高性能机床主轴的发展.pdf

Annals of the CIRP Vol 56 1 2007 395 doi 10 1016 j cirp 2007 05 092 Developments for High Performance Machine Tool Spindles C Brecher1 2 G Spachtholz1 F Paepenm ller1 1Laboratory 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 0 x106 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 varied elastically preloaded back to back arrangement HSK A 63 80 bearing diameter 70 mm max rotational speed 30 000 rpm source Weiss GmbH Figure 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 experimental investigation 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 2 out 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 0 0 5 1 1 5 2 2 5 051015202530 rotational speed 1 000 rpm 0 3 6 9 12 15 excess temperature K Hertzian pressure kN mm 1 in ring1 2 out ring4 3 in ring1 ET 4 out ring4 ET 1 2 3 4 ET axial preload 3 1 4 2 ax load 500 N Figure 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 ring s excess temperature up to 15 K see Figure 3 The internal spring preload amounts to 370 N 0 5 10 15 20 25 051015202530 rotational speed 1 000 rpm 325 345 365 385 405 425 spring preload N contact angles 1 in ring1 ET 2 in ring2 ET 2a in ring2 spring preload Ax Load 370 N 2b in ring2 ET 20 3 out ring3 ET 4 out ring4 ET 2a 3 4 22b 1 Figure 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 0 0 4 0 8 1 2 1 6 2 051015202530 rotational speed 1 000 rpm 0 3 6 9 12 15 excess temperature K Hertzian pressure kN mm 1 in ring1 2 out ring4 3 in ring1 ET ET ax load 370 N 4 out ring4 ET 5 in ring2 ET 6 in ring2 ET 20 1 2 4 5 6 3 Figure 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 drive Hydraulic unit Temperature sensor Flange with test bearing Test spindle Figure 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 40 C it3 ot3 t3 Inner outer ring excess temperature torque concept c tempering 40 C 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 Kl berSpeed 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 55 K is exceeded 0 10 20 30 40 50 60 04812162024283236 time h 0 0 2 0 4 0 6 0 8 1 1 2 friction torque Nm ax load 1000 N excess temperature rel to Tambient K 7t11t15t19t21t23t25t27t29t ot2 it1 ot1 it3 ot3 it2 t3 t1 t2 Figure 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 40 C 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