Machine Tool Spindle Units

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

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CIRP Annals - Manufacturing Technology jou rnal homep age : ht t p: // ees .e lse vi er. com/ci rp/ def a ult . asp

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

Institute of Production Management, Technology and Machine Tools (PTW), Technische Universita¨t Darmstadt, Germany Manufacturing Automation Laboratory, University of British Columbia, Vancouver, Canada Laboratory for Machine Tools and Production Engineering (WZL), RWTH Aachen, Germany

A R T I C L E I N F O

A B S T R A C T

Keywords: Spindle Mechatronic Machine tools

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. ß 2010 CIRP.

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 removal rates and quality of the machined parts. This paper reviews the current state and presents research challenges of spindle technology. 1.1. Historical review Classically, main spindles were driven by belts or gears and the rotational speeds could only be varied by changing either the transmission ratio or the number of driven poles by electrical switches. Later simple electrical or hydraulic controllers were developed and the rotational speed of the spindle could be changed by means of infinitely adjustable rotating transformers (Ward Leonard system of motor control). The need for increased productivity led to higher speed machining requirements which led to the development of new bearings, power electronics and inverter systems. The progress in the field of the power electronics (static frequency converter) led to the development of compact drives with low-cost maintenance using high frequency three-phase asynchronous motors. Through the early 1980’s high spindle speeds were achievable only by using active magnetic bearings. Continuous developments in bearings, lubrication, the rolling element materials and drive

* Corresponding author. 0007-8506/$ – see front matter ß 2010 CIRP. doi:10.1016/j.cirp.2010.05.002

systems (motors and converters) have allowed the construction of direct drive motor spindles which currently fulfill a wide range of requirements. A historical review of spindle technology is given in Fig. 1. 1.2. Principal setup Today, the overwhelming majority of machine tools are equipped with motorized spindles. Unlike externally driven spindles, the motorized spindles do not require mechanical transmission elements like gears and couplings. A motor spindle mainly consists of the elements shown in Fig. 2. The spindles have at least two sets of mainly ball bearing systems. The bearing system is the component with the greatest influence on the lifetime of a spindle. Most commonly the motor is arranged between the two bearing systems. Due to high ratio of ‘power to volume’ active cooling is often required, which is generally implemented through water based cooling. The coolant flows through a cooling sleeve around the stator of the motor and often the outer bearing rings. Seals at the tool end of the spindle prevent the intrusion of chips and cutting fluid. Often this is done with purge air and a labyrinth seal. A standardized tool interface such as HSK and SK is placed at the spindles front end. A clamping system is used for fast automatic tool changes. Ideally, an unclamping unit (drawbar) which can also monitor the clamping force is needed for reliable machining. If cutting fluid has to be transmitted through the tool to the cutter, adequate channels and a rotary union become required features of the clamping system. Today, nearly every spindle is equipped with sensors for monitoring the motor temperature (thermistors or thermocouples) and the position of the clamping system. Additional sensors for monitoring the bearings, the drive and the process stability can be attached, but are not common in many industrial applications.

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

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Fig. 1. Historical review.

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Fig. 4. Spindles available on the market [PTW].

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

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

The focus of this paper is the spindle technology. To clearly limit the topics of this paper the chosen system boundaries are illustrated in Fig. 3.

Current spindle manufacturers offer wide variety of spindle solutions for the application areas shown in Fig. 4. Spindles with high power and high speeds are mainly developed for the machining of large aluminum frames in the aerospace industry. Spindles with extremely high speeds and low power are used in electronics industry for drilling printed circuit boards (PCB).

Particular attention was paid to the increase of the reliable reachable rotational speeds in the past. However, the focus has changed towards higher torque at speeds up to 15,000 rpm. Because of Increased requirements in reliability, life-cycle and predictable maintenance the ‘condition monitoring’ systems in motor spindles have become more important. Periodic and/or continuous observation of the spindle status parameters is allowing detection of wear, overheating and imminent failures. Understanding the life cycle cost (LCC) of the spindles has steadily gained importance in predicting their service period with maintenance, failure and operational costs.

1.4. Actual development areas in industry

2. Fields of application and specific demands

Current developments in motor spindle industrial application focus on motor technology, improving total cost of ownership (TCO) and condition monitoring for predictive maintenance (see Fig. 5). Another central issue is the development of drive systems which neutralize the existing constraints of power and output frequency while reducing the heating of the spindle shaft.

Spindles are developed and manufactured for a wide range of machine tool applications with a common goal of maximizing the metal removal rates and part machining accuracy. Fig. 6 gives an overview regarding the application areas, material grades and the resultant requirements concerning rotational speed, power, torque and accuracy. The work materials range from easy to machine materials like aluminum at high speeds with high power spindles, to nickel and

1.3. State of the art

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Fig. 3. Overview of the spindle system and its integration into the machine tool.

Fig. 6. Application areas and branches.

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titanium alloys which require spindles having high torque and stiffness at low speeds. Cutting work materials with abrasive carbon or fiber-reinforced plastics (FRP) content need good seals at the spindle front end. Spindles for drilling printed circuit boards operate in the angular speed range of 100,000 to 300,000 rpm. The increase in productivity and speed in this application field over the last few years was possible with the development of precision air bearings. Spindles used in die and mould machining have to fulfill the roughing operations (high performance cutting, HPC) at high feed rates as well as the finishing processes (high-speed cutting, HSC) at high cutting speeds. Depending on the strategy and the machinery of the mould and die shop either two different machine tools equipped with two different spindles are used or one machine is equipped with a spindle changing unit. Another possibility is to use a spindle which can fulfill both, HSC and HPC conditions, but this still remains a compromise regarding overall productivity. Aerospace spindles are defined by high power as well as high rotational 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 is necessary, which requires stiff spindles with bearings having minimum runout. The present internal cylindrical grinding spindles have a runout requirement of less than 1 mm. Spindle units which are used mainly for boring and drilling operations require high axial stiffness, which is achieved by using angular contact bearings with high contact angles. On the contrary, high-speed milling operations use spindles with bearings having small contact angles in order to reduce the dependency of radial stiffness on the centrifugal forces. Contemporary machining centers tend to have multi functions where milling, drilling, grinding and sometimes honing operations can be realized on the same work piece. The bottleneck for the enhancement of the multi-technology machines is still the spindle, which cannot satisfy all the machining operations with the same degree of performance. Reconfigurable and modular machine tools require interchangeable spindles with standardized mechanical, hydraulic, pneumatic and electrical interfaces. 3. Spindle analysis The aim of modeling and analysis of spindle units is to simulate the performance of the spindle and optimize its dimensions during the design stage in order to achieve maximum dynamic stiffness and increased material removal rate with minimal dimensions and power consumption. Modeling of the cutting process and the prediction of chatter stability can be found in previous key note articles [10] and are not covered in this article. The mechanical part of the spindle assembly consists of hollow spindle shaft mounted to a housing with bearings. Angular contact ball bearings are most commonly used in high-speed spindles due to their low-friction properties and ability to withstand external loads in both axial and radial directions. The spindle shaft is modeled by beam, brick or pipe elements in finite element environment. The bearing stiffness is modeled as a function of ball bearing contact angle, preload caused by the external load or thermal expansion of the spindle during operation. The equation of motion is derived in matrix form by including gyroscopic and centrifugal effects, and solved to obtain natural frequencies, vibration mode shapes and frequency response function at the tool attached to the spindle. If the bearing stiffness is dependent on the speed, or if the spindle needs to be simulated under cutting loads, the numerical methods are used to predict the vibrations along the spindle axis as well as contact loads on the bearings. The model allows the simulation of interaction between the cutting process and spindle structure (Fig. 7). Spindle simulation models allow for the optimization of spindle design parameters either to achieve maximum dynamic stiffness at all speeds for general operation, or to reach maximum axial depth of cut at the specified speed with a designated cutter for a specific

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Fig. 7. Flowchart of spindle analysis.

machining application. The objective of cutting maximum material at the desired speed without damaging the bearings and spindle is the main goal of spindle design while maintaining all other quality and performance metrics, e.g. accuracy and reliability. 3.1. Experimental modeling The dynamic behavior of an existing spindle is most quickly obtained by measuring its frequency response function (FRF) between force and displacement at the tool tip. The measured FRF can be curve fitted to estimate the natural frequencies, damping ratios and stiffness values at a range of frequency where the spindle structure may cause vibrations during machining. The experimental measurement of FRFs is practical to assess the dynamic stiffness and identify chatter free cutting conditions in process planning of part machining operations. However, the following difficulties need to be kept in mind: only a small part of the rotating shaft is accessible for testing, hence modeling of entire spindle is not possible operational speed and temperature mainly influence the eigenvalues, but the measurement of FRFs when the spindle rotates is quit difficult curve fitting or other methods to extract parameters out of the measured input and output data does not always lead to accurate identification of the spindle’s dynamic parameters. Traditionally, the FRF is measured by exciting the spindle at the tool tip by impact hammers or shakers manually. There have been several attempts to measure the FRFs automatically. The authors in [20] use a piezo actuator connected in series with a force measurement sensor and a displacement sensor to estimate the FRF at tool tip at standstill. The authors in [2,121] use an electromagnetic actuator to excite the shaft and measure the displacement with a non-contact probe at various rotating speeds. Various authors use an impact hammer for exciting the shaft. This type of excitation is also used for hitting the rotating shaft under high speeds as presented in [80]. In this work a special mechanism was designed to hit the shaft with the impact hammer under repeatable conditions. The measured data can also be used to verify or update the results from theoretical models especially to adjust damping coefficients as shown in [87]. Direct methods to achieve parametric models for further simulations are an issue of system identification and cause problems like selecting the order of the model, models without minimum phase and stability issues [123]. 3.2. Theoretical modeling Theoretical models are based on physical laws, and used to predict and improve the performance of spindles during the design stage. The models provide mathematical relation between the inputs F (force, speed) and the outputs q (deflections, bearing loads, and temperature). The mathematical models can be

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expressed in state space forms or by a set of ordinary differential equations. In both cases linear or nonlinear behavior of the spindles can be modeled. 3.2.1. Mechanical modeling of shaft and housing Finite element (FE) methods are most commonly used to model structural mechanics and dynamics of the spindles. The method is based on discretization of the structure at finite element locations by partial derivative differential equations. The analysis belongs to the class of rotor-dynamic studies where the axis-symmetric shaft is usually modeled by beam elements, which lead to construction of mass (Me) and stiffness (Ke) matrices. Timoshenko beam element is most commonly used because it considers the bending, rotary inertia and shear effects, hence leads to 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 FEA software ANSYS is also an implementation of the Timoshenko theory and use the mass matrix from [178] and stiffness matrix from [120]. As an example in the finite element model in Fig. 8, the black dots represent nodes, and each node has three Cartesian translational displacements and two rotations [7]. The pulley is modeled as a rigid disk, the bearing spacer as a bar element, and the nut and sleeve as a lumped mass. The spindle in this case has two front bearings in tandem and three bearings in tandem at the rear. The five bearings are in overall back-to-back configuration. The tool is assumed to be rigidly connected to the tool holder which is fixed to the spindle shaft rigidly or through springs with stiffness in both directions translation and rotation. The flexibility of the spindle mounting has to be reflected in the model of the spindle-machine system. Springs are also used between the spindle housing and spindle head, whose stiffness is obtained from experience.

Fig. 9. Comparisons of experimentally determined tool-tip FRF and simulated FRF [7].

[47,99] describe this term as spin-softening effect. The skew symmetric matrix CV in case of rotating reference frame carries the damping of the rotating parts and in the other case carries the damping of the non-rotating parts. One way to match the model to the physical system is to solve the eigenvalue problem of the undamped system and then fit the analytical solution to the experimental data by adding empirical modal damping ratios [87] and/or by changing geometrical parameters to match the natural frequencies which result from experimental measurements described in the previous section [7,99]. Therefore the system can be rewritten in state space notation and transformed to a Jordan canonical form. Fig. 9 shows the result of the modeled FRF at tool tip as compared with the experimentally measured FRFs for the spindle given in Fig. 10. The closeness of the experimentally measured and simulated FRFs determine the accuracy of the theoretical models in analyzing the spindle behavior during the design stage. Several researchers have developed FE packages dedicated to the virtual design and performance analysis of spindles [26,7,27].

2

M q¨ þ ðC VGÞq˙ þ ðK V M V þ VC V Þq ¼ F The vector q is the generalized displacement vector which includes the movement of all nodes in the chosen degrees of freedom. The reference frame of the vector q can be fixed to the rotating shaft of the spindle or be an inertial frame fixed to space. The symmetric matrices M and K are assembled out of the single element matrices Me and Ke as described in detail in [49]. These matrices are independent of the chosen reference frame. The stiffness matrix KB of the bearing support is also added in K. The symmetric matrix C is also independent of the chosen reference frame and contains structural damping of the shaft (rotating) and the damping of the housing (non-rotating). The skew symmetric matrix G is often named as gyroscopic matrix but in fact only includes the case of inertial reference frame for pure gyroscopic effects. In the case of rotating reference frame, the matrix G also includes the coriolis acceleration coupling terms which are proportional to the mass as shown in [99]. The term V2MV only exists in the rotating reference frame and adds centrifugal forces to the system. The matrix MV is symmetric and positive definite and so [(Fig._8)TD$IG] reduces the stiffness of the system. Therefore the authors in

3.2.2. Mechanical modeling of tool–spindle interfaces The FRF at tool tip is important for stability predictions of the machine tool. The flexibility of the assembly/interface tool-holder and spindle usually dominates the dynamics of the spindle. Due to the large number of spindle, holder, and tool combinations that may be available in a particular production facility, the required testing time for experimental modeling can be significant. There are approaches to model the shaft and housing with finite element theory as mentioned above and to couple different tool holder and tools. Ertu¨rk et al. [41,42,108] use a receptance coupling and structural modification method to connect the tool-holder to the spindle shaft. Schmitz et al. [135] presents a receptance coupling substructure analysis method for modeling a shrink fit tool holder. This work considers distributed springs and dampers between the tool and holder along the interference contact surface. In difference to that the conventional modeling uses one concentrated spring damper element between the portions of the tool inside and outside the holder. 3.3. Modeling of angular contact ball bearings Angular contact ball bearings (Fig. 11) are commonly used in high-speed spindles. The bearings require preloading to prevent skidding in order to maintain rotational accuracy and sufficient [(Fig._10)TD$IG]

Fig. 8. The finite element model of the spindle-bearing-machine-tool system from Fig. 10 [7].

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

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

stiffness in both the radial and axial directions to support the basic operational requirements. Basically, there are two types of bearing preloads: rigid preload and constant preload (see Fig. 22). Modeling of the bearing support of the spindle shaft is important in predicting the spindle’s structural deformations during machining. The bearing parameters can vary during the operation, which in turn affect the stiffness. For example relative displacement of raceways curvature centers, relative speed between inner and outer ring, axial and radial load, temperature difference between inner and outer rings and contact angles may vary during cutting. The basic equations to evaluate the tangential stiffness matrix KB for each bearing under working conditions are based on the models proposed by Jones [76] and De Mul et al. [33]. The theory of contact between balls and rings was based on Hertz [61]. The effects of cutting loads [25,78], frictional moments [112,157] and thermal deformations [63,95] have been incorporated to the classical bearing models by spindle researchers [64,81,25,100, 60,11]. The bearing models considers the following conditions: relative speed between inner and outer rings speed-related centrifugal forces and gyroscopic moments acting on rolling elements relative displacement of centers of ring groove curvatures due to thermal expansion The resulting bearing stiffness KB depends on the size of the bearing balls or rollers, curvature of the bearing rings, and contact angle. However, the bearing dynamics changes with preload, cutting forces, spindle speeds and thermal expansion. These factors lead to a nonlinear modeling of the spindle system. The tangential stiffness matrix KB of the bearings is only valid at one operating condition; hence the system dynamics become nonlinear and appear non-repeatable. Because of the nonlinear dependency of the radial bearing stiffness on the axial or radial load, the bearing stiffness changes with cutting force dynamically. In [25] it is shown that the axial force has a larger effect in the bearing stiffness than radial forces and therefore the variation of bearing stiffness is matched with the frequency of cutting forces in axial direction. It can be shown that in case of periodic cutting loads the matrix KB(t) is periodic at the harmonics of the shaft speed. In Fig. 12 this effect is shown for different preload mechanism. Besides the aspects of preload and load of bearings in high spindle speeds, the bearing ball centrifugal forces and gyroscopic moments can be of significant magnitude such that inner ring contact angles tend to increase and outer ring contact angles tend to decrease. The natural frequency of the system is related to the bearing stiffness, hence it increases with preload due to increased bearing stiffness, but decrease with spindle speed due to centrifugal forces as experimentally demonstrated in

Fig. 12. Radial stiffness changing with cutting forces for spindle-bearing-system Fig. 10 [25].

[25,2,99,121]. Fig. 13 shows one result of experimental measurements of the dynamic behavior of a motorized spindle over a range of speed. 3.3.1. Thermal modeling of motorized spindle units Limits of a spindle’s speed, reliability and performance are usually constrained by properties of its bearings, which are affected by the uneven thermal expansion of spindle parts and degraded 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 the estimation of limiting speed of rolling bearings. When the product Dm times n exceeds 0.5 106 mm/min, the operation is considered as high-speed and the spindle design must be able to deal with the heat produced and subsequent thermal expansion. In extreme cases, the Dmn value can be as high as 4 106 mm/min (jet lubrication). The limiting speed of each application depends on the amount of heat produced by a particular design and on the mechanical ‘‘sensitivity’’ of the design to already developed temperatures (thermal expansion affecting bearings). The important factors are: bearing type/size, lubrication, bearing configuration, type of preload and overall heat management within the spindle. There is a link between thermal and mechanical behavior of spindles. Uneven thermal expansion changes the mechanical condition of bearings which in turn affects the amount of heat [(Fig._13)TD$IG]produced, which is fed back to the system and further increases the

Fig. 13. Experimentally estimated 1st (––) and 2nd (–*–) radial eigenfrequencies of a motorized spindle vs. spindle speed [2].

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thermal load and heat. The system can be unstable, and increasing thermal expansion and resulting bearing preload can damage the bearings (see also Fig. 20). Spindle builders use three basic principles to avoid excessive preloading of bearings: Constant preload of bearings by springs, hydraulic or piezo actuators ([158,36,105]) Rigid preload, theoretically thermally symmetric design Radial flexible housing [104] 3.3.2. Heat sources in spindles Principally, there are three main sources of heat in spindles: friction within the bearings power losses by the motor – depends on type (asynchronous or synchronous) motors and frequency converter [127] cutting power – since most heat is transferred to the chip or localized at the tool’s cutting edge, it usually does not play an important role. This is a more significant effect in materials with low thermal conductivity like titanium friction in tool clamping device The heat generation in bearings is challenging to predict. Theories presented in [115,55] and [74] distinguish three main sources of friction which occur in ball-groove contact: load friction caused by rolling and proportional to contact forces viscous friction caused by viscosity of lubricant – nonlinearly proportional to speed and mean diameter of bearing [115] spinning friction (spin/roll ratio) caused by kinematics of rolling elements. If contact angles differ from ‘‘zero’’ value, the rolling element necessarily spins in one of the bearing grooves. Currently none of these theoretical analyses have predicted the temperature distribution and resulting thermal deformations of the spindle bearings with sufficient accuracy. Some authors tried to derive dedicated formulations by using ‘coast tests’ [19,62] (see Fig. 14), or by other experiments [77] measuring passive moments of bearings. However, generalized models applicable to a wide range of bearings, lubrications and operating conditions have not yet been developed. 3.3.3. Heat sinks and heat transfer Heat transfer by conduction within spindle parts can be modeled using 2D or 3D FEM elements, once the boundary conditions are known. It is difficult to model heat transfer correctly at thermal joints – between surfaces mounted with clearance which changes during thermal expansion, for example the interface between the bearing outer race and the inner diameter of the housing. The authors in [104] show an example of the relationship between bearing temperature and working clearance

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as a function of spindle speed. In such cases dedicated nonlinear thermal resistance elements with properties based on heat transfer theories must be created [69]. Heat sinks are usually caused by convection, conduction and radiation on spindle surfaces with the following sources: convection to coolant fluids (motor and bearing cooling) convection and radiation to surrounding air or oil–air lubrication conduction through the housing flange to the spindle head Specifically the coefficients of heat convection between surfaces and fluids (air, coolant) are difficult to predict and depend on many parameters, hence care must be taken to specify these boundary conditions according to established thermal theories [69]. 3.3.4. History of ‘thermal preload’ prediction There has been a need to predict temperatures and thermal stability in the past, e.g. to avoid thermally induced seizure of bearings. In 1967 Burton and Staph [22] derived a general theory of thermal stability of angular contact ball bearings where the authors distinguished temperatures of various parts of a simple bearing assembly as stabilizing or destabilizing factors. In 1970 and 1972 Carmichael and Davies [29,28] investigated experimentally the effect of cooling of stationary outer parts of a two-bearing rigid assembly on preload experimentally. In 1974 Sud and Davies [153] investigated the effect of speed and the similarity of thermal preload with displacement of mechanical systems in time. They suggested a first order differential equation and a method to calculate its constant terms. These terms are based on thermal capacity, conductivity and convection as well as geometry and thermal expansion coefficients of a particular assembly. In 1983 Lacey et al. [92] conducted extensive experimental research on operational preload of rigid spindles. They investigated the effect of speed, lubrication and initial preload and indicated various types of 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 and publishing results on the thermo-mechanical behavior of highspeed spindle units: 1. Stein, Bossmanns, Lin, Tu, Harder: In 1994 they presented a predictive model of a simple two-bearing assembly [149]. The mechanical part of the model was based on simplified calculation of relative distance of bearing rings. The paper was followed by more detailed investigation of heat transfer in spindles [19] and its interface with bearing ring model [99]. 2. Jo¨rgensen, Li, Shin determine the steady state heat transfer by gradually introduced boundary conditions [77,96,97]. They use deMul’s bearing model [33] for calculating condition and properties of bearings. 3. Kim-Lee focused on the effect of radial bearing fit [82,83]. 4. Kowal, Jedrzejewski, Kwasny, Winiarski: program based on FEM and FDM [71,88,89,73]. The papers showed the results with comments, but the adopted bearing models were not presented. 5. Zverev, Eun, Chun, Lee: their model published in [180] and [181] has the same approach to the mechanical part of the problem as in [27], but they added a simplified heat transfer and thermal expansion model based on beam elements. 6. Holkup, Holy: model [63] and [62] used Jones’ bearing model, axis-symmetric heat transfer in 2D finite elements, effect of bearing and spacer radial fit, transient heat transfer (see Fig. 15). The common structure of the past models can be listed as:

Fig. 14. Frictional moment measured on a single SKF 7010CD bearing, constant preload Fa0 introduced by springs [62].

Rolling bearings are modeled with Jones’ or deMul’s bearing formulations (Groups 2, 5, 6). Steady state and transient analysis which also captured preload peaks are considered by all groups except (2).

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Modeling, prediction and optimization of spindles with respect to thermal issues has not been studied sufficiently in the literature. It is possible to [62] to build a model which performs a closed loop thermo-mechanical simulation using best available theories in each field. Special care must be taken when introducing thermal boundary conditions of such model (bearing friction and surface-fluid heat convection). Thermal displacements resulting from thermo-mechanical simulation can be passed to nonlinear ‘bearing models’ of purely mechanical models [27] and provide improved prediction of structural dynamic behavior of spindles.

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

Good model of heat transfer and thermal expansion based on FEM with 2D or 3D elements – all groups except (5). The nonlinear effect of bearing and spacer radial fit, radial flexibility of supporting structure (rings, housing, shaft) influencing heat transfer between parts as well as mechanical stiffness (3, 6). In principal, the structure of a predictive thermo-mechanical model must be integrated to simulation which updates mechanical as well as thermal nonlinearities during transient simulation [63] as shown in Fig. 16. Boundary conditions of thermal models are crucial for the correct prediction of thermal loads on the spindle. Since no general theory has been proven to be applicable for the estimation of bearing heat sources, they still need to be identified experimentally as performed by groups 1, 2 and 6. The displacement distribution in the spindle assembly caused by the temperature fields in the system is also studied in highspeed machining centers as shown in Fig. 17 [75,72].

4. Mechanical design Depending on the machine tool application area, the spindlebearing systems are subject to a complex array of requirements. For example, in high-speed cutting (HSC) applications on aluminum components, there is a need to combine high speeds with low stiffness [67] whereas heavy-duty machining of titanium- or nickel-based alloys require that the bearings must be able to absorb high cutting forces at low speeds of rotation [167]. Apart from the choice of a suitable type of bearing, optimum design of the bearing configuration makes a decisive contribution to the performance and service life of the main spindle [65]. 4.1. Bearing solution The following bearing types are used depending on the application requirements as shown in Fig. 18 [65]:

rolling bearings, electromagnetic bearings, aerostatic bearings, hydrostatic bearings, hydrodynamic bearings.

3.3.6. Summary on thermal aspects

Thermal issues significantly affect properties (stiffness, life, accuracy) of spindles.

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Fig. 16. Closed loop of thermo-mechanical transient analysis.

Fig. 17. Measurements of displacements in Z-axis over time at different speeds on a high speed machining center [72].

4.1.1. Ball bearings At specific speed coefficient (Dmn value) up to a maximum of 3.0 106 mm/min (e.g. milling operations), main spindles with high axial and radial stiffness are generally mounted on rolling bearings. High-precision spindle bearings combine good radial run-out and stiffness properties with low assembly and maintenance effort, and a good cost-effectiveness ratio [133,142,50]. Modern ball bearings produce less friction losses with easier lubrication flow since they have smaller contact surfaces [157,150]. Increased speed requirements have led to the development of special types of high speed (HS) and hybrid bearings. The HS bearings have a larger number of smaller balls. Frictional behavior improves due to the consequent improvement in contact [(Fig._18)TD$IG]parameters.

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

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Hybrid bearings possess silicon nitride balls (Si3N4) with a lower density of 3.16 g/cm3 and a higher modulus of elasticity of 320,000 N/mm2 (as compared to the usual roller bearing steel 100Cr6 with a density of 7.85 g/cm3 and a modulus of elasticity of 210,000 N/mm2). The modulus of elasticity crucially affects resilience. The higher modulus of elasticity results in increased stiffness at the same preload for hybrid bearings, but also in higher Hertzian stresses and lower acceptable loads. In order to keep the Hertzian stresses in hybrid bearings at a level comparable to that for steel bearings, preloading of the hybrid bearings is reduced. The ceramic material has excellent tribological properties in combination with steel, resulting in lower friction and reduced wear [133,142,50,16,173,151]. It is also possible to improve performance by using special high-nitrided bearing steel (HNS). The much finer microstructure of this stainless steel, combined with its greater 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 the rolling surfaces with hard thin films. The coatings are intended to improve the wear resistance further while reducing the friction coefficient of contact surfaces [40,132]. Despite their many advantages, the rolling bearings also have performance limitations, partly due to the geometry of the angular ball bearing. Radial displacement of the rolling bodies, or radial widening of the inner ring due to centrifugal forces or thermal expansion can lead to a relative axial displacement of the rings in the case of a bearing adjustment with constant preload. In the case of a rigid bearing configuration it will lead to rising internal bearing loads [157,55]. Any reduction in the bearing stiffness due to changes in the contact angle of the bearing will reduce the dynamic stiffness of the spindle, which will in turn reduce the chatter free material removal rates [1,3]. Development of new bearing concepts is aimed at countering the disadvantages of conventional spindle bearings described above. These concepts are based on the notion that axial and radial displacement of the balls (and also the inner ring) can be prevented by an additional rolling contact in the outer bearing raceway [147,172]. Bearing types with this kind of inner geometry are being used experimentally, for example in the aerospace sector (aircraft engines). These applications, however, require different lubrication and rolling contact parameters. Fig. 19 provides an overview of concepts. Apart from the 3-point bearing with two rolling contacts on the outer ring, two variants of the 4-point bearing are shown. In the 3-point bearing, the problems associated with migration of the balls to the apex as the speed of rotation rises are prevented by the double contact on the outer ring. Preloading of the inner ring by appropriate loading of the bearing continues to be necessary. The rigid 4-point contact can be used as a solid bearing. At high speeds, however, the loads at the contact points increase strongly due to elastic and thermal expansion of the spindle, and limit the allowable spindle speed. The elastically loaded bearings with 4-point contacts do not suffer from this drawback. If a change in contact kinematics due to thermal expansion or centrifugal forces occurs the divided halves

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Fig. 19. New kinematic concepts for spindle bearings.

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

of the inner ring will change their axial distance until the new equilibrium is achieved. This bearing is safeguarded against overload by springs. It is also possible to cool the outer ring without endangering the bearing through a build-up of bracing forces and bearing heat by mutual feedback (known as the ‘suicide loop’, Fig. 20). The spring force, in all cases, must be larger than the maximum axial spindle load in the direction in which the rings are pulled apart [147]. 4.1.2. Roller bearing Cylindrical and tapered roller bearings with single or multiple arrangements are most commonly used on machine tool spindles. High-precision cylindrical roller bearings have been used particularly as movable bearings on spindles, but can likewise be mounted to increase radial stiffness in the region of the spindle nose [102]. Cylindrical roller bearings are radially preloaded via adjustment of bearing clearance in the installed state [133,23]. Due to much greater contact area between the rolling elements and the races, cylindrical roller bearings are much stiffer than ball bearings, and are able to carry heavier loads [61,174]. However, the increased contact area produces higher friction and is more difficult to lubricate. As compared to that of balls in angular ball bearings, they rotate only about one direction, which is kinematically favorable [139]. Critical operating conditions occur especially when there are temperature gradients between the inner and outer rings. The outer ring can usually dissipate heat significantly better via housing components than the inner ring via the spindle body and by convection. Thermally induced radial expansion of the inner ring directly changes the preload set during mounting. An increased preload in turn increases heat generation in the bearing. When a threshold preload value is reached, the bearing is no longer able to dissipate the generated heat quickly, especially from the inner ring. The bearing is in a ‘suicide loop’ (Fig. 20) which can result in destruction of the bearing within a few seconds [23,22,163]. A number of different approaches have been adopted to increase the reliability of cylindrical roller bearings [23,122, 118,119,56]. Systematic weakening of rollers, inner and outer rings makes the bearing less sensitive to the changes in radial preloads. Smaller roller-ring contact zones, profiled rolling bodies, and the use of ceramic rollers are used to reduce the friction and hence the excessive generation of heat in the bearings (Fig. 21). By optimizing such design variables, it is possible to increase the operating speeds of the bearings prior to reaching the suicide loop. The reduction in radial stiffness for roller bearings which accompanies modification of the bearing components may be regarded as unproblematic from the viewpoint of statics and dynamics in a spindle-bearing system [23] because its initial stiffness is always higher than that for ball bearings. 4.1.3. Magnetic bearings Spindles operating with electromagnetic bearings cover a wide range of applications at high speeds. Because of the relatively large

[(Fig._21)TD$IG]

[(Fig._2)TD$IG]


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Fig. 22. Preload mechanisms and influence of a temperature difference DT [167]. Fig. 21. Constructive increase of bearing compliance and reduction of friction [167].

air gaps between the rotor and stator of the axial and radial bearings, the friction is negligible, and the bearings can be operated with minimal wear [65]. Because magnetic bearings are actively controlled, stiffness and damping properties can be adjusted as a function of the feasible control loop dynamics [155]. Due to the integral behavior of the controller, the maximum possible static stiffness of a magnetic bearing is higher than that of a comparable rolling contact bearing. However, the maximum load rating is significantly lower than that of rolling bearings. Even at high speeds of rotation, the rotor in the magnetic bearing can be moved to eccentric paths, permitting a certain degree of self-balancing [52]. Due to the high costs of a complex control system and peripherals these spindles are currently used only in special applications. In the research field, for example, magnetic bearings are used as actuator for contactless application of static or dynamic loads to the spindle-bearing system or as additional bearings mounted with conventional spindle bearings to provide an active influence on the bearing of a motor spindle [81,94]. 4.1.4. Fluid bearing Depending on the desired speed of rotation, spindles with hydrostatic or hydrodynamic bearings are frequently used for high precision production tasks in the metalworking sector. By comparison with rolling contact bearings, and given good damping and high stiffness, these have the advantage of achieving minimum radial and axial runout. As a result of increased heating of the fluid, due to internal shearing effects, there is, a limit on the speeds with spindles having larger tool interfaces (e.g. HSK 63; approximately 10.000 min1) [65,124]. Recent developments show the use of water as the hydrostatic bearing fluid to minimize frictional losses and optimize rotational speed performance [43]. Compared to hydrostatic bearings, hydrodynamic applications in machine tool construction have lost much interest. Hydrodynamic bearings are used only where operation conditions are in the purely fluid friction range without speed variations. On slow running spindles or spindles with frequent starts and stops (e.g. tool changes), operation in the mixed friction range causes increased wear and large frictional losses [65]. Aerostatic bearings are employed when much higher rotational speeds need to be achieved. These work on the same principle as liquid-lubricated bearings, but the active medium is gaseous air, with a lower viscosity than that of the liquids by two to three orders of magnitude. In order to realize a high load capacity and stiffness, very small clearances must be used within the bearings. The air fed into the system is blown off to the surrounding environment by balancing the design of the bearing land for the correct flow resistance for stiffness and length for load carrying capacity. Due to the low mass flows and small specific thermal capacity of air, the frictional heat produced by shear forces cannot be dissipated completely at high relative speeds of the bearing components. High-speed spindles with aerostatic bearings

therefore require additional cooling. The viscosity of the air is virtually independent of its temperature. At certain pressures, the compressibility of air causes pneumatic instabilities, meaning that aerostatic bearings with feed pressures between 4 and 10 105 Pa should preferably be operated in the laminar flow range. The low pressures entail relatively low rated loads and stiffness, require much 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 complete spindle-bearing system depend greatly on the chosen preload force and preload mechanism [157,65]. In terms of design, the preload can be realized in different ways, a fundamental distinction being drawn between rigid and elastic behavior. A rigid configuration is the simplest engineering design solution, since the sole need is to fix the outer and inner rings axially, either by direct opposition of the two bearings or by means of matched spacer rings. Depending on the orientation of the bearing, one differentiates between X and O arrangements. In an X arrangement, an axial thermal expansion of the shaft is transformed into an increase in internal loads, and may lead to failure of the bearing. Analogously, in an O arrangement the preload is relieved (Fig. 23). Elastic preloads keep the bearing preload constant, even when thermally induced relative movements take place between the [(Fig._23)TD$IG]

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

790

[(Fig._24)TD$IG]


spindle and the housing. This can be achieved by using disk springs, by distributing coil springs around the circumference, via hydraulic or pneumatic pistons or via high dynamic piezoactuators. When designing an elastic bearing mounting, space must be left for the bearing rings to move axially. The simplest solution is the use of slide bushings, which are inserted in the housing, where they contain the outer rings of the spindle bearings. To perform the function of a movable bearing, it is necessary to provide a minimum radial clearance, whose size depends on the collective load and on the construction of the spindle. If the clearance is lower than this minimum size due to thermal expansion of the sleeve, or if fretting corrosion occurs, sliding may become impossible or moves in a slip-stick fashion, the bearing may be destroyed. Possible solutions are to use non-ferrous metal or to coat the sleeve or the housing bore. The advantages of the bushings are simplicity of design and good heat dissipation from the bearing to the housing. To make the movable bearing with sliding bushing insensitive to radial seizing, a hydraulic counter-pressure bushing can be inserted. Here the housing and bushing are separated from one another by an oil film. Additional pressure is applied to compensate thermally or kinematically induced ball counterforces which could lead to radial seizing. Design effort is increased as opposed to a simple sliding bushing. In order to reduce friction between the elements of the movable bearing unit which move in relation to one another, it is also possible to employ an axially mobile ball-bearing linear bushing. It should be noted, that if the sleeve expands radially due to heating or centrifugal force, high Hertzian stresses may occur in contact with the housing, due to the small diameter of the balls. This can make the balls exceed the allowable stresses and penetrate the surface of the housing, leading to failure of the linear bearing function. Another disadvantage is poor heat dissipation due to the reduced surface contact. A completely different design solution is to integrate the movable bearing in a membrane spring bushing, consisting of two nested but radially independent sleeves supported in the housing by radially stiff but axially compliant membrane-spring-type elements. The axial and radial stiffness of the movable bearing can be influenced via the spring characteristic. A disadvantage of this solution is the greater difficulty in dissipating heat loss induced by the bearing [23]. If it is possible to dispense with spindle bearings, the use of cylindrical roller or floating displacement (FD) bearings provides a very simple option to realize movable bearings in terms of engineering design. Thanks to the flat outer or inner ring, an axial compensating movement for thermal expansion of the shaft can take place in the form of spiral rolling directly in the bearing itself. Disadvantages are the complex mounting procedure and high sensitivity to radial seizing [65,23]. 4.1.6. Lubrication The tribology system of a rolling contact bearing is characterized by a heavily loaded rolling contact and, in the normal case, low loaded sliding contacts (rolling body/cage, cage/bearing ring). The main task of the lubricant (interfacial medium) in a rolling contact bearing is to form a lubricating film in the respective contact zones of the rolling bodies, bearing rings and cage, so that contact between surfaces – resulting in friction and wear (DIN 50322) – is reduced [167]. It also serves to reduce corrosion and to dissipate frictional heat [142]. In principle, methods are available for lubricating the main spindle of a machine tool, depending on the range of speeds involved: grease lubrication and grease relubrication, oil–air lubrication and oil injection lubrication. Roughly 90% of all rolling contact bearings are operating with grease lubrication [86,17]. In this form of lubrication, the bearings are filled with grease prior to mounting, and it is also referred to as lifetime lubrication. Speed coefficients of up to 2.0 106 mm/min can be achieved by optimizing the chemical composition of the greases and adjusting lubrication to the ceramic ball materials employed in modern high-speed spindle bearings [133]. To

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

improve high-speed performance and service life in grease lubrication systems, relubricating systems, in which fresh grease is introduced via feed lines to points near the bearings or to the bearings themselves, are also commonly used in addition to lifetime lubrication. Oil–air lubrication (Fig. 24) can be used for applications with speed coefficients of up to 3.0 106 mm/min. The principle of lubrication is based on continuous dosing of compressed air mixed with oil to the bearing. Depending on the size of the bearing, quantities of oil lower than 60–200 mm3/h per bearing are sufficient lubrication. Oils with viscosities between 32 and 100 mm2/s can be used and the pressure range is between 2 and 6 bar, depending on the manufacturer’s recommendations and design [133]. In these assemblies, pressure is previously built up in a distribution system by a pump, or static pressure is exerted by the reservoir weight applied to the dosing valves. The valves feed a defined quantity of oil (usually 10 mm3) to a mixing chamber, which is then passed via lubricant feed lines to form a streak of oil in a constantly passing air flow in a narrow tube. This is then fed axially or radially via a feed nozzle to the bearing. In oil–air lubrication, the volumetric flow rate is determined by the cycle time, which establishes the length of time between individual lubricating pulses or valve switching times. If the selected period between 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 and oil streak themselves [111,21]. The lubricant feed to the rolling contact differs depending on the type of bearing concerned. Lubricant can either be provided by a reservoir directly on the bearing itself (capped bearing for lifetime lubrication) or supplied to the vicinity of the bearing by feed lines. Lubricant is fed axially or radially to the bearing through a hole. In the case of spindle bearings, different types of feed are employed, varying in their geometries and surrounding components. 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 which produce a large amount of lost heat. Therefore, in most applications the spindle housing is flown through by a liquid cooling medium in a closed cooling circuit (see Fig. 26, top).

[(Fig._25)TD$IG]

Fig. 25. Lubrication variants.

[(Fig._26)TD$IG]


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

Hence this cooling method is quite complex due to its peripherals (cooling unit, circulating pump). Newly developed solutions like [126] suggest a different consideration with an angular spindle housing. The housings backend is extended and equipped with a separately driven fan. The housing has drains in its corners where an air flow in axial direction is generated caused by the fan. An additional cooling effect is achieved by generating an air flow through the air gap of the motor. 4.2.2. Bearing cooling To minimize the thermal deformation of an aerostatic spindle system, the control of supply air temperature based on the concept of thermal balance has been proposed [179]. The system developed provides the thermal balance between heat generation in bearing clearance and the cooling effect of supply air, therefore, it is possible to eliminate the thermal deformation of the overall aerostatic spindle system including the surrounding parts of the system. 4.2.3. Shaft cooling A newly developed spindle (Fig. 27) is based on an interior cooled shaft [137]. The central component is a rotary union with at least three separate connections. Cooling can be applied through an axial interface and dissipated by a radial hand-over-point. By this means a cooling circuit through the shaft is realized. A substantial advantage of this system is the noticeably shortened time till a steady thermal condition is achieved and through this a reduced heat input from the spindle into the tool is reached. Therefore a thermally induced spindle extension is noticeably reduced [166]. 4.3. Internal coolant supply So-called rotary unions are needed for passing a liquid medium through the rotating shaft to the cutting tool. There are basically contacting and non-contact solutions. On the one hand Frisch [46,45] describes a contactless rotary union for use in motor spindles equipped with an air seal. This system is suitable for a rotational speed up to 60,000 rpm and media-pressure upto 4 MPa

[(Fig._27)TD$IG]

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

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with minimal leakage [46]. He also provides discussions on emerging techniques for rotational speeds up to 90,000 rpm [45]. Sykora [154] points out that many sealing materials utilized in non-contact rotary unions are designed for operating with either a cooling lubricant or minimum quantity lubrication (MQL). Changing the type of lubrication supply may damage the seals. In such a case a solution is provided which allows switching between both cooling systems. MQL has grown in popularity in recent years in an effort to minimize environmental impact. Different works illustrate that the inner supply with pre-mixed aerosol is quite problematic: Schneeweiß et al. [136] as well as Aoyama et al. [13] noticed that an increasing rotational speed leads to demixing of the aerosol. This is mainly the result of the centripetal force acting on the oil droplets which leads to an oil dispersal at the walls of the lubrication circuit [13]. For analyzing the influence of the centripetal force Aoyama et al. [13] also concentrated on the droplets size; the smaller the droplets the lesser the forces. Furthermore the decomposition is less intense. At the same time the lubricating effect is influenced because the droplets poorly stick to the cutting edge. Schneeweiß et al. [136] and Palm and Fuchs [114] identified volatile changes in cross-section, dead spaces and leakage as reasons for a significant oil losses inside the spindle. Aoyama et al. [13] invented a solution with a pivot-mounted pipe inside the spindle. While the spindle rotates the pipe stands still. In this manner the aerosol inside of the pipe is not affected by centripetal forces. A contactless rotary union provides for the transfer of the medium to the rotating tool. Air and oil are led through two separated ducts within the spindle and mixed just before the tool as shown in Fig. 28 [32]. 4.4. Tool clamping and release mechanism In most instances the clamping force is provided by axially stacked disk springs or spiral springs. However these springs are a source of unbalance in operation because the allocation of the masses may vary. One possibility is to use gas-pressurized springs instead of steel springs as shown in [70]. The benefit of this variation is the homogeneous mass distribution. One problem with gas springs is the loss of pressure resulting in a loss of clamping force. Hence it is necessary to monitor the spring load permanently [70]. A system which realizes the tool clamping with an electrical linear motor instead of a hydraulic system or spring assembly was introduced in [90,103]. In addition to shortened tool change times and increased balance quality of the spindle being achieved, the clamping 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. In addition to high stiffness, adequate damping and the ability of [(Fig._28)TD$IG]transmitting the required forces/torques, these systems should

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

[(Fig._29)TD$IG]

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

4.5.3. Hollow shaft taper, HSK For many years the HSK interface can be considered as standard in HSC milling. The static and dynamic characteristics of the HSK tool interface have been analyzed by many institutions and researchers. The basic performance of the HSK shank, such as positioning accuracy, stiffness and bending load capacity, were analyzed in Germany by WZL [171]. Later Aoyama and Inasaki [12] showed that in worst case of diameter tolerances the taper surface of a HSK A 50 tool holder at spindle speeds over 24,000 rpm can lose contact and so the radial stiffness decreases. It is proposed to increase the taper oversize as well as the pull-in-force [12]. Hanna et al. [53] also observed that the deviation in dimension of the tapers is of great importance for the characteristics of the HSK interface. They also show that an increased axial pull force positively affects the rigidity of the system as well as the ability to transfer the required torque. The authors in [70] have investigated theoretical FE analysis and measurements to show the dependencies of the mechanical stresses due to the rotating speed. Strain gauges were applied to the clamping assembly and analyzed via a contactless telemetric system. Fig. 31 shows results that stress increases with higher speeds and demonstrates an increased pull-in-force.

allow a quick, accurate and reliable tool changing process. Very comprehensive overviews of spindle/tool interfaces are given in a CIRP 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 minor contact area between the tapers of the tool holder and the spindle. Compared to HSK tool holders (HSK = Hohlschaftkegel) there is a noticeably lower rigidity. A commercially available steep taper is called ‘Big-Plus’. These tool holders eliminate the mentioned disadvantages of SK tool holders through different optimizations of the taper and the tool clamping. An additional flange contact surface is added as shown in Fig. 30(top). On the one hand this leads to high repetition accuracy in tool clamping. On the other hand a remarkable stiffness increase is achieved compared to usual steep tapers (see Fig. 30, bottom) [18]. Other references can be found on how the flange contact can be realized with short tapers. One possibility is to shorten the taper and to use a longer retention knob [113,152]. It is also suggested to insert adjusting washers. This may affect the rigidity of the connection but it improves the damping [125]. One further option for setting flange contact as well as taper contact is to use oversized and axially slotted tapers. This allows a taper deformation caused by the pull force till the flange surfaces touch [54,57,143].

4.5.4. Coromant Capto Sandvik Coromant company developed a system called ‘Coromant Capto’, which was originally designed for the modular construction of very long tools. The system is mainly characterized by a very flat taper (cone pitch 1:20) with a polygonal shaped outer contour [130]. The hollow taper is deformed during the drawing into the spindle until the flange surface of the tool holder contacts the spindle. Due to its good symmetrical characteristics, the high stiffness and the high torque transfer, this system is also adequate as an interface between spindle and tool holder [125].

[(Fig._30)TD$IG]

4.5.5. Spindle/machine tool interfaces The development of reconfigurable machine tools needs new concepts for spindle/machine tool interfaces. Abele et al. [6] compiled different interface requirements for reconfigurable machine tools and suggested an interface which is based on a modular concept where – according to stiffness demands – up to 9 coupling mechanisms could be used. With this interface a change of different spindle modules is possible within less than 30 min (Fig. 32). Further requirements for spindle-machine-interfaces are presented by Heisel and Meitzner [58]: All ports (for energy or information) should be included into the interface. All components should 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 even the machine operator himself can perform the exchange and

[(Fig._31)TD$IG]

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

Fig. 31. Measured mechanical stress via strain gauges at the clamping set of a HSK 100 A [70].

[(Fig._32)TD$IG]


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composite material. In addition to the analysis of the mechanical characteristics the optimization of fiber layers is considered. For mounting a tool at the one end of the spindle-shaft and an electric motor on the other end steel flanges are glued on both ends of the spindle. 5. Drive concepts 5.1. Motor design

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

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

The present multi-functional machine tools demand motor spindles suitable for HSC-applications with high rotational speeds and lower torques as well as spindles for heavy roughing with comparatively low rotational speeds and high torque. These requirements are fulfilled with various solutions. It was suggested to connect the shaft of the motor spindle through a shiftable coupling to an additional electric motor as shown in Fig. 34. While the main drive works at high rotational speeds and rather low torques, the second electric motor is switched on at low rotational speeds and high torque. This construction offered different advantages, the functional range (torque, speed) of the spindle enlarges and a reduction of torsional vibrations of the spindle and the tool is achieved [146]. 5.2. Frequency converter

4.6. Light weight design For the utilization in small and fast moving applications like printed circuit board applications, Ban and Lee [14] developed an aerostatic spindle with a shaft consisting of a carbon fiber

[(Fig._3)TD$IG]

Frequency converters are required to convert the constant three-phase supply into variable three-phase supply. Integral motor spindles are usually fed by a three-phase two level inverter with pulse width modulated (PWM) output. A wide field of problems emerges from the inadequate supply through the inverter. Because of the switching operating mode, the output voltages are not purely sinusoidal and contain switching harmonics. These harmonic voltages induce currents which do not contribute to torque formation but solely to an undesirable heating of the various spindle elements. To reduce the harmonic content in the supply voltages and currents three-phase three-level inverters were introduced [93]. Due to additional power semi-conductors these inverters have the ability to apply an additional voltage potential to the motor. This significantly reduces the harmonic content of the supply voltage. Fig. 35 shows the set up of the two different inverters and the achievable voltages and currents [127]. Another possibility to overcome the mentioned problems which arise from inadequate voltage supply is to apply a LC output filter between frequency converter and motor. These second order filters damp the harmonic content in the supply. This results in voltage and current which are very close to the ideal sinusoidal form. In this filter application one must pay attention to the characteristic resonance frequency of the filter is not excited by either the inverter or by the motor, as this could destroy the inverter, the filter or the motor. Hence, a control system for the

[(Fig._34)TD$IG]

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).

Fig. 34. Motor spindle with auxiliary drive [146].

[(Fig._35)TD$IG]

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chatter detection index g. This index gives information about the detection accuracy and the permissible computational efficiency; hence it is appropriate for online implementation. Suitable threshold values of g for turning and milling operations are presented. Kuljanic et al. [91] tested several sensors like rotating dynamometers, accelerometers, acoustic emission and electrical power sensors. The sensors were compared in terms of accuracy and robustness. The best results were achieved with a multisensory system composed of axial force sensor and accelerometers. The research of [34] concluded that microphones, which are placed in the ambience of the milling machine to record the ideal noise emission, achieved the best results compared to other sensors.

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 the resonance frequency of the filter is sufficiently damped [4,5]. The influence of the inverter design concept on the temperature behavior of a single integral motor spindle with a permanent magnet synchronous motor is depicted in Fig. 36. The highest temperatures are achieved with the two-level inverter while the filter 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 several methods have been developed over the years to improve the stability of milling. A comprehensive overview is given in [10,141]. The use of microphones and accelerometers are successfully applied in detection chatter in milling operations [34,35,107]. 6.1.2. Dynamometers For measuring the cutting force during the milling process two variations 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 located between the tool clamping and the milling spindle [134,91]. First the sensor signals are analyzed for determining the characteristics of the sensor signals in stable and unstable cases. Secondly chatter indicators and the critical values are composed and finally the characteristics are compared. 6.1.3. Accelerometers According to [31,34,91] the acceleration based chatter detection is carried out with accelerometers. The characteristics of acceleration signals are similar to cutting force signals: The signals are periodic in stable and non-periodic in unstable machining processes. [(Fig._36)TD$IG]Choi and Shin [31] use the measured signals for estimating the

Fig. 36. Stationary temperatures of motor and front bearing with three different frequency inverters (according to [127]).

6.1.4. Displacement meter Another method for measuring and identification of chatter during milling is described in [129] where a laser displacement meter is used in a prototype of a milling system. A laser beam is aimed at the cutting edge and reflected. The intensity and the angle of incidence of the reflected laser beam are recorded and used to interpret the occurring vibrations and to monitor the tool geometry 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 identify chatter during the grinding process. Laser triangulation sensors, pneumatic, radar and waviness sensors are used to measure the geometry of the grinding disk, its wear and the surface [68]. Another possibility for monitoring and diagnosing machining processes is to record the acoustic emissions during the grinding process. A sensor-integrated grinding disk and an acoustic emission (AE) sensor are used for this purpose in [160]. The different AE sensor integrated in the control loop of grinding machines, are shown in Fig. 37. The two automatic identification methods of chatter, which are presented in [51], have the following indicators: Entropy and coarse-grained information rate (CIR). Signals from piezo-electric and acoustic emission sensors, which capture the normal grinding force and the acoustic emissions, are stored for further analyses. During this sequence the entropy is counted from the service spectrum. The CIR is obtained directly from the fluctuation of the recorded signal. 6.3. Preventive detection of bearing damages One of the main problems which suspend the production line is the [(Fig._37)TD$IG] early failure of the spindle bearings. Characteristically this

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

[(Fig._39)TD$IG]


involves high costs for repairs and downtime. Detection and prevention of bearing damage can be taken as precaution. A simple and low cost method for predicting the right time for the replacement of the bearing is to monitor the geometrical damage occurring at the rolling surfaces of the ball bearings [66]. Sinking into and out of a cavity creates an acceleration signal which is analogous to a pulsed sine wave shown in Fig. 38(top). The vibration signals emitted by the rolling motion of the bearing elements (outer and inner rings, cage and balls) are analyzed. A piezo-electric accelerometer is attached at the spindle housing to sense the mentioned acceleration signal. The construction of the test bench is shown in Fig. 38(bottom). After converting the signal it is monitored by a failure prediction processor. This processor transmits a failure prediction alarm when a reference value is exceeded. With measuring the temperature of the outer bearing race during the acceleration of the spindle Spur and Feil [148] assess the actual status and the aging of the bearing as well as the contamination of the lubrication. Damage of the bearing running surface can also be estimated. For detection of axial forces in the fixed bearing a specific force measuring ring was developed by [24,84,134]. This force ring consists of a distance tube equipped with piezo-electric force sensor elements which are circularly adjusted at its front (Fig. 39). This device delivers information about the axial cutting force, the dynamic bearing forces during the operation and the actual preload force during the assembly process. Excessive load on the bearings which causes destruction can be detected. The arrangement can be completed by using a flange sensor ring for measuring the forces in three directions. The project ‘Intelligent Spindle Unit (ISPI)’ [169] links sensors for [(Fig._38)TD$IG] bearing cooling, stator temperature, spindle rotational speed

[(Fig._40)TD$IG]

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Fig. 39. Force measuring ring based on piezo-electric force sensors [134].

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

and outer load conditions. A CAD-model of the spindle is shown in Fig. 40. A microcontroller is used to evaluate the sensor data. Another application is presented in [109]. Sensors for detecting bearing temperature, vibrations and axial displacement of the spindle shaft and tool change control are included. 6.4. Collisions and tool fractures

[(Fig._41)TD$IG]

Fig. 38. Top: Acceleration signal (R: amplitude, L: time duration, P: time period). Bottom: Bearing test bench with variable axial load (according to [66]).

Collisions, tool fractures or overload at high-speed applications can cause serious damage in today’s high-speed machining centers with 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] a piezo-electric force ring is integrated into the spindle housing as shown in Fig. 41 to capture the force signals with minimum time

Fig. 41. Installation of the force ring within the spindle [85].

[(Fig._43)TD$IG]


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lag. Typically, the structural dynamic modes of the spindle reduce the measurement bandwidth of the force rings. Altintas et al. developed a Kalman filter to compensate the structural modes, and increased the measurement bandwidth of the spindle integrated sensors significantly [117,9]. As a result, the dynamometers and spindle integrated force sensors can be used to measure milling forces at high rotational speeds. For the detection of a broken tool an acoustic emission sensor is added to the machine table. A peak in the amplitude of the recorded signal shows the breakage of the tool. In this context a force ring is compared with a table dynamometer in [79]. The ring is composed of eight uniaxial piezo-electric force sensors contained in a circular frame using a special epoxy-material to measure the axial and radial force components. The result is that the force ring is as good as a table dynamometer. Its use is recommended if a table dynamometer is not applicable due to the workpiece size or its geometry. In [24] two piezo-electric force sensing rings (bearing sensor ring and flange sensor ring) are developed and integrated into a direct driven motor spindle for the online process monitoring of machining processes. With this method it is possible to detect the process forces during drilling with tool diameter of less than 4 mm. A spindle integrated data logging unit can be utilized to identify and to avoid spindle damage quickly [38]. A thumb-sized data logger is permanently integrated into the spindle. The unit records the signals of acceleration, temperature and tool change sensors. Connecting the device through a serial port to a computer allows the reading and the parameterization of the logging unit. This simple and safe measuring device enables the detection of spindle irregularities for avoiding serious damages. Operational hours are recorded in relation to the rotation speed which is useful for the prediction of the maintenance interval.

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

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. For determining the thermal growth of a spindle it is common practice to measure the bearing temperature and to perform the needed correlations and corrections. The thermal expansion is just part of the problem. For operations with high-precision spindles, it is also necessary to include the shift caused by speed dependent motion of the bearing to predict more precisely the displacement. The displacement measurement device DMD, developed by Fischer [44], allows for measuring the displacement of the spindle relative to the housing. Attaching a special sensor into the tool flange enables detection of displacement within 1 mm accuracy. The measured deviation is subsequently automatically transmitted to the CNC.

[(Fig._42)TD$IG]

6.5. Axial displacement The spindle’s axial displacement consists of a speed dependent axial displacement of the bearing races and thermal expansion of the spindle. Fig. 17 illustrates these concepts. An overview of measuring axial displacements is given in Fig. 42. In [30] Chen and Hsu characterized the thermal growth of a HSC spindle. Compared to a conventional spindle, the HSC spindles are subject to complex dynamic and speed-dependent influences. A thermal error model is developed to show the mechanical growth and the scheme of Fig. 43 is used. Six sensors are attached to the housing near the bearings and the cooling, and to detect the thermal expansion of the cutting tool and the front-end-cover. Another procedure for improving the machine accuracy by means of temperature control is described in [138], where temperature and flow rate of the cooling are monitored and controlled simultaneously.

Fig. 42. Three methods for spindle’s axial displacement compensation.

7. Mechatronic concepts 7.1. Active balancing High-speed spindles used for grinding or milling demand a higher degree of balance of the spindle system, especially, tool holders and tools. The higher quality and precision can be observed in workpieces and a longer life of the machines components is common. For achieving these goals it is necessary to lower the vibrations caused by unbalance. In addition to the ‘classical’ method of balancing by removing or adding mass in two planes, spindle integrated balancing systems have been developed [110,106]. For reducing vibrations during the operation of the spindle, an active balancing program using influence coefficient method is used to calculate the optimal position of the correction masses and an active balancing device can be used [106]. Fig. 44(bottom) shows the schematic of the device. The active balancing program controls the device by measuring vibration magnitude, phase angle, rotation speed, the position of the balancing rotor and calculates the adjustments to set the correction masses appropriately for balancing the spindle during the operation. Without power the balancing rotor retains its position and circulates with the rotation axis. The pole plate rotates along with the rotation axis and has a magnetized surface. It serves as a pathway for the magnetic flux which is created by the coil flow. For moving the attached unbalance mass the stator is supplied with a current. The operating principle of the device is shown in Fig. 44(top). Exciting current on the driver coil increases the upper-side magnetic flux and decreases the lower-sides flux as well. The rotor is impacted by the downward force normal to the magnetic flux (a). Positioning the permanent magnet at center of the pole plate leads to the greatest density of the magnetic flux and the smallest magnetic resistance (b). Removing the current temporarily brings the rotor into the next step due to the inertia (c).

[(Fig._4)TD$IG]

[(Fig._45)TD$IG]


797

Fig. 44. Top: Principle of the rotor mechanism for active balancing. Bottom: Schematic representation of the active balancing device [106].

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

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$IG]

7.2. Active preload control The preload has a wide influence on bearing life but also on the dynamic runout and thus also impacts surface finish quality. The objective of some research works was to define an ideal preload for the spindle with an active preload-control, and to drive the system near an optimal working point. Prestressing spindles through passive mechanisms is the most conventional method to preload bearings. The issue is that these spindles cannot be adapted to changing conditions. To solve this problem a test spindle was developed in [36] using a novel piezoelectric based actuator module. By using a spindle with an active preload-control as shown in Fig. 46(bottom) the optimal preload for different operating conditions can be determined. A controlled mechanism for preload adjustment has been developed to achieve the preload values suggested by the manufacturer. It consists of a piezo-electric based actuator module with integrated miniaturized hydraulic transmission as shown in Fig. 46(top). In combination with force sensors a continuously controlled preload shall be obtained [36,159,158]. The actuator pushes against a membrane piston which displaces a small part of the hydraulic fluid. The fluid pressure is exerted on metal bellow piston which moves the push rod [36]. Temperature sensors are also integrated for indirectly detecting the heat generation in the bearings. To compensate the thermal deflections of the preloaded spindle the actuator module generates a correlated force. Within the joint research project ‘ISPI’ (see Fig. 40) a piezoactuator is used to achieve a constant preload force. This mechatronic concept allows adjusting the bearing preload to

Fig. 46. Top: Design of the actuator module. Bottom: Test spindle (schematic view) [36].

[(Fig._47)TD$IG]

798

[(Fig._48)TD$IG]


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

establish an optimal working point if the spindle is overloaded [169]. Fig. 47 shows the measured preload force at different rotational speeds with and without active preload control. In [164] a preload model is developed to describe how external cooling (or heating) affects the bearing preload. Based on the model, an active thermal preload regulation scheme is proposed and its feasibility is validated experimentally. The proposed preload regulation scheme is achieved by circulating a cooling (or heating) flow around the spindle housing to manipulate the housing and the outer ring temperatures. 7.3. Active tool deflection compensation Increasing demands on productivity and consequently on increasing feed rates can cause undesirable tool deflections. Machining operations with long slender tools are especially affected. A mechanism for a multiaxial positioning of a HSC spindle and to correct tool deflections is invented in [176]. The combination of a parallel kinematic basic drive arrangement and piezo-ceramic actuators was developed. Complementary precision positioning and additional movements like tilting the tool for deflection compensation are possible benefits of this concept. Denkena et al. [37] also use piezo-actuators as active elements for tool deflection compensation. The experimental setup consists of an adaptronic spindle-system which operates with three prestressed piezo-actuators. The schematic setup is displayed in Fig. 48. Process forces in the x- and y-direction cause the tool deflection (Fig. 48, bottom). The static part of these forces is evaluated by averaging the process forces measured with a dynamometer. Combined with the stiffness of the tool, the deflection can be determined. Using the inverse kinematics the required position is converted into the needed actuator positions to achieve the desired tool tip position. The actual position is steadily measured and corrected by the joint control of the piezoactuators.

[(Fig._49)TD$IG]

Fig. 48. Active tool deflection compensation. Top: Prototype. Bottom: Tool deflection (exaggerated demonstration) [37].

the research, Abele et al. [2] developed a model for the identification of the systems dynamics during machining. Ries et al. [124] introduced a prototype of an active milling spindle. They integrated additional sensors and piezo-ceramic stack actuators for the induction of forces into a common milling spindle. Two piezo-ceramic stack actuators working perpendicular to each other are attached to the outer bearing ring of the front bearing. With this construction it is possible to apply radial loads in the range of 1 kN. The actuators are powered by two highvoltage power amplifiers. A modification of the front bearing allows the generated radial movement. The prototype and the machine tool are shown in Fig. 50. An appropriate control scheme is used for driving the actuators in a way that additional damping is provided. A similar work was presented previously in 1998 by Shankar et al. [140] discussing a ‘Smart Spindle Unit’ for active chatter

7.4. Vibration damping/chatter control The occurrence of chatter is the consequence of an unstable cutting operation. Chatter in machine tools can lead to poor surface finish, high loads and damage to spindles, tools and workpieces. For suppressing these unwanted vibrations three basic approaches are introduced: Active, semi-active and passive control of the process. 7.4.1. Active compensation Within the research project ‘AdHyMo’ Abele et al. [81] developed a hybrid bearing motor spindle (Fig. 49). An active magnetic bearing (AMB) is integrated in a HSC spindle in addition to the conventional ball bearings. The AMB is controlled to increase the damping of the spindle. An increase of 50% in MRR was reached in a laboratory set up. Robust feedback and adaptive feed forward control using m-synthesis improves the process stability. Based on

Fig. 49. AdHyMo spindle [81].

[(Fig._50)TD$IG]

[(Fig._52)TD$IG]


799

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

suppression. This system utilized a state-space model for active control and demonstrated 100% increase in axial depth of cut. The authors in [39] also use piezo actuators to control vibrations during milling. In contrast to the concept of Ries et al. [124] the feedback for the control is the measured strain at the tool root as shown in Fig. 51 Some research was done in the field of spindle speed regulation for stabilization of the cutting process [101,144]. The dynamic interaction of a spindle-tool set and a thin-walled workpiece was analyzed by a finite element approach for the purpose of stability [(Fig._51)TD$IG]

Fig. 52. Designed damper [98].

prediction by [101]. The proposed approach indicates that spindle speed regulation is a necessary constraint to guarantee optimum stability during machining of thin-walled structures. The theoretical basics for the elimination of chatter in milling through the automatic regulation of the spindle speed are presented in [144]. The system described here does not require knowledge of the system dynamics, and it selects stable speeds where no chatter will occur based on current dynamics. 7.4.2. Semi-active compensation A semi-active method utilizing an intelligent material: Electrorheological (ER) fluid is explored in [98]. The medium is nonconducting oil which contains dielectric particles. The fluid can immediately convert its phase from liquid to solid upon exposure to an electric field. A compact damper which contains the ER fluid was designed (as shown in Fig. 52) and mathematical models were developed. Furthermore a semi-active artificial intelligence (AI) feedback controller was established. 7.4.3. Passive compensation Placing the rolling bearings in an additional non-rotating hydrostatic configuration is the procedure used in [168]. This leads to an optimization of the damping characteristics of the spindle, which improves the dynamic behavior of the spindle and tool system. 8. Conclusions and further potentials

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

The heart of every modern machine tool is the main spindle unit which is often designed as a motorized spindle. This core component contributes heavily to productivity, precision and quality of the machined products. Numerous projects have been carried out and reviewed here which had the goal to increase the performance, productivity and reliability of spindles. Through these numerous research activities the topics mechanical/thermal modeling, bearing and drives have been summarized. Also many concepts were developed to integrate additional sensors and actuators in the spindle unit. In general, it appears the state-of-the-art is advancing in the recent past. Especially apparent advances in modeling the dynamic behavior which indicates the potential to more accurately predict the real world.

800


Further progress in machine tools in future will be directly related to the spindle technology. The market demands are still pulling many requirements: Enhancing torque and speed for multi-functional technology applications (grinding, milling and drilling in one spindle) Minimization of energy consumption including the peripheral equipment for drive, bearing and cooling Hard to cut materials are demanding better solutions, e.g. cutting titanium with gearless spindles. Derived from these demands are the challenges for the future research. Exemplary challenges in rolling element bearings are to realize Dmn values of 3 million mm/min with lifetime grease lubrication. Also system design has to cover spindle units which implement multi technologies resulting in high torque and high speed in one system. Further development will be required to allow sensor actuator integration to use the spindle unit as an inherent quality insuring system. The prediction of thermal expansion of spindles and associated changes in the structural dynamic behavior during high-speed machining operations have yet to be solved with satisfactory accuracy. Acknowledgements The authors would like to thank all CIRP colleagues and experts from industry who have given input to this keynote paper. A special thank to following colleagues who have contributed with detailed information about their work related to spindle technology: Budak E., Cao H., Denkena B., Heisel U., Holkup T., Jedrzejewski J., Kolar P., Neugebauer R., Shinno H., Uriarte L., Wertheim R., Winfough W.R., Yamazaki K. This paper would not have been realized without the dedicated effort of Dipl.-Ing. A. Schiffler and Dipl.-Ing. S. Rothenbu¨cher.

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, Rothenbu¨cher S (2007) System Identification During Milling via Active Magnetic Bearing. Production Engineering 1(2):219–226. [3] Abele E, Kreis M, Roth M (2006) Non-Contact Identification of Spindle-Tool Frequency Response Functions. Production Engineering 13(2):207–210. [4] Abele E, Rothenbu¨cher S, Schiffler A (2007) Digitally Controlled Output Filters for Improvement of Temperature Behavior of Machine Tool Main Spindle Drives. Production Engineering 1(2007):219–226. [5] Abele E, Rothenbu¨cher S, Schiffler A (2009) Ausgangsfilter fu¨r SynchronHauptspindelantriebe - Theoretische Grundlagen fu¨r ein neuartiges Regelungskonzept. wt Werkstattstechnik online 99/5:324–330. [6] Abele E, Wo¨rn A, Fleischer J, Wieser J, Martin P, Klo¨pper R (2007) Mechanical Module Interfaces for Reconfigurable Machine Tools. Production Engineering 1:421–428. [7] Altintas Y, Cao Y (2005) Virtual Design and Optimization of Machine Tool Spindles. CIRP Annals Manufacturing Technology 54(1):379–382. [8] Altintas Y, Chan PK (1992) In-process Detection and Suppression of Chatter in Milling. International Journal of Machine Tools and Manufacture 32(3):329– 347. [9] Altintas Y, Park S (2004) Dynamic Compensation of Spindle-Integrated Force Sensors. 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 Angle Calculation in Angular Contact Ball Bearing. Journal of Mechanical Design 128(2):468–478. [12] Aoyama T, Inasaki I (2001) Performance of HSK Tool Interfaces under High Rotational Speed. CIRP Annals Manufacturing Technology 50(1):281–284. [13] Aoyama T, et al, (2008) Development of a New Lean Lubrication System for Near Dry Machining Process. CIRP Annals Manufacturing Technology 57(1):125–128. [14] Ban KG, Lee DG (2002) Design of Carbon Fiber Composite Shafts for High Speed Air Spindles. Composite Structures 55(2):247–259. [15] Bartz W (1993) Luftlagerung Grundlagen und Anwendungen. Expert, Ehningen. [16] Bayer O (1997) Keramik im Wa¨lzlager – ein wirtschaftlicher Weg. Tagungsband ‘‘Innovationen fu¨r Gleitlager, Wa¨lzlager. Dichtungen und Fu¨hrungen’’ (Reihe: VDI Berichte) 1331:61-69. [17] Berger T (2004) Ermittlung der theoretischen Fettgebrauchsdauer fu¨r Langzeit- oder lebensdauergeschmierte Wa¨lzlager. 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 Motorized Spindles—Heat Generation Characterization. Journal of Manufacturing Science and Engineering 123(3):494–505. [20] Brecher C, Esser M, Queins M, Voll H (2005) Motorspindeln sind Schlu¨sselkomponenten. WB - Werkstatt und Betrieb 138(5):45–49. ¨ berwachte Schmierung, Hochgeschwindigkeits[21] Bruenner C, Spieß G (2005) U ¨ l-Luft-Schmiersystem. Antriebstechnik 9:340. spindeln mit integriertem O [22] Burton RA, Staph HE (1967) Thermally Activated Seizure of Angular Contact Bearings. ASLE Transactions 10:408–417. [23] Butz F (2007) Lo¨sungen zur Gestaltung der Loslagerung von Werkzeugmaschinenspindeln. Dissertation, RWTH Aachen. [24] Byrne G, O’Donnell G (2007) An Integrated Force Sensor Solution for Process Monitoring of Drilling Operations. CIRP Annals Manufacturing Technology 56(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 Tool Systems for Virtual Simulation of Milling Operations. International Journal of Machine Tools and Manufacture 47(9):1342–1350. [27] Cao Y, Altintas Y (2004) A General Method for the Modeling of SpindleBearing Systems. ASME Journal of Mechanical Design 126:1089–1104. [28] Carmichael G, Davies P (1970) Factors Which Affect the Transient Behavior of Preloaded Ball Bearing Assemblies. ASLE Transactions 15:1–7. [29] Carmichael G, Davies P (1970) Measurement of Thermally Induced Preloads in Bearings. Strain 6:162–165. [30] Chen JS, Hsu WY (2003) Characterizations and Models for the Thermal Growth of a Motorized High Speed Spindle. International Journal of Machine Tools and Manufacture 43(11):1163–1170. [31] Choi TJ, Shin YC (2003) On-line Chatter Detection Using Wavelet-Based Parameter Estimation. Journal of Manufacturing Science and Engineering 125:21–28. [32] CyTec. Internetpra¨senz. http://www.cytecsystems.de/cytecsystems/german/ produkte/cyspeed/cyspeed_kuehlung.htm. [33] de Mul J, Vree J, Maas D (1989) Equilibrium and Associated Load Distribution in Ball and Roller Bearings Loaded in Five Degrees of Freedom While Neglecting Friction-Part 1: General Theory and Application to Ball Bearings. Journal of Tribology 111:142–148. [34] Delio T, Tlusty J, Smith S (1992) Use of Audio Signals for Chatter Detection and Control. Journal of Engineering for Industry 114(2):146–157. [35] Delio T (1992) Method of Controlling Chatter in Machine Tool, US Patent 5170358, issued 08.12.1992. [36] Denkena B, Go¨tz T (2004) Effects of a Preload Control Unit on Machine Tool Spindles. Production Engineering 11(2):175–178. [37] Denkena B, Mo¨hring HC, Will J (2007) Tool Deflection Compensation with an Adaptronic Milling Spindle. International Conference on Smart Machining Systems (ICSMS), 13.03-15.03, Gaithersburg, USA, . [38] Dittel W (2006) Logger in der Spindel - Condition Monitoring gegen Spindelscha¨den. Werkstatt und Betrieb 139(9):109. [39] Dohner JL, Lauffer JP, Hinnerichs TD, Shankar N, Regelbrugge M, Kwan CM, Xu R, Winterbauer B, Bridger K (2004) Mitigation of Chatter Instabilities in Milling by Active Structural Control. Journal of Sound and Vibration 269(1– 2):197–211. ¨ ber die [40] Erkens G, Rasa T, Mu¨ller J, Brecher C, Spachtholz G (2006) U Vorteile gepulster Plasmen zur Abscheidung innovativer tribologischer Hochleistungsschichten auf Lagerkomponenten. Galvanotechnik 97:1224– 1235. ¨ zgu¨ven H (2007) Selection of Design and Operational [41] Ertu¨rk A, Budak E, O Parameters in Spindle-Holder-Tool Assemblies for Maximum Chatter Stability by Using a New Analytical Model. International Journal of Machine Tools and Manufacture 47(9):1401–1409. ¨ zgu¨ven H, Budak E (2006) Analytical Modeling of Spindle-Tool [42] Ertu¨rk A, O Dynamics on Machine Tools Using Timoshenko Beam Model And Receptance Coupling for the Prediction of Tool Point FRF. International Journal of Machine Tools and Manufacture 46(15):1901–1912. [43] Fecht N (2005) Spurtstarke Spindeln. Antriebspraxis 4:20–22. [44] Fischer. (2000) Alles fu¨r die Spindel. Werkstatt und Betrieb HSC Sonderteil 133(9):34. [45] Frisch T (2001) Ku¨hlmitteleinsatz beim HSC-Bearbeiten - SchnellfrequenzMotorspindeln mit Drehdurchfu¨hrung. Schweizer Maschinenmarkt ;(9)21–23. [46] Frisch T (2001) Minimierter Ku¨hlmitteleinsatz 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 Engineering 129(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) Hochpra¨zisionslager. Katalog-Nr. 40000805, Firmenschrift. [51] Gradisek J, Baus A, Govekar E, Klocke F, Grabec I (2003) Automatic Chatter Detection in Grinding. International Journal of Machine Tools and Manufacture 43(14):1397–1403. [52] Großmann K, Reinschke K, Mu¨hl A, Rudolph J, Rudolph H, Lo¨ser M, Eckhardt S (2006) Auslegung aktiv magnetisch gelagerter Werkzeugmaschinen-Hauptspindeln – Entwurfssystem zur Simulation von Spindeln und Zerspanprozess. ZWF – Zeitschrift fu¨r wirtschaftlichen Fabrikbetrieb 101(9):486–491. [53] Hanna IM, Agapiou JS, Stephenson DA (2002) Modeling the HSK ToolholderSpindle Interface. Journal of Manufacturing Science and Engineering 124(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.

E. Abele et al. / CIRP Annals - Manufacturing Technology 59 (2010) 781–802 [56] Hat Y, Qin Y, Baelendra R, Jiang Q (2004) FE Analysis of a Novel Roller Form: a Deep endcavity roller for roller type bearings. Journal of Materials Processing Technology 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 Integrated Device for Fast Spindle Alignment. Production Engineering 3(3):313– 321. [60] Hernot X, Sartor M, Guillot J (2000) Calculation of the Stiffness Matrix of Angular Contact Ball Bearings by Using the Analytical Approach. ASME Journal of Mechanical Design 122:83–89. ¨ ber die Beru¨hrung fester elastischer Ko¨rper. Journal fu¨r reine [61] Hertz H (1882) U und angewandte Mathematik 92:156–171. [62] Holkup T (2007) Komplexni teplotne mechanicky model vysokorychlostnich sestav valivych lozisek (Complex thermo-mechanical modelling of high-speed rolling 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 Manufacturing Machines 6(3):48–61. [64] Holkup T, Kolar P (2008) Influence of Thermo-Mechanical Conditions of Rolling Bearings on the Dynamics of Machine Tool Spindles. MATAR PRAHA 2008? Proceedings of Part 1, 141–146. [65] Holz M (2006) European Titanium Market – Current and Future Szenario. ITA 22nd Annual Conference & Exhibition, San Diege, 1–3 Oktober 2006/Thyssen Krupp, . [66] Hoshi T (2006) Damage Monitoring of Ball Bearing. CIRP Annals Manufacturing Technology 57(1):427–430. [67] Howes T (1996) The High-Speed Aluminium Milling Challenge. Cutting Tool Engineering 48(7):64–70. [68] Inasaki I, Karpuschewski B, Lee HS (2001) Grinding Chatter – Origin and Suppression. CIRP Annals Manufacturing Technology 50:515–534. [69] Incropera FP, Dewitt DP, Bergman TL, Lavine AS (2007) Fundamentals of Heat And Mass Transfer. John Willey and sons, New York. [70] Jakob F, Sykora H (2008) Special Tool Clamping Requirements in the Field of High 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 of High Speed Machining Centre Headstock. CIRP Annals Manufacturing Technology 53(1):285–288. [72] Jedrzejewski J, Kwasny W, Modrzycki W (2008) Limiting Precision Distortions In Spindle Unit of HSC Machining Centre. Journal of Machine Engineering 8(1):44–53. [73] Jedrzejewski J, Kwasny W, Modrzycki W (2008) Precise model of HSC Machining Centre For Aerospace Parts Machining. Journal of Machine Engineering 8(3):29–41. [74] Jedrzejewski J, Kwasny W, Potrykus J (1989) Beurteilung der Berechnungsmethoden fu¨r die Bestimmung der Energieverluste in Wa¨lzlagern. Schmierungstechnik 8:243–244. [75] Jedrzejewski J, Modrzycki W, Kowal Z, Kwasny W, Winiarski Z (2007) Precise Modelling of HSC Machine Tool Thermal Behaviour. Journal of Achievements in Materials and Manufacturing Engineering 24:245–252. [76] Jones AB (1960) A General Theory for Elastically Restrained Ball and Radial Roller Bearings Under Arbitrary Load and Speed Conditions. ASME Journal of Basic Engineering 82:309–320. [77] Jorgensen BR, Shin YC (1997) Dynamics of Machine Tool Spindle/Bearing Systems Under Thermal Growth. Journal of Tribology 119(4):875–882. [78] Jorgensen BR, Shin YC (1998) Dynamics of Spindle-Bearing Systems at High Speeds Including Cutting Load Effects. Journal of Manufacturing Science and Engineering 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 Fault Diagnosis of Machining Operations. International Journal of Machine Tools and Manufacture 42(6):741–751. [80] Kalveram M (2005) Analyse und Vorhersage der Rotordynamik und Prozessstabilita¨t beim Hochgeschwindigkeitsfra¨sen. Ph.D. thesis, Universita¨t Dortmund. [81] Kern S, Schiffler A, Nordmann R, Abele E (2008) Modelling and Active Damping of a Motor Spindle with Speed-Dependent Dynamics. Vibration in Rotating Machinery, International conference, vol. 1, 465–475. [82] Kim SM, Lee KJ, Lee SK (2002) Effect of Bearing Support Structure on the HighSpeed Spindle Bearing Compliance. International Journal of Machine Tools and Manufacture 42(3):365–373. [83] Kim SM, Lee SK (2001) Prediction of Thermo-Elastic Behavior in a SpindleBearing System Considering Bearing Surroundings. International Journal of Machine Tools and Manufacture 41(6):809–831. [84] Kirchheim A, Schaffner G, Jeck N (2000) Zerspankraftu¨berwachung mit spindelintegrierten Kraftsensoren. Kistler Instrumente AG Winterthur, Schweiz. [85] Klocke F, Reuber M (2000) Sicheres Freiformfra¨sen 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 Linearmotors. http://www.maschinenmarkt.vogel.de/themenkanaele/produktion/ spanendefertigung/spannmittelundsysteme/articles/97570/. [91] Kuljanic E, Sortino M, Totis G (2008) Multisensor Approaches for Chatter Detection in Milling. Journal of Sound and Vibration 321(4–5):672– 693.

801

[92] Lacey S, Wardle F, Poon S (1983) High Speed Bearing for CNC Machine Tool Spindles. Chartered Mechanical Engineering 30:51–56. [93] LEViTEC. Internetpra¨senz. 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 a Piecewise mu-Synthesis Controller. 6th Symposium on Magnetic Suspension Technology, Turin, Italy, October 2001, . [95] Li H, Shin YC (2004) Analysis of Bearing Configuration Effects on High Speed Spindles Using an Integrated Dynamic Thermo-Mechanical Spindle Model. International Journal of Machine Tools and Manufacture 44(4):347–364. [96] Li H, Shin YC (2004) Integrated Dynamic Thermo-Mechanical Modeling of High Speed Spindles, Part 1: Model Development. Journal of Manufacturing Science and Engineering 126(1):148–158. [97] Li H, Shin YC (2004) Integrated Dynamic Thermo-Mechanical Modeling of High Speed Spindles, Part 2: Solution Procedure and Validations. Journal of Manufacturing Science and Engineering 126(1):159–168. [98] Lim S, Park SM, Kim KI (2005) AI Vibration Control of High-Speed Rotor Systems 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-Dynamic Model to Characterize Motorized Machine Tool Spindles During Very High Speed Rotation. International Journal of Machine Tools and Manufacture 43(10):1035–1050. [100] Lin NTLJF (2001) A New Method for the Analysis of Deformation and Load in a Ball Bearing With Variable Contact Angle. Journal of Mechanical Design 123:304–312. ˜ e´ I, Gagnol V, Bouzgarrou B, Ray P (2008) Stability-Based Spindle Speed [101] Man Control During Flexible Workpiece High-Speed Milling. International Journal of Machine Tools and Manufacture 48:184–194. [102] Makino (2005) MAG Series, Five-Axis Horizontal Machining Centers. Katalog: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 and Error Assessment. CRC Press, Boca Raton. (Chapter 3, pp. 75–126). [105] Metzele M (2004) Piezoaktoren regeln die Lagervorspannung. wt - Werkstattstechnik online 94/2004/7-8:344–349. [106] Moon JD, Kim BS, Lee SH (2006) Development of the Active Balancing Device for High-Speed Spindle System Using Influence Coefficients. International Journal of Machine Tools and Manufacture 46(9):978–987. [107] Morgan G, Cheng RQ, Altintas Y, Ridgway K (2007) An Expert Troubleshooting System for the Milling Process. International Journal of Machine Tools and Manufacture 47(9):1417–1425. [108] Namazi M, Altintas Y, Abe T, Rajapakse N (2007) Modeling and Identification of Tool Holder-Spindle Interface Dynamics. International Journal of Machine Tools and Manufacture 47(9):1333–1341. [109] Neuenschwander U (2005) Wie man sich spindelt, so fra¨st man!. Werkstatt und 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 Machining Proceedings, 28. und 29. Mai 2008, Darmstadt, Meisenbach Verlag GmbH, Bamberg, ISBN: , pp. 978-3-87525-274-3179–184. [111] N.N. (2007) MMS Lagerschmierung (u¨berwacht) BAL 30995090. Bedienungsanleitung, Bielomatik Leuze GmbH+Co KG. [112] Ophey L (1985) Da¨mpfungs- und Steifigkeitseigenschaften vorgespannter Schra¨gkugellager/Lothar Ophey. Lehrstuhl fu¨r 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 Materials Handling Engineering. Paper no. IMECE2004-60395 (ASME 2004 International Mechanical Engineering Congress and Exposition (IMECE2004) November 13–19, 2004, Anaheim, California, USA). pp. 41–48. [117] Park SS, Altintas Y (2004) Dynamic Compensation of Spindle Integrated Force Sensors With Kalman Filter. Journal of Dynamic Systems Measurement and Control 126(3):443–452. [118] Poplawski JV, Peters SM, Zaretsky EV (2001) Effect of Roller Profile on Cylindrical Roller Bearing Life Prediction – Part I: Comparison of Bearing Life Theories. Tribology Transactions 44(3):339–350. [119] Poplawski JV, Peters SM, Zaretsky EV (2001) Effect of Roller Profile on Cylindrical Roller Bearing Life Prediction – Part II: Comparison of Roller Profiles. Tribology Transactions 44(3):417–427. [120] Przemieniecki JS (1985) Theory of Matrix Structural Analysis. Courier Dover Publications. [121] Rantatalo M, Aidanpaä¨ JO, Go¨ransson B, Norman P (2007) Milling Machine Spindle Analysis Using FEM and non-Contact Spindle Excitation and Response Measurement. International Journal of Machine Tools and Manufacture 47(7–8):1034–1045. [122] Reusner H (1987) The Logarithmic Roller Profile – The Key to Superior Performance of Cylindrical and Tapered Roller Bearings. Ball Bearing Journal 230:2–10. [123] Richardson MH, Formenti DL (1982) Parameter Estimation From Frequency Response Measurements Using Rational Fraction Polynominals. IMAC Conference, Orlando, FL, . [124] Ries M, Pankoke S, Gebert K (2006) Increase of Material Removal Rate with an Active HSC Milling Spindle. Adaptronic Congress 2006, 3rd–4th May, Go¨ttingen, .

802


[125] Rivin E (2000) Tooling Structure: Interface between Cutting Edge and Machine Tool. CIRP Annals Manufacturing Technology 49(2):591–634. [126] Ronde´ U, Kades H (2007) Motorspindel mit einer Ku¨hleinheit. Patentschrift DE102006043550A1, issued 30.08.2007. [127] Rothenbu¨cher 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 Dynamics Analysis. Journal of Sound and Vibration 66(1):119–140. [129] Ryabov O, Mori K, Kasashima N (1998) Laser Displacement Meter Application for 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 for Stability Analysis of the Variable Speed Face-Milling Process. Journal of Manufacturing Science and Engineering 124(1):1–10. [132] Schaeffler GruppeIndustrie H (2005) Ho¨heres Leistungsvermo¨gen durch Beschichtungen–Beschichtete Wa¨lzlager der Schaeffler Gruppe. Firmenschrift. [133] Schaeffler Gruppe Industrie S (2007) Hochgenauigkeitslager–Super Precision Bearings. Firmenschrift. [134] Scheer C, Hoffmann P (1999) Spindle-Integrated Force Sensors for Monitoring Drilling and Milling Processes. 9th Int. Conf. for Sensors, Transducers & Systems May 18–20, 1999, Nu¨rnberg, Germany, . [135] Schmitz TL, Powell K, Won D, Duncan GS, Sawyer WG, Ziegert JC (2007) Shrink Fit Tool Holder Connection Stiffness/Damping Modeling for Frequency Response Prediction In Milling. International Journal of Machine Tools and Manufacture 47(July (9)):1368–1380. [136] Schneeweiß M, Ko¨hler S, Schwarzburger R, Pospischil T (2004) Ku¨hlschmierung bedarfsgerecht dosieren. WB - Werkstatt und Betrieb 137(11):54–58. [137] Scholl S (2007) Spindelvorrichtung mit Innenku¨hlung. Patentschrift DE102005030277B4, issued 31.10.2007. [138] Schossig HP (2002) Neue Motorspindeln fu¨r die Komplettbearbeitung. WB Werkstatt und Betrieb 135(9):70–74. [139] Schu¨tt P (1978) Bewegungsverha¨ltnisse in leichtbelasteten schnell laufenden Zylinderrollenlagern. Dissertation, TU Braunschweig. [140] Shankar N, Bridger K, Regelbrugge ME, Winfough WR (1998) Smart Spindle Unit For Active Chatter Suppression Of A Milling Machine: I. Overview, Fabrication and Assembly. Smart Structures and Materials 1998 Industrial and Commercial Applications of Smart Structures Technologies 3326(1):160– 166. [141] Shiraishi M (1989) Scope of In-Process Measurement, Monitoring and Control Techniques in Machining Processes – Part 3: In-Process Techniques for Cutting Processes and Machine Tools. Precision Engineering 11(1):39–47. [142] SKF Gruppe, SPWVG (2006) Pra¨zisions-Schra¨gkugellager. Hauptkatalog, Firmenschrift. [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 Engineering 120(4):674–683. [146] Soshi M, Ishiguro H, Yamazaki K (2009) A Study on the Development of a Multi-Purpose Spindle System for Quality Productive Machining. CIRP Annals Manufacturing 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 Thermally Induced Preload in Anti-Friction Spindle Bearings of High-Speed Machine Tools. Journal of Dynamic Systems Measurement and Control 116(3): 372–386. [150] Steinert T (1995) Das Reibmoment von Kugellagern mit bordgefu¨hrtem Ka¨fig. Dissertation, RWTH-Aachen. [151] Sternagel R, Kreuser J (2001) Hybrid- und Keramikwa¨lzlager fu¨r extreme Beanspruchungen. Tribologiecehsp sp025/cehsp sp025/Schmierungstechnik 48(4):11–20. [152] Stoll R (1990) Grenze nicht erreicht - Optimierte Steilkegelspannung hat noch Reserven. MM - Maschinenmarkt Das Industriemagazin 96(38):68–70.

[153] Sud O, Davies P, Halling J (1974) The Thermal Behavior of Rolling Bearing Assemblies subjected to preload. Wear 27:240–246. [154] Sykora H (2004) Diese Systeme bestehen die HSC-Probe. WB - Werkstatt und Betrieb 137(9):56–59. [155] Tamisier V, Font S, Lacour M, Carre`re F, Dumur D (2001) Attenuation of Vibrations due to Unbalance of an Active Magnetic Bearings Milling ElectroSpindle. 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] Tu¨llmann U (1999) Das Verhalten axial verspannter schnelldrehender Schra¨gkugellager. Ph.D. thesis, WZL RWTH Aachen. [158] To¨nshoff H, Denkena B, Go¨tz T (2002) Piezoelectric Actuator Based Preload Control Unit for Machine Tool Spindles. Production Engineering 9(1):117–122. [159] To¨nshoff H, Go¨tz T (2002) Piezoaktoren zur Vorspannungsregelung in Werkzeugspindeln. Antriebstechnik 41(3):45–49. [160] To¨nshoff H, Karpuschewski B, Mandrysch T, Inasaki I (1998) Grinding Process Achievements and their Consequences on Machine Tools Challenges and Opportunities. CIRP Annals Manufacturing Technology 47(2):651–668. (18 pp.). [161] To C (1981) A Linearly Tapered Beam Finite Element Incorporating Shear Deformation And Rotary Inertia For Vibration Analysis. Journal of Sound and Vibration 78(4):475–484. [162] Trojahn W, Zoch H (2000) Entwicklung und Anwendung von Wa¨lzlagern aus Cronidur130. Stahl ;(3)44–46. [163] Tu JF, Katter JG (1996) Bearing Force Monitoring in a Three-Shift Production Environment. Tribology Transactions 39:201–207. [164] Tu JF, Stein JL (1996) Active Thermal Preload Regulation for Machine Tool Spindles with Rolling Element Bearings. Journal of Manufacturing Science and Engineering 118(4):499–505. [165] Young DL, Weaver W, Timoshenko S (1990) Vibration Problems in Engineering. John Wiley & Sons, Incorporated. [166] Walter R (2006) Mit direkter Ku¨hlung zu mehr Genauigkeit. WB - Werkstatt und 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 - Werkstattstechnik online 93/2003/7–8:517–521. [170] Weck M, Reinartz T (2004) Determination of the Load Capacity of HSK Interfaces. Production Engineering 11(1):99–102. [171] Weck M, Schubert I (1994) New Interface Machine/Tool: Hollow Shank. CIRP Annals – Manufacturing Technology 43(1):345–348. [172] Weck M, Spachtholz G (2003) 3- and 4-Contact Point Spindle Bearings – A New Approach For High Speed Spindle Systems. CIRP Annals Manufacturing Technology 52:311–316. [173] Wemho¨ner J (2001) Keramik–ein Meilenstein auf dem Weg zum Hochleistungswa¨lzlager. Tagungsband zum Symposium ‘‘Pulvermetallurgie–Schlu¨ssel zur Effizienzsteigerung’’. Hagen:15–16. [174] Wiche E (1967) Die radiale Federung von Wa¨lzlagern bei beliebiger Lagerluft. Konstruktion 19:184–192. [175] Wiemer AL (1969) Luftlagerung. VEB Verlag Technik, Berlin. [176] Wittstock V, Mu¨nzer J, Neugebauer R (2007) Aktive Lagekorrektur einer HSCHauptspindel. 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 of Materials Processing Technology 138(1–3):379–384. [178] Yokoyama T (1990) Vibrations of a Hanging Timoshenko Beam Under Gravity. Journal of Sound and Vibration 141(2):245–258. [179] Yoshioka H, Matsumara S, Hashizume H, Shinno H (2006) Minimizing Thermal Deformation of Aerostatic Spindle System by Temperature Control of Supply Air. JSME International Journal Series C Mechanical Systems Machine Elements and Manufacturing 49(2):606–611. [180] Zverev IA, Eun IU, Chung W, Lee CM (2003) Simulation of Spindle Units Running on Rolling Bearings. The International Journal of Advanced Manufacturing Technology 21(10):889–895. [181] Zverev IA, Eun IU, Hwang Y, Chung W, Lee CM (2006) An Elastic Deformation Model of High-Speed Spindle Units. International Journal of Precision Engineering and Manufacturing 7:39–46.

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