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Annals of the CIRP Vol. 56/1/2007 -395- doi:10.1016/j.cirp.2007.05.092Developments for High Performance Machine Tool Spindles C. Brecher1(2), G. Spachtholz1, F. Paepenmller11Laboratory for Machine Tools and Production Engineering, Aachen, Germany Abstract One important demand on spindle systems in modern machine tools is to realise higher rotational speeds in order to increase the machining efficiency. Additionally, for a given speed range a better robustness is demanded so that the spindle system is desensitised against improper operating conditions. The paper presents research results in various fields which contribute to the improvement of spindle-bearing systems. At first, new results for alternative spindle bearing kinematics with four contact points are presented. Secondly, a new solution for floating bearing arrangements is discussed. A modified cylindrical roller bearing is presented which can be operated at higher speeds. Finally, the potential of coated bearing components is discussed in the context of improved fail-safe properties. In this paper both analytic studies and experimental tests are presented. Keywords:Machine, Spindle, Bearing 1 INTRODUCTION The productivity of modern machine tools is mainly determined by the rotational speed limits and the load carrying capacities of their main spindle units. On the one hand, the machining of aluminium or magnesium with modern cutting tools, equipped with cubic boron nitride (CBN) or polycrystalline diamond (PCD) inserts, allows cutting rates from 5,000 m/min up to 10,000 m/min. In the case of the application of end mills with diameters between 20 and 30 mm the realisation of these very high cutting speeds requires spindle speeds of more than 100,000 rpm. According to the actual state of the art of rolling bearing technology, this demand can currently only be realised by spindle bearings with a mean diameter of 30 mm. However, due to these extreme operating conditions, all functional components of a main spindle unit the spindle bearings, the rotor of the motor as well as the rotating unions are loaded up to their physical limits. On the other hand, the main spindles also need to be suitable for versatile machine tool applications. These are characterised by varying demands. The rough machining of steel, for example, is characterised by high cutting forces and moments and moderate rotational speeds. In those cases, bigger spindle and bearing diameters are essential to bear these loads. The following basic approaches for the design of main spindles can be derived from the diverging demands presented above. To fulfil those demands, the characteristic speed coefficient n x dm has to be increased up to 3.5 4.0x106 mm/min, ensuring a sufficient stiffness and robustness of the spindle body and the spindle bearings.Figure 1 presents a typical motor spindle with a power output of 80 kW and a maximum rotational speed of 30,000 rpm. The stator of the drive is water-cooled. The spindle body has a hollow shaft taper and is rotationally supported by an elastically preloaded back-to-back spindle bearing arrangement. In order to develop an improved spindle and bearing design the optimisation of the fixed bearing unit (variedelastically preloaded back-to-back arrangementHSK A 63/80bearing diameter: 70 mmmax. rotational speed: 30,000 rpmsource: Weiss GmbHFigure 1: Milling motor spindle with rolling bearings. inner bearing geometry), of the movable bearing unit(elastic cylindrical roller bearings) as well as of the tribological properties (surface coatings, lubrication) shall be analysed. These topics are discussed in the following. 2 MULTIPOINT(3P,4P)-SPINDLE BEARINGS 2.1 Motivation for the Optimisation of the Bearing Geometry Spindle bearings for the application in main spindles of modern machine tools have to fulfil highest demands on running accuracy and stiffness. In the past, various modifications have been developed in order to improve the bearing performance. Among others, one can enumerate special lubricant supplies through the outer ring, smaller or ceramic balls as well as optimised cages. Nevertheless, highest speeds extremely reduce the life time of spindle bearings. The underlying main effects during the operation process were investigated by various authors and were, e.g., summarised in 1, 2, 3, 4. Especially, the speed dependent deviation of the contact angles on the inner and outer ring causes a decrease of axial and radial stiffness. In addition, the contact areas on -396- the outer ring are strongly loaded by the centrifugal forces acting on the balls. Weck et al. presented in 1 spindle bearings with varied inner geometries in order to reduce the negative effects of the centrifugal forces acting on the balls. In addition, the robustness of the bearings should be improved. Weck et al. provided additional contact points on the raceways of the inner and outer ring of conventional spindle bearings. Due to this, the axial and radial movement of the balls is prevented and constant contact angles and a reduced axial displacement of the inner ring can be ensured over a wide speed range. The bearing concepts introduced in 1 are shown in Figure 2 (a), (b). Figure 2 (c) presents a third, novel concept introduced in this paper.(a)Multipoint(3P)-bearing(b)Multipoint(4P)-bearing(c)Multipoint(4P)-bearing, internal spring preload Figure 2: Different alternatives of multipoint(3P,4P)-spindle bearings. 2.2 Analytic Investigation of Multipoint(4P)-Bearings The contents of 1 focus on the operating behaviour of multipoint(3P)-bearings, both theoretic and experimentalinvestigation. The test bearings were manufactured at the Laboratory for Machine Tools and Production Engineering (WZL) based on conventional spindle bearings. Also some characteristics of multipoint(4P)-Bearings were analysed by means of numerical calculations. Subsequently, further results regarding the development of new bearing kinematics with four contact points per ball under consideration of thermal effects will be presented. All calculations are done for the bearing size 7014 with ceramic balls of the diameter 12.7 mm. The contact angles amount to 15. The abbreviations used in the following diagrams are listed in Table 1. in.ring1/2 Inner ring, contact point 1 or 2out.ring3/4 Outer ring, contact point 3 or 4 ET Excess temperature Table 1: Abbreviations used in Figures 3, 4, 5. In 1 it was shown by calculations that the axial displacement of the inner ring can be reduced to less than two microns and that a speed dependent change of the contact angles can be prevented. However, the multipoint(4P)-bearings are mounted with zero radial clearance. Therefore, they are extremely sensitive to thermal effects. Especially excess temperatures of the inner ring may cause jamming of the bearing. This effect also occurs in cylindrical roller bearings under high rotational speeds. Figure 3 illustrates the Hertzian pressures in the contact 1, inner ring, and contact 4, outer ring of a multipoint(4P)-bearing. These contacts directly transfer the axial load (see pictogram) and are stressed to the highest extent. The bearing shows a radial clearance of 22 microns. The fit between inner ring and spindle amounts to 35 microns in order to prevent lifting-off of the inner ring. Therefore, the bearing is slightly preloaded. The curves 1 and 2 represent a bearing calculation without consideration of thermal effects. The higher stress values on the raceway of the inner ring result from a wider curvature. The increase of the Hertzian pressures is caused by the centrifugal expansion of the inner ring as well as by the centrifugal forces acting on the balls. However, the stresses do not exceed critical values of the conventional bearing material 100Cr6 (2,000 N/mm). 00,511,522,5051015202530rotational speed 1,000 rpm03691215excess temperature KHertzian pressure kN/mm 1: in. ring12: out. ring43: in. ring1, ET4: out. ring4, ET1234ETaxial preload3142ax. load 500 NFigure 3: Hertzian pressures in a multipoint(4P)-bearing dependent on speed and excess temperature. In contrast, the curves 3 and 4 show the influence of the inner ring excess temperature (linear increase) in combination with the centrifugal effects. These curves were calculated by assuming a gradient of 1 K per 2,000 rpm. A significant rise of the internal stresses can be noticed. At a maximum rotational speed of 30,000 rpm the Hertzian pressures on the inner ring rise above the limiting value of 2,000 N/mm. In addition to these theoretical results, one has to take into consideration that the increase of internal stresses on the one hand and the thermal overloading on the other are coupled self-energising effects. Therefore, one can expect jamming of a multipoint(4P)-bearing with a pronounced excess temperature of the inner ring in case of higher rotational speeds. This assumption will be investigated in experimental tests in section 2.3.Figure 2 (c) shows a third concept of a new bearing geometry. It was developed in order to prevent the internal overloading of the multipoint(4P)-bearing. The concept is characterised by a divided inner ring. One half of the ring which is oriented towards the tool side of the spindle is fixed to the spindle body. By that, it can bear the forces resulting from the machining process. The second half of the ring is axially movable and pressed against the balls by a disc spring, creating an internal preloading of the bearing.The diagram in Figure 4 illustrates the calculated speed dependent development of kinematics in the preloaded multipoint(4P)-bearing assuming a linear increase of the inner rings excess temperature up to 15 K (see Figure 3). The internal spring preload amounts to 370 N. 0510152025051015202530rotational speed 1,000 rpm325345365385405425spring preload Ncontact angles 1: in.ring1, ET2: in.ring2, ET2a: in.ring2spring preloadAx. Load 370 N2b: in.ring2, ET, 203: out.ring3, ET4: out.ring4, ET2a3, 422b1Figure 4: Contact angles in a multipoint(4P)-bearing with internal spring preload of 370 N. -397- The outer axial load is decreased to 370 N in order to assure the same initial loading of the contact points as in Figure 3. As explained in 1 the outer contact angles (curves 3, 4) stay constant over the whole speed range. Caused by the thermal and centrifugal effects the first half of the inner ring expands (contact 1). Therefore, a slight decrease of the corresponding contact angle (curve 1) and, consequently, a spindle displacement occur. The second movable half of the inner ring is not mounted by press fit to the spindle resulting in a strong centrifugal expansion and a strongly decreasing contact angle (curve 2). Curve 2a shows the reduction of the contact angle without consideration of thermal effects. In addition to the contact angles the spring preload curve is shown. The load increases to 406 N caused by the displacement of the spindle body and the movable ring half.The diagram in Figure 5 shows the Hertzian pressures in the preloaded multipoint(4P)-bearing. Again, curves 1 to 4 represent the loading of contacts 1 (inner ring) and 4 (outer ring) with and without an excess temperature on the inner ring. It becomes obvious, that the stress level is clearly lower than in the rigid multipoint(4P)-bearing variant (see Figure 3). The maximum stress value is 1,600 N/mm. To understand the effect of the decreasing contact angle on the movable half of the inner ring, the Hertzian pressure in contact 2 (movable inner ring) is also analysed. Curve 5 illustrates a maximum value of about 2,000 N/mm. This Hertzian pressure can be reduced by providing a larger contact angle. This effect is shown by the curve 6 in the diagrams in Figures 4 and 5. By increasing the contact angle to 20 a pressure reduction to 1,200 N/mm is possible. The theoretical analysis of the multipoint(4P)-bearing with internal spring preload shows that the elastic arrangement of the two halves of the inner ring can prevent the bearing from jamming. The negative effects which result from inner ring excess temperatures can be reduced. In the following, experimental results regarding the operating behaviour of multipoint(4P)-bearings under the influence of an excess temperature are presented. 00,40,81,21,62051015202530rotational speed 1,000 rpm03691215excess temperature KHertzian pressure kN/mm 1: in.ring12: out.ring43: in.ring1, ETETax. load 370 N4: out.ring4, ET5: in.ring2, ET6: in.ring2, ET, 20124563Figure 5: Hertzian pressure in a multipoint(4P)-bearing with internal spring preload of 370 N. 2.3 Experimental Investigation of Multipoint(4P)-Bearings The test stand used for the experimental investigations is shown in Figure 6. The direct drive can realise maximum rotational speeds of up to 40,000 rpm. The nominal torque amounts to 4.2 Nm for a nominal speed of 23,000 rpm. The test spindle and the drive are connected by a jaw clutch. The test bearings can be loaded axially by a hydraulic piston. A tempering of the outer ring is realised by a water circulation in the flange. Thereby, the heating of the outer ring caused by the additional rolling contact can be reduced. The inner bearing temperature is measured by a non-contacting sensor positioned closely to the rotating inner ring.Direct driveHydraulicunitTemperaturesensorFlange withtest bearingTest spindleFigure 6: Test stand. The diagram in Figure 7 presents experimental results for both a rigid and an elastic multipoint(4P)-bearing.At first, the rigid bearing (concept (b), Figure 2) was tested. The tests were performed with and without tempering of the outer ring. Subsequently, the flexible bearing (concept (c), Figure 2) was tested with tempering of the outer ring. The measured temperatures are shown related to ambient temperature. The torque values are derived from the motor current during the tests. The abbreviations used in Figure 7 are explained in Table 2. it1, ot1, t1 Inner / outer ring excess temperature / torque, concept (b), no tempering it2, ot2, t2 Inner / outer ring excess temperature / torque, concept (b), tempering 40C it3, ot3, t3 Inner / outer ring excess temperature / torque, concept (c), tempering 40C Table 2: Abbreviations used in Figure 7. The curves it1 and ot1 illustrate the inner and outer temperatures of first test bearing, concept (b). The axial load amounts to 1,000 N. The bearing is grease-lubricated with 5 g of KlberSpeed BF72-22. Up to 19,000 rpm the rotational speed is increased by 2,000 rpm every 2 hours, then by 1,000 rpm every 2 hours. The test is stopped if a shut-off temperature of 55K is exceeded.010203040506004812162024283236time h00,20,40,60,811,2friction torque Nmax. load: 1000 Nexcess temperature (rel. to Tambient) K7t11t15t19t21t23t25t27t29tot2it1, ot1it3ot3it2t3t1t2Figure 7: Operational behaviour of multipoint(4P)-bearings.With this first bearing alternative, a maximum rotational speed of 21,000 rpm can be realised. Inner and outer -398- bearing temperatures reach the same values and rise up to the shut-off temperature of 55 K (curves it1, ot1). The torque amounts to 0.25 Nm. In the next step, the same bearing was tested tempering the outer ring. At the tempering unit, a supply temperature of 40C was adjusted. Compared to the first test, up to 17,000 rpm the bearing temperatures are clearly reduced (curves it2, ot2). The temperature of the outer ring is up to 4 K lower than the one of the inner ring. This is caused by the optimised heat dissipation through the water circulation in the flange. The operating state of the bearing corresponds to the calculation in Figure 3. After increasing the speed to the next level (19,000 rpm) a sudden rise of the temperatures and the friction torque occurs. The test is stopped immediately. This effect corroborates the assumption that the bearing concept (b) will fail in case of high exceed temperatures of the inner ring.Finally, the third bearing concept (c) with the elastically preloaded, movable half of the inner ring was examined. The axial load was decreased to 800 N in order to adapt the loading of the rolling contacts to the first tests. Again, the outer ring was tempered by water circulation. The measured temperatures of the concept (c) are in between the results of the two foregoing tests. They are lower than those of test 1 due to tempering and higher than test 2 because of the additional internal spring preload of about 380 N. Again, an excess temperature of the inner ring can be observed. Nevertheless, the bearing reaches the final rotational speed of the test cycle of 30,000 rpm with maximum excess temperatures relative to ambient temperature of 50 K (inner ring) and 43 K (outer ring),respectively. The tests with the bearing concepts (b) and (c) were performed in order to verify the results of the calculations. Although a direct measurement of the internal bearing kinematics is not possible during operation the basic findings regarding the operational behaviour of the multipoint(4P)-bearings could be proven. 3 ELASTIC CYLINDRICAL ROLLER BEARING AS MOVABLE BEARING Main spindle units for highest rotational speeds are usually designed based on an elastic arrangement of angular contact ball bearings. This spindle design is characterised by a fixed and a movable bearing unit in order to compensate thermal and kinematic elongations of the spindle. Exemplary solutions for the floating unit are the so called sliding bush or the ball bush. In case of temperature gradients in the spindle, the bearing rings, the bushes or the housing the axial movement of the bushes may be reduced or even prevented causing a spindle failure. In this context the cylindrical roller bearing can be called an “ideal” movable bearing. A relative axial movement of inner and outer ring is enabled by a helical rotation of the rollers. However, the reachable rotational speeds are limited due to radial jamming caused by thermal and centrifugal effects acting on the bearing components. Therefore, approaches for high speed cylindrical roller bearings have to reduce the internal heat generation based on power losses and increasing Hertzian pressures in the line contacts. In 5 Yangang et al. present deep end-cavity rollers for cylindrical roller bearings reducing the sharp edge-stresses.In current research work at WZL the internal loads of cylindrical roller bearings are reduced similar to the approach of the multipoint(4P)-bearing. Additionally, the internal friction is reduced. Figure 8 shows three different bearing modifications which aim at a reduction of the internal friction (narrow raceway, ceramic rollers) and at a reduction of the speed and heat-dependent Hertzian pressures in the line contacts (elastic bearing rings). (a)Narrow raceway (b)Ceramic rollers (c)Elastic rings Figure 8: Different alternatives of modified cylindrical roller bearings. In the following section test results for the first (a) and the third bearing concept (c) are presented. The cylindrical roller bearings are equipped with crowned steel rollers. The modifications are ground grooves in the outer or inner ring or a narrow inner ring raceway, respectively. By these measures the stiffness of the bearings and the heat generation are reduced. The design of the grooves is developed by means of FE-calculations 5 so that the rings comply equally in case of rising internal loads. All test bearings are mounted with an initial radial clearance of -2 m and are air-oil lubricated (oil viscosity 32 mm/s, air pressure 1.7 bar). During the tests the rotational speed is increased stepwise. Curve 1 in Figure 9 illustrates the operating behaviour of a reference bearing. The temperatures measured at the outer ring clearly rise after passing rotational speeds of 12,000 rpm. The final rotational speed is 15,000 rpm with an absolute outer ring temperature of 70C. Curves 2 to 4 show the test results for the modified bearings. The best temperature behaviour and, in consequence, the highest rotational speeds are realised with the bearing with a narrow inner ring raceway. The maximum speed amounts to 23,000 rpm. The bearings with elastic rings show a significant increase in temperature after passing 16,000 rpm (curve 4) and 18,000 rpm (curve 3). Both reach final speeds of 18,000 rpm. According to the test results all modified bearing types allow for the realisation of higher rotational speeds. The positive effects of the developed optimisations are proven.rotational speed 1000 rpm2824201284000:0009:3619:1228:4838:24time h807060403020100absolute temperature Cclearance: -2 mviskosity: 32 mm/sair pressure: 1.7 bar1: no modification2: narrow raceway3: groove in outer ring4: groove in inner ringrotational speedFigure 9: Test results for modified cylindrical roller bearings (type N1014K). 4 OPTIMISATION OF THE FAIL-SAFE PROPERTIES OF SPINDLE BEARINGS BY A-C:H:W COATINGS Besides the bearing design development, additional research work was done focussing on the optimisation of the fail-safe properties of conventional spindle bearings. -399- Spindle breakdowns are often caused by insufficient lubrication. Especially in case of grease lubrication, an insufficient lubricant supply may result in cage breakage or overheating of the bearing. 010203040506070048121620242832absolute temperature C1: coated, without any lubrication2: uncoated, 1mm grease applicated to the raceways 8t29tcage breakage 20t21time hFigure 10: Hybrid spindle bearing, uncoated / a-C:H:W, without lubrication. Thus, the application of a-C:H:W coatings for improved failure resistance was analysed by performing tests on a coated spindle bearing operated without any lubrication. An a-C:H:W coating (formerly known as DLC, new notation according to guideline VDI 2840) was deposited to the raceways of the bearing rings. It is characterised by a carbon concentration increasing towards the surface of the coating. Thereby, a carbon transfer to the counterpart within the tribological contact pairing and, in consequence, a reduction of the friction coefficient shall be initialised. Hence, the coating of the bearing raceways is expected to reduce wear of the surfaces and avoid, or at least delay, bearing failure. Figure 10 shows the test of the a-C:H:W coated spindle bearing (type HC7014E, ceramic balls) in comparison to an uncoated bearing, both loaded with 800 N. The critical bearing conditions in the uncoated reference are provoked by minimizing the used grease quantity (1 mm, weight less than 0.001 g usually 5 g). The tests were performed with increasing speeds starting from 8,000 rpm. The coated spindle bearing shows a good operating behaviour in all speed ranges up to 29,000 rpm. The temperature of the outer ring increases to approx. 70C. In contrast, the uncoated reference bearing fails after 12 h of testing during acceleration to 20,000 rpm due to cage breakage. Most likely, insufficient lubrication of the contact pair cage outer ring causes a massive heat generation. In consequence, the cage expansion reaches critical values and the cage rotation is prevented. Thus, ball forces acting on the cage bars cause the breakdown (Figure 11 (a). These results could be confirmed in a second test series. Again, the uncoated bearing failed due to cage breakage and high temperatures. 6 0 m (a) (b) Figure 11: Cage breakage and raceway surface of an uncoated spindle bearing tested without lubrication. Subsequently, the same bearings were used for a long-term test (250 h) with constant speeds of 23,500 rpm at similar load and lubrication conditions. The coated bearing endured the whole test time. The uncoated bearing was not functional, even in case of mounting a new cage. The micrograph in Figure 11 (b) shows the surface of the uncoated raceway after 23 hours of operation. The surface is massively eroded and bearing is inoperable. The raceway of the bearing coated with the a-C:H:W (not illustrated) even after 250 hours of testing shows no excessive wear apart from single isolated holes in the coating of about 10 30 m. Its outer ring temperature did not exceed 50C during the whole test.