex mounting procedure andhigh sensitivity to radial seizing [65,23].

4.1.6. Lubrication

The tribology system of a rolling contact bearing is character-ized by a heavily loaded rolling contact and, in the normal case, lowloaded sliding contacts (rolling body/cage, cage/bearing ring).

The main task of the lubricant (interfacial medium) in a rollingcontact bearing is to form a lubricating film in the respectivecontact zones of the rolling bodies, bearing rings and cage, so thatcontact between surfaces – resulting in friction and wear (DIN50322) – is reduced [167]. It also serves to reduce corrosion and todissipate frictional heat [142]. In principle, methods are availablefor lubricating the main spindle of a machine tool, depending onthe range of speeds involved: grease lubrication and greaserelubrication, oil–air lubrication and oil injection lubrication.

Roughly 90% of all rolling contact bearings are operating withgrease lubrication [86,17]. In this form of lubrication, the bearingsare filled with grease prior to mounting, and it is also referred to aslifetime lubrication. Speed coefficients of up to 2.0 � 106 mm/mincan be achieved by optimizing the chemical composition of thegreases and adjusting lubrication to the ceramic ball materialsemployed in modern high-speed spindle bearings [133]. To

improve high-speed performance and service life in greaselubrication systems, relubricating systems, in which fresh greaseis introduced via feed lines to points near the bearings or to thebearings themselves, are also commonly used in addition tolifetime lubrication.

Oil–air lubrication (Fig. 24) can be used for applications withspeed coefficients of up to 3.0 � 106 mm/min. The principle oflubrication is based on continuous dosing of compressed air mixedwith oil to the bearing. Depending on the size of the bearing,quantities of oil lower than 60–200 mm3/h per bearing aresufficient lubrication. Oils with viscosities between 32 and100 mm2/s can be used and the pressure range is between 2and 6 bar, depending on the manufacturer’s recommendations anddesign [133].

In these assemblies, pressure is previously built up in adistribution system by a pump, or static pressure is exerted by thereservoir weight applied to the dosing valves. The valves feed adefined quantity of oil (usually 10 mm3) to a mixing chamber,which is then passed via lubricant feed lines to form a streak of oilin a constantly passing air flow in a narrow tube. This is then fedaxially or radially via a feed nozzle to the bearing. In oil–airlubrication, the volumetric flow rate is determined by the cycletime, which establishes the length of time between individuallubricating pulses or valve switching times. If the selected periodbetween cycles is too long, the streak of oil may be interrupted.Some systems in use currently, use sensors to monitor the oil level,oil pressure and air pressure, together with the dosing valves andoil streak themselves [111,21].

The lubricant feed to the rolling contact differs depending onthe type of bearing concerned. Lubricant can either be provided bya reservoir directly on the bearing itself (capped bearing forlifetime lubrication) or supplied to the vicinity of the bearing byfeed lines. Lubricant is fed axially or radially to the bearing througha hole. In the case of spindle bearings, different types of feed areemployed, varying in their geometries and surrounding compo-nents. Fig. 25 provides an overview of the types in use [133].

4.2. Cooling

4.2.1. Motor cooling

Motor spindles are equipped with high power motors whichproduce a large amount of lost heat. Therefore, in mostapplications the spindle housing is flown through by a liquidcooling medium in a closed cooling circuit (see Fig. 26, top).

[(Fig._26)TD$FIG]

Fig. 26. Motor and bearing cooling [courtesy: GMN].


Hence this cooling method is quite complex due to itsperipherals (cooling unit, circulating pump). Newly developedsolutions like [126] suggest a different consideration with anangular spindle housing. The housings backend is extended andequipped with a separately driven fan. The housing has drains in itscorners where an air flow in axial direction is generated caused bythe fan. An additional cooling effect is achieved by generating anair flow through the air gap of the motor.

4.2.2. Bearing cooling

To minimize the thermal deformation of an aerostatic spindlesystem, the control of supply air temperature based on the conceptof thermal balance has been proposed [179]. The system developedprovides the thermal balance between heat generation in bearingclearance and the cooling effect of supply air, therefore, it is possibleto eliminate the thermal deformation of the overall aerostaticspindle system including the surrounding parts of the system.

4.2.3. Shaft cooling

A newly developed spindle (Fig. 27) is based on an interiorcooled shaft [137]. The central component is a rotary union with atleast three separate connections. Cooling can be applied throughan axial interface and dissipated by a radial hand-over-point. Bythis means a cooling circuit through the shaft is realized. Asubstantial advantage of this system is the noticeably shortenedtime till a steady thermal condition is achieved and through this areduced heat input from the spindle into the tool is reached.Therefore a thermally induced spindle extension is noticeablyreduced [166].

4.3. Internal coolant supply

So-called rotary unions are needed for passing a liquid mediumthrough the rotating shaft to the cutting tool. There are basicallycontacting and non-contact solutions. On the one hand Frisch[46,45] describes a contactless rotary union for use in motorspindles equipped with an air seal. This system is suitable for arotational speed up to 60,000 rpm and media-pressure upto 4 MPa

[(Fig._27)TD$FIG]

Fig. 27. Shaft cooling [courtesy: Fischer AG].

with minimal leakage [46]. He also provides discussions onemerging techniques for rotational speeds up to 90,000 rpm [45].

Sykora [154] points out that many sealing materials utilized innon-contact rotary unions are designed for operating with either acooling lubricant or minimum quantity lubrication (MQL). Chan-ging the type of lubrication supply may damage the seals. In such acase a solution is provided which allows switching between bothcooling systems.

MQL has grown in popularity in recent years in an effort tominimize environmental impact. Different works illustrate thatthe inner supply with pre-mixed aerosol is quite problematic:Schneeweiß et al. [136] as well as Aoyama et al. [13] noticed thatan increasing rotational speed leads to demixing of the aerosol.This is mainly the result of the centripetal force acting on the oildroplets which leads to an oil dispersal at the walls of thelubrication circuit [13]. For analyzing the influence of thecentripetal force Aoyama et al. [13] also concentrated on thedroplets size; the smaller the droplets the lesser the forces.Furthermore the decomposition is less intense. At the same timethe lubricating effect is influenced because the droplets poorlystick to the cutting edge.

Schneeweiß et al. [136] and Palm and Fuchs [114] identifiedvolatile changes in cross-section, dead spaces and leakage asreasons for a significant oil losses inside the spindle. Aoyama et al.[13] invented a solution with a pivot-mounted pipe inside thespindle. While the spindle rotates the pipe stands still. In thismanner the aerosol inside of the pipe is not affected by centripetalforces. A contactless rotary union provides for the transfer of themedium to the rotating tool. Air and oil are led through twoseparated ducts within the spindle and mixed just before the toolas shown in Fig. 28 [32].

4.4. Tool clamping and release mechanism

In most instances the clamping force is provided by axiallystacked disk springs or spiral springs. However these springs are asource of unbalance in operation because the allocation of themasses may vary. One possibility is to use gas-pressurized springsinstead of steel springs as shown in [70]. The benefit of thisvariation is the homogeneous mass distribution. One problem withgas springs is the loss of pressure resulting in a loss of clampingforce. Hence it is necessary to monitor the spring load permanently[70].

A system which realizes the tool clamping with an electricallinear motor instead of a hydraulic system or spring assembly wasintroduced in [90,103]. In addition to shortened tool change timesand increased balance quality of the spindle being achieved, theclamping force can be monitored continuously (see Fig. 29).

4.5. Interfaces

4.5.1. Spindle/tool interfaces

The tool holder is the interface between tool and spindle. Inaddition to high stiffness, adequate damping and the ability oftransmitting the required forces/torques, these systems should[(Fig._28)TD$FIG]

Fig. 28. Avoidance of aerosol demixing by separated ducts for air and oil [32].

[(Fig._29)TD$FIG]

Fig. 29. Electrical clamping system [90,103].


allow a quick, accurate and reliable tool changing process. Verycomprehensive overviews of spindle/tool interfaces are given in aCIRP keynote paper [125] and by Weck and Reinartz [170].

4.5.2. Steep taper, SK taper

The main drawback of conventional steep tapers is the minorcontact area between the tapers of the tool holder and the spindle.Compared to HSK tool holders (HSK = Hohlschaftkegel) there is anoticeably lower rigidity. A commercially available steep taper iscalled ‘Big-Plus’. These tool holders eliminate the mentioneddisadvantages of SK tool holders through different optimizations ofthe taper and the tool clamping. An additional flange contactsurface is added as shown in Fig. 30(top). On the one hand thisleads to high repetition accuracy in tool clamping. On the otherhand a remarkable stiffness increase is achieved compared to usualsteep tapers (see Fig. 30, bottom) [18].

Other references can be found on how the flange contact can berealized with short tapers. One possibility is to shorten the taperand to use a longer retention knob [113,152]. It is also suggested toinsert adjusting washers. This may affect the rigidity of theconnection but it improves the damping [125]. One further optionfor setting flange contact as well as taper contact is to use oversizedand axially slotted tapers. This allows a taper deformation causedby the pull force till the flange surfaces touch [54,57,143].

[(Fig._30)TD$FIG]

Fig. 30. SK-Slot ‘Big Plus’ and rigidity advantages [18].

4.5.3. Hollow shaft taper, HSK

For many years the HSK interface can be considered as standardin HSC milling. The static and dynamic characteristics of the HSKtool interface have been analyzed by many institutions andresearchers. The basic performance of the HSK shank, such aspositioning accuracy, stiffness and bending load capacity, wereanalyzed in Germany by WZL [171]. Later Aoyama and Inasaki [12]showed that in worst case of diameter tolerances the taper surfaceof a HSK A 50 tool holder at spindle speeds over 24,000 rpm canlose contact and so the radial stiffness decreases. It is proposed toincrease the taper oversize as well as the pull-in-force [12]. Hannaet al. [53] also observed that the deviation in dimension of thetapers is of great importance for the characteristics of the HSKinterface. They also show that an increased axial pull forcepositively affects the rigidity of the system as well as the ability totransfer the required torque.

The authors in [70] have investigated theoretical FE analysisand measurements to show the dependencies of the mechanicalstresses due to the rotating speed. Strain gauges were applied tothe clamping assembly and analyzed via a contactless telemetricsystem. Fig. 31 shows results that stress increases with higherspeeds and demonstrates an increased pull-in-force.

4.5.4. Coromant Capto

Sandvik Coromant company developed a system called‘Coromant Capto’, which was originally designed for the modularconstruction of very long tools. The system is mainly characterizedby a very flat taper (cone pitch 1:20) with a polygonal shaped outercontour [130]. The hollow taper is deformed during the drawinginto the spindle until the flange surface of the tool holder contactsthe spindle. Due to its good symmetrical characteristics, the highstiffness and the high torque transfer, this system is also adequateas an interface between spindle and tool holder [125].

4.5.5. Spindle/machine tool interfaces

The development of reconfigurable machine tools needs newconcepts for spindle/machine tool interfaces. Abele et al. [6]compiled different interface requirements for reconfigurablemachine tools and suggested an interface which is based on amodular concept where – according to stiffness demands – up to 9coupling mechanisms could be used. With this interface a changeof different spindle modules is possible within less than 30 min(Fig. 32).

Further requirements for spindle-machine-interfaces are pre-sented by Heisel and Meitzner [58]: All ports (for energy orinformation) should be included into the interface. All componentsshould offer a high stiffness and damping.

Heisel and Tonn developed a new model of a lathe main spindle[59]. The interfaces inside the spindle were adjusted so that eventhe machine operator himself can perform the exchange and
[(Fig._31)TD$FIG]
Fig. 31. Measured mechanical stress via strain gauges at the clamping set of a HSK

100 A [70].

[(Fig._32)TD$FIG]

Fig. 32. SST60 interface (according to [6]).


alignment. For this purpose a self-centering device with appro-priate profile in the area of the rear bearing was built (Fig. 33, top).In the newly developed spindle module a conventional counterspindle was combined with an adjusting device (Fig. 33, bottom).This allows alignment of the two spindles. This method allowedthe exchange of a spindle with the eccentric spindle module inunder two hours.

4.6. Light weight design

For the utilization in small and fast moving applications likeprinted circuit board applications, Ban and Lee [14] developed anaerostatic spindle with a shaft consisting of a carbon fiber
[(Fig._33)TD$FIG]
Fig. 33. Installation of the plug-in motor spindle into the headstock and developed

counter spindle module with double eccentric units:(source: IfW, University of

Stuttgart/INDEX-Werke Esslingen).

composite material. In addition to the analysis of the mechanicalcharacteristics the optimization of fiber layers is considered. Formounting a tool at the one end of the spindle-shaft and an electricmotor on the other end steel flanges are glued on both ends of thespindle.

5. Drive concepts

5.1. Motor design

The present multi-functional machine tools demand motorspindles suitable for HSC-applications with high rotational speedsand lower torques as well as spindles for heavy roughing withcomparatively low rotational speeds and high torque. Theserequirements are fulfilled with various solutions. It was suggestedto connect the shaft of the motor spindle through a shiftablecoupling to an additional electric motor as shown in Fig. 34. Whilethe main drive works at high rotational speeds and rather lowtorques, the second electric motor is switched on at low rotationalspeeds and high torque. This construction offered differentadvantages, the functional range (torque, speed) of the spindleenlarges and a reduction of torsional vibrations of the spindle andthe tool is achieved [146].

5.2. Frequency converter

Frequency converters are required to convert the constantthree-phase supply into variable three-phase supply. Integralmotor spindles are usually fed by a three-phase two level inverterwith pulse width modulated (PWM) output. A wide field ofproblems emerges from the inadequate supply through theinverter. Because of the switching operating mode, the outputvoltages are not purely sinusoidal and contain switching harmo-nics. These harmonic voltages induce currents which do notcontribute to torque formation but solely to an undesirable heatingof the various spindle elements. To reduce the harmonic content inthe supply voltages and currents three-phase three-level inverterswere introduced [93]. Due to additional power semi-conductorsthese inverters have the ability to apply an additional voltagepotential to the motor. This significantly reduces the harmoniccontent of the supply voltage. Fig. 35 shows the set up of the twodifferent inverters and the achievable voltages and currents [127].

Another possibility to overcome the mentioned problemswhich arise from inadequate voltage supply is to apply a LCoutput filter between frequency converter and motor. Thesesecond order filters damp the harmonic content in the supply. Thisresults in voltage and current which are very close to the idealsinusoidal form. In this filter application one must pay attention tothe characteristic resonance frequency of the filter is not excited byeither the inverter or by the motor, as this could destroy theinverter, the filter or the motor. Hence, a control system for the
[(Fig._34)TD$FIG]
Fig. 34. Motor spindle with auxiliary drive [146].

[(Fig._35)TD$FIG]

Fig. 35. Setup, phase-to-phase voltages, phase currents of two-level (top) and three

level inverters (bottom) [127].


filter voltages and currents is introduced to make sure that theresonance frequency of the filter is sufficiently damped [4,5].

The influence of the inverter design concept on the temperaturebehavior of a single integral motor spindle with a permanentmagnet synchronous motor is depicted in Fig. 36. The highesttemperatures are achieved with the two-level inverter while thefilter control achieves lower temperatures.

6. Sensor integration/condition monitoring

6.1. Vibrations/chatter detection

6.1.1. Chatter detection during the milling process

In conjunction with the detection of chatter vibrations severalmethods have been developed over the years to improve thestability of milling. A comprehensive overview is given in [10,141].The use of microphones and accelerometers are successfullyapplied in detection chatter in milling operations [34,35,107].

6.1.2. Dynamometers

For measuring the cutting force during the milling process twovariations of dynamometers are often used: Plate dynamometers,which are placed under the workpiece (used by [8,34,145,31,131,134]), and rotating dynamometers, which are locatedbetween the tool clamping and the milling spindle [134,91].

First the sensor signals are analyzed for determining thecharacteristics of the sensor signals in stable and unstable cases.Secondly chatter indicators and the critical values are composedand finally the characteristics are compared.

6.1.3. Accelerometers

According to [31,34,91] the acceleration based chatter detectionis carried out with accelerometers. The characteristics of accelera-tion signals are similar to cutting force signals: The signals areperiodic in stable and non-periodic in unstable machining processes.Choi and Shin [31] use the measured signals for estimating the[(Fig._36)TD$FIG]

Fig. 36. Stationary temperatures of motor and front bearing with three different

frequency inverters (according to [127]).

chatter detection index g. This index gives information about thedetection accuracy and the permissible computational efficiency;hence it is appropriate for online implementation. Suitable thresh-old values of g for turning and milling operations are presented.Kuljanic et al. [91] tested several sensors like rotating dynam-ometers, accelerometers, acoustic emission and electrical powersensors. The sensors were compared in terms of accuracy androbustness. The best results were achieved with a multisensorysystem composed of axial force sensor and accelerometers. Theresearch of [34] concluded that microphones, which are placed inthe ambience of the milling machine to record the ideal noiseemission, achieved the best results compared to other sensors.

6.1.4. Displacement meter

Another method for measuring and identification of chatterduring milling is described in [129] where a laser displacementmeter is used in a prototype of a milling system. A laser beam isaimed at the cutting edge and reflected. The intensity and the angleof incidence of the reflected laser beam are recorded and used tointerpret the occurring vibrations and to monitor the toolgeometry during the milling process.

6.2. Spindle integrated force measurement sensor system

Another method of chatter detection is presented in [116],which is based on piezo-electric force measurement sensors,which are integrated into a spindle.

6.2.1. Chatter detection during the grinding process

There are several methods for process monitoring to identifychatter during the grinding process. Laser triangulation sensors,pneumatic, radar and waviness sensors are used to measure thegeometry of the grinding disk, its wear and the surface [68].Another possibility for monitoring and diagnosing machiningprocesses is to record the acoustic emissions during the grindingprocess. A sensor-integrated grinding disk and an acousticemission (AE) sensor are used for this purpose in [160]. Thedifferent AE sensor integrated in the control loop of grindingmachines, are shown in Fig. 37.

The two automatic identification methods of chatter, which arepresented in [51], have the following indicators: Entropy andcoarse-grained information rate (CIR). Signals from piezo-electricand acoustic emission sensors, which capture the normal grindingforce and the acoustic emissions, are stored for further analyses.During this sequence the entropy is counted from the servicespectrum. The CIR is obtained directly from the fluctuation of therecorded signal.

6.3. Preventive detection of bearing damages

One of the main problems which suspend the production line isthe early failure of the spindle bearings. Characteristically this[(Fig._37)TD$FIG]

Fig. 37. Sensor concepts for acoustic emission [160]].

[(Fig._39)TD$FIG]

Fig. 39. Force measuring ring based on piezo-electric force sensors [134].[(Fig._40)TD$FIG]


involves high costs for repairs and downtime. Detection andprevention of bearing damage can be taken as precaution.

A simple and low cost method for predicting the right time forthe replacement of the bearing is to monitor the geometricaldamage occurring at the rolling surfaces of the ball bearings [66].Sinking into and out of a cavity creates an acceleration signal whichis analogous to a pulsed sine wave shown in Fig. 38(top).

The vibration signals emitted by the rolling motion of thebearing elements (outer and inner rings, cage and balls) areanalyzed. A piezo-electric accelerometer is attached at the spindlehousing to sense the mentioned acceleration signal. The construc-tion of the test bench is shown in Fig. 38(bottom). After convertingthe signal it is monitored by a failure prediction processor. Thisprocessor transmits a failure prediction alarm when a referencevalue is exceeded.

With measuring the temperature of the outer bearing raceduring the acceleration of the spindle Spur and Feil [148] assess theactual status and the aging of the bearing as well as thecontamination of the lubrication. Damage of the bearing runningsurface can also be estimated.

For detection of axial forces in the fixed bearing a specific forcemeasuring ring was developed by [24,84,134]. This force ringconsists of a distance tube equipped with piezo-electric forcesensor elements which are circularly adjusted at its front (Fig. 39).This device delivers information about the axial cutting force, thedynamic bearing forces during the operation and the actualpreload force during the assembly process. Excessive load on thebearings which causes destruction can be detected. The arrange-ment can be completed by using a flange sensor ring for measuringthe forces in three directions.

The project ‘Intelligent Spindle Unit (ISPI)’ [169] links sensorsfor bearing cooling, stator temperature, spindle rotational speed[(Fig._38)TD$FIG]

Fig. 38. Top: Acceleration signal (R: amplitude, L: time duration, P: time period).

Bottom: Bearing test bench with variable axial load (according to [66]).

Fig. 40. ‘Intelligent Spindle Unit’ ISPI with exemplary sensors and actuators.

and outer load conditions. A CAD-model of the spindle is shown inFig. 40. A microcontroller is used to evaluate the sensor data.

Another application is presented in [109]. Sensors for detectingbearing temperature, vibrations and axial displacement of thespindle shaft and tool change control are included.

6.4. Collisions and tool fractures

Collisions, tool fractures or overload at high-speed applicationscan cause serious damage in today’s high-speed machining centerswith high feed rates.

Spindle integrated force sensing rings can be used as shown in[134,4,85,79] for monitoring the machining process. In [85] apiezo-electric force ring is integrated into the spindle housing asshown in Fig. 41 to capture the force signals with minimum time
[(Fig._41)TD$FIG]
Fig. 41. Installation of the force ring within the spindle [85].

[(Fig._43)TD$FIG]

Fig. 43. Test bed for evaluation of the thermal model [30].


lag. Typically, the structural dynamic modes of the spindle reducethe measurement bandwidth of the force rings. Altintas et al.developed a Kalman filter to compensate the structural modes, andincreased the measurement bandwidth of the spindle integratedsensors significantly [117,9]. As a result, the dynamometers andspindle integrated force sensors can be used to measure millingforces at high rotational speeds.

For the detection of a broken tool an acoustic emission sensor isadded to the machine table. A peak in the amplitude of therecorded signal shows the breakage of the tool. In this context aforce ring is compared with a table dynamometer in [79]. The ringis composed of eight uniaxial piezo-electric force sensorscontained in a circular frame using a special epoxy-material tomeasure the axial and radial force components. The result is thatthe force ring is as good as a table dynamometer. Its use isrecommended if a table dynamometer is not applicable due to theworkpiece size or its geometry.

In [24] two piezo-electric force sensing rings (bearing sensorring and flange sensor ring) are developed and integrated into adirect driven motor spindle for the online process monitoring ofmachining processes. With this method it is possible to detect theprocess forces during drilling with tool diameter of less than 4 mm.

A spindle integrated data logging unit can be utilized to identifyand to avoid spindle damage quickly [38]. A thumb-sized datalogger is permanently integrated into the spindle. The unit recordsthe signals of acceleration, temperature and tool change sensors.Connecting the device through a serial port to a computer allowsthe reading and the parameterization of the logging unit. Thissimple and safe measuring device enables the detection of spindleirregularities for avoiding serious damages. Operational hours arerecorded in relation to the rotation speed which is useful for theprediction of the maintenance interval.

6.5. Axial displacement

The spindle’s axial displacement consists of a speed dependentaxial displacement of the bearing races and thermal expansion ofthe spindle. Fig. 17 illustrates these concepts. An overview ofmeasuring axial displacements is given in Fig. 42.

In [30] Chen and Hsu characterized the thermal growth of a HSCspindle. Compared to a conventional spindle, the HSC spindles aresubject to complex dynamic and speed-dependent influences. Athermal error model is developed to show the mechanical growthand the scheme of Fig. 43 is used. Six sensors are attached to thehousing near the bearings and the cooling, and to detect thethermal expansion of the cutting tool and the front-end-cover.

Another procedure for improving the machine accuracy bymeans of temperature control is described in [138], wheretemperature and flow rate of the cooling are monitored andcontrolled simultaneously.
[(Fig._42)TD$FIG]
Fig. 42. Three methods for spindle’s axial displacement compensation.

Commercially available products are introduced in [44,109]. In[109] a sensor for measuring the axial displacement of the shaft(called ‘AMS’, Step-Tec AG) is added to the front part of the spindle.The CNC-unit enables the compensation of this axial extension. Fordetermining the thermal growth of a spindle it is common practiceto measure the bearing temperature and to perform the neededcorrelations and corrections. The thermal expansion is just part ofthe problem. For operations with high-precision spindles, it is alsonecessary to include the shift caused by speed dependent motionof the bearing to predict more precisely the displacement. Thedisplacement measurement device DMD, developed by Fischer[44], allows for measuring the displacement of the spindle relativeto the housing. Attaching a special sensor into the tool flangeenables detection of displacement within 1 mm accuracy. Themeasured deviation is subsequently automatically transmitted tothe CNC.

7. Mechatronic concepts

7.1. Active balancing

High-speed spindles used for grinding or milling demand ahigher degree of balance of the spindle system, especially, toolholders and tools. The higher quality and precision can be observedin workpieces and a longer life of the machines components iscommon. For achieving these goals it is necessary to lower thevibrations caused by unbalance.

In addition to the ‘classical’ method of balancing by removing oradding mass in two planes, spindle integrated balancing systemshave been developed [110,106]. For reducing vibrations during theoperation of the spindle, an active balancing program usinginfluence coefficient method is used to calculate the optimalposition of the correction masses and an active balancing devicecan be used [106]. Fig. 44(bottom) shows the schematic of thedevice. The active balancing program controls the device bymeasuring vibration magnitude, phase angle, rotation speed, theposition of the balancing rotor and calculates the adjustments toset the correction masses appropriately for balancing the spindleduring the operation. Without power the balancing rotor retains itsposition and circulates with the rotation axis. The pole platerotates along with the rotation axis and has a magnetized surface.It serves as a pathway for the magnetic flux which is created by thecoil flow. For moving the attached unbalance mass the stator issupplied with a current. The operating principle of the device isshown in Fig. 44(top). Exciting current on the driver coil increasesthe upper-side magnetic flux and decreases the lower-sides flux aswell. The rotor is impacted by the downward force normal to themagnetic flux (a). Positioning the permanent magnet at center ofthe pole plate leads to the greatest density of the magnetic flux andthe smallest magnetic resistance (b). Removing the currenttemporarily brings the rotor into the next step due to the inertia(c).

[(Fig._44)TD$FIG]

Fig. 44. Top: Principle of the rotor mechanism for active balancing. Bottom:

Schematic representation of the active balancing device [106].

[(Fig._45)TD$FIG]

Fig. 45. Top: Dual plane balancing system (a: balancer ring at the spindle, b: sensors

at the stator). Bottom: Schematic representation (a: balancer ring, b: stator with

sensors for speed and position, c and d: vibration sensor) [110].

[(Fig._46)TD$FIG]

Fig. 46. Top: Design of the actuator module. Bottom: Test spindle (schematic view)

[36].


The dual plane balancing system depicted in Fig. 45 is used fordynamic balancing which is necessary if long tools are used [110].As opposed to the single balancing plane method, this systemmeasured the vibrations in a front and a rear plane at the spindlefor balancing, hence the vibrations can be reduced moreeffectively. This active balancing system can be used for balancingspindles at run time within a few seconds.

7.2. Active preload control

The preload has a wide influence on bearing life but also on thedynamic runout and thus also impacts surface finish quality. Theobjective of some research works was to define an ideal preload forthe spindle with an active preload-control, and to drive the systemnear an optimal working point.

Prestressing spindles through passive mechanisms is the mostconventional method to preload bearings. The issue is that thesespindles cannot be adapted to changing conditions. To solve thisproblem a test spindle was developed in [36] using a novel piezo-electric based actuator module. By using a spindle with an activepreload-control as shown in Fig. 46(bottom) the optimal preloadfor different operating conditions can be determined. A controlledmechanism for preload adjustment has been developed to achievethe preload values suggested by the manufacturer. It consists of apiezo-electric based actuator module with integrated miniaturizedhydraulic transmission as shown in Fig. 46(top). In combinationwith force sensors a continuously controlled preload shall beobtained [36,159,158].

The actuator pushes against a membrane piston whichdisplaces a small part of the hydraulic fluid. The fluid pressureis exerted on metal bellow piston which moves the push rod [36].Temperature sensors are also integrated for indirectly detectingthe heat generation in the bearings. To compensate the thermaldeflections of the preloaded spindle the actuator module generatesa correlated force.

Within the joint research project ‘ISPI’ (see Fig. 40) a piezo-actuator is used to achieve a constant preload force. Thismechatronic concept allows adjusting the bearing preload to

[(Fig._47)TD$FIG]

Fig. 47. Speed dependant preload force with and without active preload control.

[(Fig._48)TD$FIG]

Fig. 48. Active tool deflection compensation. Top: Prototype. Bottom: Tool

deflection (exaggerated demonstration) [37].

[(Fig._49)TD$FIG]

Fig. 49. AdHyMo spindle [81].


establish an optimal working point if the spindle is overloaded[169]. Fig. 47 shows the measured preload force at differentrotational speeds with and without active preload control.

In [164] a preload model is developed to describe how externalcooling (or heating) affects the bearing preload. Based on themodel, an active thermal preload regulation scheme is proposedand its feasibility is validated experimentally. The proposedpreload regulation scheme is achieved by circulating a cooling(or heating) flow around the spindle housing to manipulate thehousing and the outer ring temperatures.

7.3. Active tool deflection compensation

Increasing demands on productivity and consequently onincreasing feed rates can cause undesirable tool deflections.Machining operations with long slender tools are especiallyaffected.

A mechanism for a multiaxial positioning of a HSC spindle andto correct tool deflections is invented in [176]. The combination ofa parallel kinematic basic drive arrangement and piezo-ceramicactuators was developed. Complementary precision positioningand additional movements like tilting the tool for deflectioncompensation are possible benefits of this concept.

Denkena et al. [37] also use piezo-actuators as active elementsfor tool deflection compensation. The experimental setup consistsof an adaptronic spindle-system which operates with three pre-stressed piezo-actuators. The schematic setup is displayed inFig. 48.

Process forces in the x- and y-direction cause the tooldeflection (Fig. 48, bottom). The static part of these forces isevaluated by averaging the process forces measured with adynamometer. Combined with the stiffness of the tool, thedeflection can be determined. Using the inverse kinematics therequired position is converted into the needed actuator positionsto achieve the desired tool tip position. The actual position issteadily measured and corrected by the joint control of the piezo-actuators.

7.4. Vibration damping/chatter control

The occurrence of chatter is the consequence of an unstablecutting operation. Chatter in machine tools can lead to poor surfacefinish, high loads and damage to spindles, tools and workpieces.For suppressing these unwanted vibrations three basic approachesare introduced: Active, semi-active and passive control of theprocess.

7.4.1. Active compensation

Within the research project ‘AdHyMo’ Abele et al. [81]developed a hybrid bearing motor spindle (Fig. 49). An activemagnetic bearing (AMB) is integrated in a HSC spindle in additionto the conventional ball bearings. The AMB is controlled to increasethe damping of the spindle. An increase of 50% in MRR was reachedin a laboratory set up. Robust feedback and adaptive feed forwardcontrol using m-synthesis improves the process stability. Based on

the research, Abele et al. [2] developed a model for theidentification of the systems dynamics during machining.

Ries et al. [124] introduced a prototype of an active millingspindle. They integrated additional sensors and piezo-ceramicstack actuators for the induction of forces into a common millingspindle. Two piezo-ceramic stack actuators working perpendi-cular to each other are attached to the outer bearing ring of thefront bearing. With this construction it is possible to apply radialloads in the range of 1 kN. The actuators are powered by two high-voltage power amplifiers. A modification of the front bearingallows the generated radial movement. The prototype and themachine tool are shown in Fig. 50. An appropriate control schemeis used for driving the actuators in a way that additional dampingis provided.

A similar work was presented previously in 1998 by Shankaret al. [140] discussing a ‘Smart Spindle Unit’ for active chatter

[(Fig._50)TD$FIG]

Fig. 50. CAD model and prototype of an active milling spindle with piezo stacks.

[(Fig._52)TD$FIG]

Fig. 52. Designed damper [98].


suppression. This system utilized a state-space model for activecontrol and demonstrated 100% increase in axial depth of cut.

The authors in [39] also use piezo actuators to controlvibrations during milling. In contrast to the concept of Rieset al. [124] the feedback for the control is the measured strain atthe tool root as shown in Fig. 51

Some research was done in the field of spindle speed regulationfor stabilization of the cutting process [101,144]. The dynamicinteraction of a spindle-tool set and a thin-walled workpiece wasanalyzed by a finite element approach for the purpose of stability[(Fig._51)TD$FIG]

Fig. 51. Cross-section view of the Smart Spindle Unit [39].

prediction by [101]. The proposed approach indicates that spindlespeed regulation is a necessary constraint to guarantee optimumstability during machining of thin-walled structures. The theore-tical basics for the elimination of chatter in milling through theautomatic regulation of the spindle speed are presented in [144].The system described here does not require knowledge of thesystem dynamics, and it selects stable speeds where no chatter willoccur based on current dynamics.

7.4.2. Semi-active compensation

A semi-active method utilizing an intelligent material: Electro-rheological (ER) fluid is explored in [98]. The medium is non-conducting oil which contains dielectric particles. The fluid canimmediately convert its phase from liquid to solid upon exposureto an electric field. A compact damper which contains the ER fluidwas designed (as shown in Fig. 52) and mathematical models weredeveloped. Furthermore a semi-active artificial intelligence (AI)feedback controller was established.

7.4.3. Passive compensation

Placing the rolling bearings in an additional non-rotatinghydrostatic configuration is the procedure used in [168]. This leadsto an optimization of the damping characteristics of the spindle,which improves the dynamic behavior of the spindle and toolsystem.

8. Conclusions and further potentials

The heart of every modern machine tool is the main spindle unitwhich is often designed as a motorized spindle. This corecomponent contributes heavily to productivity, precision andquality of the machined products. Numerous projects have beencarried out and reviewed here which had the goal to increase theperformance, productivity and reliability of spindles.

Through these numerous research activities the topics mechan-ical/thermal modeling, bearing and drives have been summarized.Also many concepts were developed to integrate additionalsensors and actuators in the spindle unit. In general, it appearsthe state-of-the-art is advancing in the recent past. Especiallyapparent advances in modeling the dynamic behavior whichindicates the potential to more accurately predict the real world.


Further progress in machine tools in future will be directlyrelated to the spindle technology. The market demands are stillpulling many requirements:

� E
nhancing torque and speed for multi-functional technologyapplications (grinding, milling and drilling in one spindle) � M inimization of energy consumption including the peripheral
equipment for drive, bearing and cooling
� H ard to cut materials are demanding better solutions, e.g. cutting
titanium with gearless spindles.

Derived from these demands are the challenges for the futureresearch. Exemplary challenges in rolling element bearings are torealize Dmn values of 3 million mm/min with lifetime greaselubrication. Also system design has to cover spindle units whichimplement multi technologies resulting in high torque and highspeed in one system. Further development will be required toallow sensor actuator integration to use the spindle unit as aninherent quality insuring system. The prediction of thermalexpansion of spindles and associated changes in the structuraldynamic behavior during high-speed machining operations haveyet to be solved with satisfactory accuracy.

Acknowledgements

The authors would like to thank all CIRP colleagues and expertsfrom industry who have given input to this keynote paper. Aspecial thank to following colleagues who have contributed withdetailed information about their work related to spindle technol-ogy: Budak E., Cao H., Denkena B., Heisel U., Holkup T., JedrzejewskiJ., Kolar P., Neugebauer R., Shinno H., Uriarte L., Wertheim R.,Winfough W.R., Yamazaki K. This paper would not have beenrealized without the dedicated effort of Dipl.-Ing. A. Schiffler andDipl.-Ing. S. Rothenbucher.

References

[1] Abele E, Fiedler U (2004) Creating Stability Lobe Diagrams during Milling.CIRP Annals - Manufacturing Technology 53(1):309–312.

[2] Abele E, Schiffler A, Rothenbucher S (2007) System Identification DuringMilling via Active Magnetic Bearing. Production Engineering 1(2):219–226.

[3] Abele E, Kreis M, Roth M (2006) Non-Contact Identification of Spindle-ToolFrequency Response Functions. Production Engineering 13(2):207–210.

[4] Abele E, Rothenbucher S, Schiffler A (2007) Digitally Controlled Output Filtersfor Improvement of Temperature Behavior of Machine Tool Main SpindleDrives. Production Engineering 1(2007):219–226.

[5] Abele E, Rothenbucher S, Schiffler A (2009) Ausgangsfilter fur Synchron-Hauptspindelantriebe - Theoretische Grundlagen fur ein neuartiges Rege-lungskonzept. wt Werkstattstechnik online 99/5:324–330.

[6] Abele E, Worn A, Fleischer J, Wieser J, Martin P, Klopper R (2007) MechanicalModule Interfaces for Reconfigurable Machine Tools. Production Engineering1:421–428.

[7] Altintas Y, Cao Y (2005) Virtual Design and Optimization of Machine ToolSpindles. CIRP Annals Manufacturing Technology 54(1):379–382.

[8] Altintas Y, Chan PK (1992) In-process Detection and Suppression of Chatter inMilling. International Journal of Machine Tools and Manufacture 32(3):329–347.

[9] Altintas Y, Park S (2004) Dynamic Compensation of Spindle-Integrated ForceSensors. CIRP Annals – Manufacturing Technology 53(1):305–308.

[10] Altintas Y, Weck M (2004) Chatter Stability of Metal Cutting and Grinding.CIRP Annals Manufacturing Technology 53(2):619–642.

[11] Antoine JF, Abba G, Molinari A (2006) A New Proposal for Explicit AngleCalculation in Angular Contact Ball Bearing. Journal of Mechanical Design128(2):468–478.

[12] Aoyama T, Inasaki I (2001) Performance of HSK Tool Interfaces under HighRotational Speed. CIRP Annals Manufacturing Technology 50(1):281–284.

[13] Aoyama T, et al, (2008) Development of a New Lean Lubrication System forNear Dry Machining Process. CIRP Annals Manufacturing Technology57(1):125–128.

[14] Ban KG, Lee DG (2002) Design of Carbon Fiber Composite Shafts for HighSpeed Air Spindles. Composite Structures 55(2):247–259.

[15] Bartz W (1993) Luftlagerung Grundlagen und Anwendungen. Expert, Ehningen.[16] Bayer O (1997) Keramik im Walzlager – ein wirtschaftlicher Weg. Tagungs-

band ‘‘Innovationen fur Gleitlager, Walzlager. Dichtungen und Fuhrungen’’(Reihe: VDI Berichte) 1331:61-69.

[17] Berger T (2004) Ermittlung der theoretischen Fettgebrauchsdauer fur Lang-zeit- oder lebensdauergeschmierte Walzlager. Konstruktion 5:64–66.

[18] Big-Daishowa. Catalog No. EXm48-3. http://www.big-daishowa.com/E_products/pdf/bigplus.pdf.

[19] Bossmanns B, Tu JF (2001) A Power Flow Model for High Speed MotorizedSpindles—Heat Generation Characterization. Journal of Manufacturing Scienceand Engineering 123(3):494–505.

[20] Brecher C, Esser M, Queins M, Voll H (2005) Motorspindeln sind Schlusselk-omponenten. WB - Werkstatt und Betrieb 138(5):45–49.

[21] Bruenner C, Spieß G (2005) Uberwachte Schmierung, Hochgeschwindigkeits-spindeln mit integriertem Ol-Luft-Schmiersystem. Antriebstechnik 9:340.

[22] Burton RA, Staph HE (1967) Thermally Activated Seizure of Angular ContactBearings. ASLE Transactions 10:408–417.

[23] Butz F (2007) Losungen zur Gestaltung der Loslagerung von Werkzeugmaschi-nenspindeln. Dissertation, RWTH Aachen.

[24] Byrne G, O’Donnell G (2007) An Integrated Force Sensor Solution for ProcessMonitoring of Drilling Operations. CIRP Annals Manufacturing Technology56(1):89–92.

[25] Cao Y (2006) Modeling of High-speed Machine-tool Spindle System. Ph.D. thesis,The University of British Columbia.

[26] Cao Y, Altintas Y (2007) Modeling of Spindle-Bearing and Machine ToolSystems for Virtual Simulation of Milling Operations. International Journalof Machine Tools and Manufacture 47(9):1342–1350.

[27] Cao Y, Altintas Y (2004) A General Method for the Modeling of Spindle-Bearing Systems. ASME Journal of Mechanical Design 126:1089–1104.

[28] Carmichael G, Davies P (1970) Factors Which Affect the Transient Behavior ofPreloaded Ball Bearing Assemblies. ASLE Transactions 15:1–7.

[29] Carmichael G, Davies P (1970) Measurement of Thermally Induced Preloadsin Bearings. Strain 6:162–165.

[30] Chen JS, Hsu WY (2003) Characterizations and Models for the ThermalGrowth of a Motorized High Speed Spindle. International Journal of MachineTools and Manufacture 43(11):1163–1170.

[31] Choi TJ, Shin YC (2003) On-line Chatter Detection Using Wavelet-BasedParameter Estimation. Journal of Manufacturing Science and Engineering125:21–28.

[32] CyTec. Internetprasenz. http://www.cytecsystems.de/cytecsystems/german/produkte/cyspeed/cyspeed_kuehlung.htm.

[33] de Mul J, Vree J, Maas D (1989) Equilibrium and Associated Load Distributionin Ball and Roller Bearings Loaded in Five Degrees of Freedom While Neglect-ing Friction-Part 1: General Theory and Application to Ball Bearings. Journal ofTribology 111:142–148.

[34] Delio T, Tlusty J, Smith S (1992) Use of Audio Signals for Chatter Detection andControl. Journal of Engineering for Industry 114(2):146–157.

[35] Delio T (1992) Method of Controlling Chatter in Machine Tool, US Patent5170358, issued 08.12.1992.

[36] Denkena B, Gotz T (2004) Effects of a Preload Control Unit on Machine ToolSpindles. Production Engineering 11(2):175–178.

[37] Denkena B, Mohring HC, Will J (2007) Tool Deflection Compensation with anAdaptronic Milling Spindle. International Conference on Smart MachiningSystems (ICSMS), 13.03-15.03, Gaithersburg, USA, .

[38] Dittel W (2006) Logger in der Spindel - Condition Monitoring gegen Spin-delschaden. Werkstatt und Betrieb 139(9):109.

[39] Dohner JL, Lauffer JP, Hinnerichs TD, Shankar N, Regelbrugge M, Kwan CM, XuR, Winterbauer B, Bridger K (2004) Mitigation of Chatter Instabilities inMilling by Active Structural Control. Journal of Sound and Vibration 269(1–2):197–211.

[40] Erkens G, Rasa T, Muller J, Brecher C, Spachtholz G (2006) Uber dieVorteile gepulster Plasmen zur Abscheidung innovativer tribologischerHochleistungsschichten auf Lagerkomponenten. Galvanotechnik 97:1224–1235.

[41] Erturk A, Budak E, Ozguven H (2007) Selection of Design and OperationalParameters in Spindle-Holder-Tool Assemblies for Maximum Chatter Stabi-lity by Using a New Analytical Model. International Journal of Machine Toolsand Manufacture 47(9):1401–1409.

[42] Erturk A, Ozguven H, Budak E (2006) Analytical Modeling of Spindle-ToolDynamics on Machine Tools Using Timoshenko Beam Model And ReceptanceCoupling for the Prediction of Tool Point FRF. International Journal of MachineTools and Manufacture 46(15):1901–1912.

[43] Fecht N (2005) Spurtstarke Spindeln. Antriebspraxis 4:20–22.[44] Fischer. (2000) Alles fur die Spindel. Werkstatt und Betrieb HSC Sonderteil

133(9):34.[45] Frisch T (2001) Kuhlmitteleinsatz beim HSC-Bearbeiten - Schnellfrequenz-

Motorspindeln mit Drehdurchfuhrung. Schweizer Maschinenmarkt ;(9)21–23.[46] Frisch T (2001) Minimierter Kuhlmitteleinsatz bei Motorspindeln. dima - die

maschine 55(3):10–11.[47] Gagnol V, Bouzgarrou BC, Ray P, Barra C (2007) Stability-Based Spindle

Design Optimization. Journal of Manufacturing Science and Engineering129(2):407–415.

[48] Genta. (1985) Consistent Matrices in Rotor Dynamics. MECCANICA 20:235–248.

[49] Genta. (2005) Dynamics of rotating Systems. Springer.[50] GMN, P.M.I.G.C.K. (2005) Hochprazisionslager. Katalog-Nr. 40000805, Fir-

menschrift.[51] Gradisek J, Baus A, Govekar E, Klocke F, Grabec I (2003) Automatic Chatter

Detection in Grinding. International Journal of Machine Tools and Manufacture43(14):1397–1403.

[52] Großmann K, Reinschke K, Muhl A, Rudolph J, Rudolph H, Loser M, Eckhardt S(2006) Auslegung aktiv magnetisch gelagerter Werkzeugmaschinen-Haupt-spindeln – Entwurfssystem zur Simulation von Spindeln und Zerspanprozess.ZWF – Zeitschrift fur wirtschaftlichen Fabrikbetrieb 101(9):486–491.

[53] Hanna IM, Agapiou JS, Stephenson DA (2002) Modeling the HSK Toolholder-Spindle Interface. Journal of Manufacturing Science and Engineering124(3):734–744.

[54] Hansen P (2001) Solid Contact. Cutting Tool Engineering 53(7).[55] Harris TA (2001) Rolling Bearing Analysis, vol. 4. John Wiley & Sons, New York.

http://www.big-daishowa.com/E_products/pdf/bigplus.pdf

http://www.big-daishowa.com/E_products/pdf/bigplus.pdf

http://www.cytecsystems.de/cytecsystems/german/produkte/cyspeed/cyspeed_kuehlung.htm

http://www.cytecsystems.de/cytecsystems/german/produkte/cyspeed/cyspeed_kuehlung.htm


[56] Hat Y, Qin Y, Baelendra R, Jiang Q (2004) FE Analysis of a Novel Roller Form: aDeep endcavity roller for roller type bearings. Journal of Materials ProcessingTechnology 145:233–241.

[57] Heel H, Effenberger W (1987). Werkzeugkupplung.[58] Heisel U, Meitzner M (2004) Rekonfigurierbare Bearbeitungssysteme. wt -

Werkstattstechnik online 94(10):517–520.[59] Heisel U, Tonn N (2009) Reconfigurable Counter Spindle Module with Inte-

grated Device for Fast Spindle Alignment. Production Engineering 3(3):313–321.

[60] Hernot X, Sartor M, Guillot J (2000) Calculation of the Stiffness Matrix ofAngular Contact Ball Bearings by Using the Analytical Approach. ASME Journalof Mechanical Design 122:83–89.

[61] Hertz H (1882) Uber die Beruhrung fester elastischer Korper. Journal fur reineund angewandte Mathematik 92:156–171.

[62] Holkup T (2007) Komplexni teplotne mechanicky model vysokorychlostnichsestav valivych lozisek (Complex thermo-mechanical modelling of high-speedrolling bearings). Ph.D. thesis, Technical University in Prague.

[63] Holkup T, Holy S (2006) Complex Modelling of Spindle Rolling Bearings.Journal of Machine Engineering? Eficiency Development of ManufacturingMachines 6(3):48–61.

[64] Holkup T, Kolar P (2008) Influence of Thermo-Mechanical Conditions ofRolling Bearings on the Dynamics of Machine Tool Spindles. MATAR PRAHA2008? Proceedings of Part 1, 141–146.

[65] Holz M (2006) European Titanium Market – Current and Future Szenario. ITA22nd Annual Conference & Exhibition, San Diege, 1–3 Oktober 2006/ThyssenKrupp, .

[66] Hoshi T (2006) Damage Monitoring of Ball Bearing. CIRP Annals ManufacturingTechnology 57(1):427–430.

[67] Howes T (1996) The High-Speed Aluminium Milling Challenge. Cutting ToolEngineering 48(7):64–70.

[68] Inasaki I, Karpuschewski B, Lee HS (2001) Grinding Chatter – Origin andSuppression. CIRP Annals Manufacturing Technology 50:515–534.

[69] Incropera FP, Dewitt DP, Bergman TL, Lavine AS (2007) Fundamentals of HeatAnd Mass Transfer. John Willey and sons, New York.

[70] Jakob F, Sykora H (2008) Special Tool Clamping Requirements in the Field ofHigh Speed Spindles. 7th International Conference on High Speed Machining,Darmstadt, 28–29 Mai 2008, .

[71] Jedrzejewski J, Kowal Z, Kwasny W, Modrzycki W (2004) Hybrid Model ofHigh Speed Machining Centre Headstock. CIRP Annals Manufacturing Tech-nology 53(1):285–288.

[72] Jedrzejewski J, Kwasny W, Modrzycki W (2008) Limiting Precision Distor-tions In Spindle Unit of HSC Machining Centre. Journal of Machine Engineering8(1):44–53.

[73] Jedrzejewski J, Kwasny W, Modrzycki W (2008) Precise model of HSCMachining Centre For Aerospace Parts Machining. Journal of Machine Engi-neering 8(3):29–41.

[74] Jedrzejewski J, Kwasny W, Potrykus J (1989) Beurteilung der Berechnungs-methoden fur die Bestimmung der Energieverluste in Walzlagern. Schmier-ungstechnik 8:243–244.

[75] Jedrzejewski J, Modrzycki W, Kowal Z, Kwasny W, Winiarski Z (2007) PreciseModelling of HSC Machine Tool Thermal Behaviour. Journal of Achievements inMaterials and Manufacturing Engineering 24:245–252.

[76] Jones AB (1960) A General Theory for Elastically Restrained Ball and RadialRoller Bearings Under Arbitrary Load and Speed Conditions. ASME Journal ofBasic Engineering 82:309–320.

[77] Jorgensen BR, Shin YC (1997) Dynamics of Machine Tool Spindle/BearingSystems Under Thermal Growth. Journal of Tribology 119(4):875–882.

[78] Jorgensen BR, Shin YC (1998) Dynamics of Spindle-Bearing Systems at HighSpeeds Including Cutting Load Effects. Journal of Manufacturing Science andEngineering 120(2):387–394.

[79] Jun MB, Ozdoganlar OB, DeVor RE, Kapoor SG, Kirchheim A, Schaffner G(2002) Evaluation of a Spindle-Based Force Sensor for Monitoring and FaultDiagnosis of Machining Operations. International Journal of Machine Tools andManufacture 42(6):741–751.

[80] Kalveram M (2005) Analyse und Vorhersage der Rotordynamik und Prozessst-abilitat beim Hochgeschwindigkeitsfrasen. Ph.D. thesis, Universitat Dortmund.

[81] Kern S, Schiffler A, Nordmann R, Abele E (2008) Modelling and ActiveDamping of a Motor Spindle with Speed-Dependent Dynamics. Vibrationin Rotating Machinery, International conference, vol. 1, 465–475.

[82] Kim SM, Lee KJ, Lee SK (2002) Effect of Bearing Support Structure on the High-Speed Spindle Bearing Compliance. International Journal of Machine Tools andManufacture 42(3):365–373.

[83] Kim SM, Lee SK (2001) Prediction of Thermo-Elastic Behavior in a Spindle-Bearing System Considering Bearing Surroundings. International Journal ofMachine Tools and Manufacture 41(6):809–831.

[84] Kirchheim A, Schaffner G, Jeck N (2000) Zerspankraftuberwachung mit spin-delintegrierten Kraftsensoren. Kistler Instrumente AG Winterthur, Schweiz.

[85] Klocke F, Reuber M (2000) Sicheres Freiformfrasen mit On-line-Prozessu-berwachung. wt - Werkstattstechnik online 04-2000:119.

[86] Klueber Lubrication MK (2006) Das Geheimnis des Rundlaufs. Firmenschrift.[87] Kolar P (2007) Modeling of a Machine Tools Spindle using a Hybrid Model .[88] Kowal Z, Kwasny W, Jedrzejewski J, Winiarski Z (2005) Increasing Machining

Centre’s Operational Flexibility. Machine Engineering 5:155–164.[89] Kowal Z, Kwasny W, Winiarski Z (2004) The Modelling of Motorized Spindle

Properties with Preload Bearing Sets. Machine Engineering 4:296–303.[90] Kroh R (2007) Schneller und sicherer Werkzeugwechsel mit Hilfe eines Linear-

motors. http://www.maschinenmarkt.vogel.de/themenkanaele/produktion/spanendefertigung/spannmittelundsysteme/articles/97570/.

[91] Kuljanic E, Sortino M, Totis G (2008) Multisensor Approaches forChatter Detection in Milling. Journal of Sound and Vibration 321(4–5):672–693.

[92] Lacey S, Wardle F, Poon S (1983) High Speed Bearing for CNC Machine ToolSpindles. Chartered Mechanical Engineering 30:51–56.

[93] LEViTEC. Internetprasenz. http://levitec.lt-i.com/.[94] Li G, Lin Z, Allaire PE, Huang B, Jiang W, Zorzi SE, Bartlett RO (2001)

Stabilization of a High Speed Rotor With Active Magnetic Bearings by aPiecewise mu-Synthesis Controller. 6th Symposium on Magnetic SuspensionTechnology, Turin, Italy, October 2001, .

[95] Li H, Shin YC (2004) Analysis of Bearing Configuration Effects on HighSpeed Spindles Using an Integrated Dynamic Thermo-MechanicalSpindle Model. International Journal of Machine Tools and Manufacture44(4):347–364.

[96] Li H, Shin YC (2004) Integrated Dynamic Thermo-Mechanical Modeling ofHigh Speed Spindles, Part 1: Model Development. Journal of ManufacturingScience and Engineering 126(1):148–158.

[97] Li H, Shin YC (2004) Integrated Dynamic Thermo-Mechanical Modeling ofHigh Speed Spindles, Part 2: Solution Procedure and Validations. Journal ofManufacturing Science and Engineering 126(1):159–168.

[98] Lim S, Park SM, Kim KI (2005) AI Vibration Control of High-Speed RotorSystems Using Electrorheological Fluid. Journal of Sound and Vibration 685–703. 284/2005/3-5.

[99] Lin CW, Tu JF, Kamman J (2003) An integrated Thermo-Mechanical-DynamicModel to Characterize Motorized Machine Tool Spindles During Very HighSpeed Rotation. International Journal of Machine Tools and Manufacture43(10):1035–1050.

[100] Lin NTLJF (2001) A New Method for the Analysis of Deformation and Load in aBall Bearing With Variable Contact Angle. Journal of Mechanical Design123:304–312.

[101] Mane I, Gagnol V, Bouzgarrou B, Ray P (2008) Stability-Based Spindle SpeedControl During Flexible Workpiece High-Speed Milling. International Journalof Machine Tools and Manufacture 48:184–194.

[102] Makino (2005) MAG Series, Five-Axis Horizontal Machining Centers. Kata-log:5.

[103] Mapal (2007) Mapal Neuheiten. News EMO 2007. http://www.mapal.com/download/news_09_07.pdf.

[104] Mekid S, Jedrzejewski J (2009) Introduction to Precision Machine Design andError Assessment. CRC Press, Boca Raton. (Chapter 3, pp. 75–126).

[105] Metzele M (2004) Piezoaktoren regeln die Lagervorspannung. wt - Werk-stattstechnik online 94/2004/7-8:344–349.

[106] Moon JD, Kim BS, Lee SH (2006) Development of the Active Balancing Devicefor High-Speed Spindle System Using Influence Coefficients. InternationalJournal of Machine Tools and Manufacture 46(9):978–987.

[107] Morgan G, Cheng RQ, Altintas Y, Ridgway K (2007) An Expert TroubleshootingSystem for the Milling Process. International Journal of Machine Tools andManufacture 47(9):1417–1425.

[108] Namazi M, Altintas Y, Abe T, Rajapakse N (2007) Modeling and Identificationof Tool Holder-Spindle Interface Dynamics. International Journal of MachineTools and Manufacture 47(9):1333–1341.

[109] Neuenschwander U (2005) Wie man sich spindelt, so frast man!. Werkstattund Betrieb 138(7–8):75–79.

[110] Neumeuer D, Kaufeld M (2008) Active Balancing for High Speed Machining –Objectives and Limits. 7th International Conference on High Speed MachiningProceedings, 28. und 29. Mai 2008, Darmstadt, Meisenbach Verlag GmbH,Bamberg, ISBN: , pp. 978-3-87525-274-3179–184.

[111] N.N. (2007) MMS Lagerschmierung (uberwacht) BAL 30995090. Bedienung-sanleitung, Bielomatik Leuze GmbH+Co KG.

[112] Ophey L (1985) Dampfungs- und Steifigkeitseigenschaften vorgespannterSchragkugellager/Lothar Ophey. Lehrstuhl fur Werkzeugmaschinen. Ph.D. thesis,Aachen, Techn. Hochsch.

[113] o.V. (1991) Kurzkegel mit Plananlage. Industrieanzeiger, 113(53–54):95.[114] Palm J, Fuchs G (2004) Trockenbearbeitung/MMS wird Industrie-Standard.

WB - Werkstatt und Betrieb 137(12):13–18.[115] Palmgren A (1959) Ball and Roller Engineering. Burbank, Philadelphia.[116] Park SS, Altintas Y (2004) Adaptive Control and Monitoring Using the Spindle

Integrated Force Sensor System. Manufacturing Engineering and MaterialsHandling Engineering. Paper no. IMECE2004-60395 (ASME 2004 InternationalMechanical Engineering Congress and Exposition (IMECE2004) November13–19, 2004, Anaheim, California, USA). pp. 41–48.

[117] Park SS, Altintas Y (2004) Dynamic Compensation of Spindle Integrated ForceSensors With Kalman Filter. Journal of Dynamic Systems Measurement andControl 126(3):443–452.

[118] Poplawski JV, Peters SM, Zaretsky EV (2001) Effect of Roller Profile onCylindrical Roller Bearing Life Prediction – Part I: Comparison of BearingLife Theories. Tribology Transactions 44(3):339–350.

[119] Poplawski JV, Peters SM, Zaretsky EV (2001) Effect of Roller Profile onCylindrical Roller Bearing Life Prediction – Part II: Comparison of RollerProfiles. Tribology Transactions 44(3):417–427.

[120] Przemieniecki JS (1985) Theory of Matrix Structural Analysis. Courier DoverPublications.

[121] Rantatalo M, Aidanpaa JO, Goransson B, Norman P (2007) Milling MachineSpindle Analysis Using FEM and non-Contact Spindle Excitation andResponse Measurement. International Journal of Machine Tools and Manufac-ture 47(7–8):1034–1045.

[122] Reusner H (1987) The Logarithmic Roller Profile – The Key to SuperiorPerformance of Cylindrical and Tapered Roller Bearings. Ball Bearing Journal230:2–10.

[123] Richardson MH, Formenti DL (1982) Parameter Estimation From FrequencyResponse Measurements Using Rational Fraction Polynominals. IMAC Con-ference, Orlando, FL, .

[124] Ries M, Pankoke S, Gebert K (2006) Increase of Material Removal Rate with anActive HSC Milling Spindle. Adaptronic Congress 2006, 3rd–4th May, Gottin-gen, .

http://www.maschinenmarkt.vogel.de/themenkanaele/produktion/spanendefertigung/spannmittelundsysteme/articles/97570/

http://www.maschinenmarkt.vogel.de/themenkanaele/produktion/spanendefertigung/spannmittelundsysteme/articles/97570/

http://levitec.lt-i.com/

http://www.mapal.com/download/news_09_07.pdf

http://www.mapal.com/download/news_09_07.pdf


[125] Rivin E (2000) Tooling Structure: Interface between Cutting Edge andMachine Tool. CIRP Annals Manufacturing Technology 49(2):591–634.

[126] Ronde U, Kades H (2007) Motorspindel mit einer Kuhleinheit. PatentschriftDE102006043550A1, issued 30.08.2007.

[127] Rothenbucher S, Schiffler A, Bauer J (2009) Die Speisung macht’s. WB -Werkstatt und Betrieb 142(7–8):62–65.

[128] Rouch KE, Kao JS (1979) A Tapered Beam Finite Element for Rotor DynamicsAnalysis. Journal of Sound and Vibration 66(1):119–140.

[129] Ryabov O, Mori K, Kasashima N (1998) Laser Displacement Meter Applicationfor Milling Diagnostics. Optics and Lasers in Engineering 30(3–4):251–263.

[130] Sandvik (2009) Coromant Capto Modular Tooling System (catalogue). http://www2.coromant.sandvik.com/coromant/downloads/catalogue/GER/Capto_2008.pdf.

[131] Sastry S, Kapoor S, DeVor R (2002) Floquet Theory Based Approach forStability Analysis of the Variable Speed Face-Milling Process. Journal ofManufacturing Science and Engineering 124(1):1–10.

[132] Schaeffler GruppeIndustrie H (2005) Hoheres Leistungsvermogen durchBeschichtungen–Beschichtete Walzlager der Schaeffler Gruppe. Firmenschrift.

[133] Schaeffler Gruppe Industrie S (2007) Hochgenauigkeitslager–Super PrecisionBearings. Firmenschrift.

[134] Scheer C, Hoffmann P (1999) Spindle-Integrated Force Sensors for Monitor-ing Drilling and Milling Processes. 9th Int. Conf. for Sensors, Transducers &Systems May 18–20, 1999, Nurnberg, Germany, .

[135] Schmitz TL, Powell K, Won D, Duncan GS, Sawyer WG, Ziegert JC (2007)Shrink Fit Tool Holder Connection Stiffness/Damping Modeling for FrequencyResponse Prediction In Milling. International Journal of Machine Tools andManufacture 47(July (9)):1368–1380.

[136] Schneeweiß M, Kohler S, Schwarzburger R, Pospischil T (2004) Kuhlschmier-ung bedarfsgerecht dosieren. WB - Werkstatt und Betrieb 137(11):54–58.

[137] Scholl S (2007) Spindelvorrichtung mit Innenkuhlung. PatentschriftDE102005030277B4, issued 31.10.2007.

[138] Schossig HP (2002) Neue Motorspindeln fur die Komplettbearbeitung. WB -Werkstatt und Betrieb 135(9):70–74.

[139] Schutt P (1978) Bewegungsverhaltnisse in leichtbelasteten schnell laufendenZylinderrollenlagern. Dissertation, TU Braunschweig.

[140] Shankar N, Bridger K, Regelbrugge ME, Winfough WR (1998) Smart SpindleUnit For Active Chatter Suppression Of A Milling Machine: I. Overview,Fabrication and Assembly. Smart Structures and Materials 1998 Industrialand Commercial Applications of Smart Structures Technologies 3326(1):160–166.

[141] Shiraishi M (1989) Scope of In-Process Measurement, Monitoring and Con-trol Techniques in Machining Processes – Part 3: In-Process Techniques forCutting Processes and Machine Tools. Precision Engineering 11(1):39–47.

[142] SKF Gruppe, SPWVG (2006) Prazisions-Schragkugellager. Hauptkatalog, Fir-menschrift.

[143] Smith GT (2008) Cutting Tool Technology, vol. 1. Springer, London.[144] Smith S, Tlusty J (1992) Stabilizing Chatter by Automatic Spindle Speed

Regulation. CIRP Annals Manufacturing Technology 41(1):433–436.[145] Soliman E, Ismail F (1998) A Control System for Chatter Avoidance by

Ramping the Spindle Speed. Journal of Manufacturing Science and Engineering120(4):674–683.

[146] Soshi M, Ishiguro H, Yamazaki K (2009) A Study on the Development of aMulti-Purpose Spindle System for Quality Productive Machining. CIRP AnnalsManufacturing Technology 58(1):327–330.

[147] Spachtholz G (2008) Erweiterung des Leistungsbereichs von Spindellagern.Dissertation, WZL RWTH Aachen.

[148] Spur G, Feil A (1994) Diagnosis of Grease Lubricated Main Spindle Bearings.Production Engineering I(2):79–80.

[149] Stein JL, Tu JF (1994) A State-Space Model for Monitoring ThermallyInduced Preload in Anti-Friction Spindle Bearings of High-Speed MachineTools. Journal of Dynamic Systems Measurement and Control 116(3):372–386.

[150] Steinert T (1995) Das Reibmoment von Kugellagern mit bordgefuhrtem Kafig.Dissertation, RWTH-Aachen.

[151] Sternagel R, Kreuser J (2001) Hybrid- und Keramikwalzlager fur extremeBeanspruchungen. Tribologiecehsp sp025/cehsp sp025/Schmierungstechnik48(4):11–20.

[152] Stoll R (1990) Grenze nicht erreicht - Optimierte Steilkegelspannung hatnoch Reserven. MM - Maschinenmarkt Das Industriemagazin 96(38):68–70.

[153] Sud O, Davies P, Halling J (1974) The Thermal Behavior of Rolling BearingAssemblies subjected to preload. Wear 27:240–246.

[154] Sykora H (2004) Diese Systeme bestehen die HSC-Probe. WB - Werkstatt undBetrieb 137(9):56–59.

[155] Tamisier V, Font S, Lacour M, Carrere F, Dumur D (2001) Attenuation ofVibrations due to Unbalance of an Active Magnetic Bearings Milling Electro-Spindle. CIRP Annals – Manufacturing Technology 50(1):255–258.

[156] Thomas D, Wilson J, Wilson R (1973) Timoshenko Beam Finite Elements.Journal of Sound and Vibration 31(3):315–330.

[157] Tullmann U (1999) Das Verhalten axial verspannter schnelldrehender Schrag-kugellager. Ph.D. thesis, WZL RWTH Aachen.

[158] Tonshoff H, Denkena B, Gotz T (2002) Piezoelectric Actuator Based PreloadControl Unit for Machine Tool Spindles. Production Engineering 9(1):117–122.

[159] Tonshoff H, Gotz T (2002) Piezoaktoren zur Vorspannungsregelung in Werk-zeugspindeln. Antriebstechnik 41(3):45–49.

[160] Tonshoff H, Karpuschewski B, Mandrysch T, Inasaki I (1998) Grinding ProcessAchievements and their Consequences on Machine Tools Challenges andOpportunities. CIRP Annals Manufacturing Technology 47(2):651–668. (18pp.).

[161] To C (1981) A Linearly Tapered Beam Finite Element Incorporating ShearDeformation And Rotary Inertia For Vibration Analysis. Journal of Sound andVibration 78(4):475–484.

[162] Trojahn W, Zoch H (2000) Entwicklung und Anwendung von Walzlagern ausCronidur130. Stahl ;(3)44–46.

[163] Tu JF, Katter JG (1996) Bearing Force Monitoring in a Three-Shift ProductionEnvironment. Tribology Transactions 39:201–207.

[164] Tu JF, Stein JL (1996) Active Thermal Preload Regulation for Machine ToolSpindles with Rolling Element Bearings. Journal of Manufacturing Science andEngineering 118(4):499–505.

[165] Young DL, Weaver W, Timoshenko S (1990) Vibration Problems in Engineer-ing. John Wiley & Sons, Incorporated.

[166] Walter R (2006) Mit direkter Kuhlung zu mehr Genauigkeit. WB - Werkstattund Betrieb 139(6):129–130.

[167] Weck M, Brecher C (2006) Werkzeugmaschinen–Konstruktion und Berechnung,vol. 8 of 2. Springer, Berlin.

[168] Weck M, Krell M (2000) Tool and Spindle Systems with High Damping.Production Engineering 7(2):79–84.

[169] Weck M, Platen S (2003) Intelligente Spindelfunktionen. wt - Werkstattstech-nik online 93/2003/7–8:517–521.

[170] Weck M, Reinartz T (2004) Determination of the Load Capacity of HSKInterfaces. Production Engineering 11(1):99–102.

[171] Weck M, Schubert I (1994) New Interface Machine/Tool: Hollow Shank. CIRPAnnals – Manufacturing Technology 43(1):345–348.

[172] Weck M, Spachtholz G (2003) 3- and 4-Contact Point Spindle Bearings – ANew Approach For High Speed Spindle Systems. CIRP Annals ManufacturingTechnology 52:311–316.

[173] Wemhoner J (2001) Keramik–ein Meilenstein auf dem Weg zum Hochleistungs-walzlager. Tagungsband zum Symposium ‘‘Pulvermetallurgie–Schlussel zurEffizienzsteigerung’’. Hagen:15–16.

[174] Wiche E (1967) Die radiale Federung von Walzlagern bei beliebiger Lagerluft.Konstruktion 19:184–192.

[175] Wiemer AL (1969) Luftlagerung. VEB Verlag Technik, Berlin.[176] Wittstock V, Munzer J, Neugebauer R (2007) Aktive Lagekorrektur einer HSC-

Hauptspindel. VDI-Z integrierte Produktion 149(6):71–74.[177] Xiong GL, Yi JM, Zeng C, Guo HK, Li LX (2003) Study of the Gyroscopic Effect of

the Spindle on the Stability Characteristics of the Milling System. Journal ofMaterials Processing Technology 138(1–3):379–384.

[178] Yokoyama T (1990) Vibrations of a Hanging Timoshenko Beam UnderGravity. Journal of Sound and Vibration 141(2):245–258.

[179] Yoshioka H, Matsumara S, Hashizume H, Shinno H (2006) MinimizingThermal Deformation of Aerostatic Spindle System by Temperature Controlof Supply Air. JSME International Journal Series C Mechanical Systems MachineElements and Manufacturing 49(2):606–611.

[180] Zverev IA, Eun IU, Chung W, Lee CM (2003) Simulation of Spindle UnitsRunning on Rolling Bearings. The International Journal of Advanced Manufac-turing Technology 21(10):889–895.

[181] Zverev IA, Eun IU, Hwang Y, Chung W, Lee CM (2006) An Elastic DeformationModel of High-Speed Spindle Units. International Journal of Precision Engi-neering and Manufacturing 7:39–46.

http://www2.coromant.sandvik.com/coromant/downloads/catalogue/GER/Capto_2008.pdf



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