機(jī)床外文翻譯-高性能機(jī)床主軸的發(fā)展【中文3140字】【PDF+中文WORD】,中文3140字,PDF+中文WORD,機(jī)床,外文,翻譯,性能,主軸,發(fā)展,中文,3140,PDF,WORD 【中文3140字】 高性能機(jī)床主軸的發(fā)展 摘要：主軸系統(tǒng)在現(xiàn)代機(jī)床中的一個(gè)重要要求是實(shí)現(xiàn)更高的轉(zhuǎn)速?gòu)亩岣呒庸ば?。此外，要使主軸系統(tǒng)在一個(gè)給定的轉(zhuǎn)速范圍內(nèi)免受不正當(dāng)?shù)牟僮鳁l件且具有較好的穩(wěn)定性。本文提出了有助于主軸軸承在不同的領(lǐng)域的改進(jìn)研究系統(tǒng)。首先，提出了替代主軸軸承運(yùn)動(dòng)學(xué)四觸點(diǎn)的新結(jié)果。其次，對(duì)于浮動(dòng)軸承的配置進(jìn)行了新的解決方案的討論。提出一種改進(jìn)的圓柱滾子軸承，可以在更高的速度下操作。最后，討論了在改進(jìn)后的涂層軸承組件的故障安全性能下的潛能。在本文中介紹了這兩個(gè)的分析研究和實(shí)驗(yàn)測(cè)試。 關(guān)鍵詞：機(jī)械，主軸，軸承 1.介紹 現(xiàn)代機(jī)床的生產(chǎn)率主要取決于轉(zhuǎn)速限制和主軸單元的負(fù)荷能力。 一方面，現(xiàn)代切削刀具采用鋁或鎂，并配有立方氮化硼（ CBN）或多晶金剛石（PCD）的刀片，這樣的切削加工工具，使得切削速率從5000m/min高達(dá)10000m/min。銑削刀具應(yīng)用20至30毫米的直徑來(lái)實(shí)現(xiàn)非常高的切削速度，從而達(dá)到主軸速度超過(guò)100000 rpm的要求。根據(jù)滾動(dòng)軸承的技術(shù)在本領(lǐng)域的實(shí)際狀態(tài)，這種需求目前只能用平均直徑為30毫米的主軸軸承來(lái)實(shí)現(xiàn)。然而，由于這些極端工作條件下，主軸單元的所有功能部件主軸軸承、電動(dòng)機(jī)旋轉(zhuǎn)轉(zhuǎn)子會(huì)加載到其物理極限。 另一方面，主軸還被應(yīng)用于通用機(jī)床。這些特點(diǎn)是由不同的需求所決定的。例如，具有高的切割力和力矩適中的轉(zhuǎn)速的鋼的粗加工，在這種情況下，大直徑的主軸和軸承來(lái)承受這些載荷。 主軸設(shè)計(jì)方法的不同是源于需求的不同。為了滿足這些需求，速度特性系數(shù)nxdm必須增加高達(dá)3.5 ～4.0×106mm/min，來(lái)充分確保主軸本身和主軸軸承的剛度和穩(wěn)定性。 圖1 銑刀電機(jī)主軸滾動(dòng)軸承 2.多位（3P，4P）主軸軸承 2.1軸承幾何優(yōu)化的推動(dòng) 軸承被主要應(yīng)用在現(xiàn)代主軸機(jī)床中，必須履行最高要求運(yùn)行精度和剛度。在過(guò)去，為了提高軸承的性能開發(fā)了各種修改方法。其中，通過(guò)給較小的陶瓷球以及優(yōu)化的外圈使用特殊潤(rùn)滑劑。盡管如此，高速運(yùn)轉(zhuǎn)還是極大地降低樂(lè)軸承的使用壽命。在操作過(guò)程中底層的主要作用是由不同的用途來(lái)決定的。尤其是，高速運(yùn)轉(zhuǎn)的內(nèi)圈和外圈上的接觸角的依賴性偏差導(dǎo)致軸向和徑向剛度的減少。另外，在外圈上的接觸區(qū)域的離心力作用下，陶瓷球受到強(qiáng)烈的負(fù)荷。 軸承之所以具有不同的內(nèi)部幾何形狀是為了減少作用在滾珠上離心力所帶來(lái)的負(fù)面影響。此外，對(duì)軸承的穩(wěn)定性也有所提高，并對(duì)傳統(tǒng)軸承的內(nèi)圈和外圈滾道提供額外的接觸點(diǎn)。因此，滾珠的軸向和徑向移動(dòng)被阻止，恒定的接觸角和軸向位移使內(nèi)圈可以保證在很寬的速度范圍內(nèi)運(yùn)動(dòng)。上述這種思想被引入到軸承的設(shè)計(jì)概念中，如圖2所示（a）和（b）。圖2中的第三個(gè)（c）這種新概念的設(shè)計(jì)方法在文中也介紹了。 圖2 軸承的不同多點(diǎn)(3 p,4 p)選擇 2.2多點(diǎn)分析研究（4P）軸承 表一主要研究多點(diǎn)（3P）軸承的操作過(guò)程及理論和實(shí)踐調(diào)查。實(shí)驗(yàn)室測(cè)試的軸承制造機(jī)床和生產(chǎn)工程(WZL)是在傳統(tǒng)主軸軸承的基礎(chǔ)上的。也有一些通過(guò)數(shù)值計(jì)算分析特性的多點(diǎn)（4P)軸承。 隨后,關(guān)于新軸承的發(fā)展有了進(jìn)一步的結(jié)果,運(yùn)動(dòng)學(xué)和四個(gè)接觸分球正在考慮被使用。所有的計(jì)算都是在軸承型號(hào)為7014、直徑為12.7mm的陶瓷球上完成的。下面的圖中使用縮略語(yǔ)，在表1的中列出。 表1 3.、4、5使用的是縮用詞 在[1]的計(jì)算中表明，內(nèi)圈的軸向位移可以減少不到兩個(gè)微米，接觸角的變化依賴速度是可以預(yù)防的。但是，多點(diǎn)軸承(4 p)的安裝與徑向間是有間隙的。因此，他們對(duì)熱有非常敏感的影響，特別是過(guò)度的高溫使得軸承的內(nèi)圈可能會(huì)引起相互干擾。這些影響也發(fā)生在高速旋轉(zhuǎn)的圓柱滾子軸承。圖3說(shuō)明了多點(diǎn)（4P）軸承的內(nèi)圈和外圈在1赫茲和4赫茲下的接觸壓力。這些直接接觸直接傳遞到軸向荷載（見圖表）和最大程度所受壓力上。軸承的徑向間隙顯示為22微米的。為了防止內(nèi)圈彈出，應(yīng)選擇合適的內(nèi)圈并與軸之間有35微米的間隙量。因此，軸承是提前被安裝上的。曲線1和2不考慮軸承計(jì)算的熱效應(yīng)。高應(yīng)力值在內(nèi)圈滾道曲率的結(jié)果上是很廣的。赫茲壓力的增加造成的內(nèi)圈離心的擴(kuò)張以及作用在球上的離心力增大。 圖3 在多點(diǎn)（4P）軸承在赫茲壓力下速度和溫度 相反，曲線3和4顯示內(nèi)圈結(jié)合離心力超溫（線性增加）的影響。通過(guò)假設(shè)并計(jì)算一個(gè)梯度為1K/2000 rpm的曲線，其內(nèi)部應(yīng)力顯著增加，這是可以注意到的。在最大轉(zhuǎn)速30000 rpm的赫茲壓力下，內(nèi)環(huán)上升超過(guò)限制值為2000 N/mm2。除了這些理論結(jié)果，還要必須考慮到內(nèi)部應(yīng)力的熱增加和過(guò)載的耦合效應(yīng)。因此，人們可以看到一點(diǎn)對(duì)多點(diǎn)（4P）軸承的干擾在高速運(yùn)轉(zhuǎn)的情況下內(nèi)圈溫度明顯過(guò)剩。這個(gè)假設(shè)將在第2.3節(jié)中進(jìn)行研究。 圖2（c）顯示出第三個(gè)概念的新軸承幾何形狀的。它是為了防止在多點(diǎn)（4P）軸承的內(nèi)部超載而開發(fā)的。概念的特征是分割內(nèi)環(huán)，它是一半固定在主軸體面向主軸的刀具側(cè)面的環(huán)。于是，它可以承受從加工過(guò)程中產(chǎn)生的力。環(huán)下半部分可沿軸軸向移動(dòng)，并通過(guò)碟形彈簧壓向球，形成軸承的內(nèi)部預(yù)壓功能。 圖4中說(shuō)明了假設(shè)多余的溫度高達(dá)15 K線性（見圖3），多點(diǎn)（4P）軸承的內(nèi)圈在計(jì)算速度依賴性和運(yùn)動(dòng)學(xué)預(yù)裝中的發(fā)展，內(nèi)部彈簧預(yù)緊量為370N。 圖4 內(nèi)部彈簧預(yù)緊為370N的多點(diǎn)（4P）軸承 2.3多點(diǎn)（4P）軸承的試驗(yàn)及研究 該試驗(yàn)臺(tái)用于實(shí)驗(yàn)研究如圖6，其直接驅(qū)動(dòng)可實(shí)現(xiàn)高達(dá)40,000 rpm的最大轉(zhuǎn)速，額定扭矩達(dá)4.2Nm，額定轉(zhuǎn)速為23,000 rpm。測(cè)試主軸和驅(qū)動(dòng)由一個(gè)爪式離合器相連接，試驗(yàn)軸承可沿軸向由一個(gè)液壓活塞被加載，外圈的回火是通過(guò)水的循環(huán)在凸緣上實(shí)現(xiàn)的。由此，引起的附加滾動(dòng)接觸的外圈的加熱可被減小。內(nèi)軸承的溫度是由一個(gè)接近內(nèi)圈旋轉(zhuǎn)的非接觸式傳感器來(lái)測(cè)量的。 圖6 試驗(yàn)臺(tái) 圖7中顯示的是實(shí)驗(yàn)結(jié)果為對(duì)多點(diǎn)（4P）軸承的一個(gè)剛性和一個(gè)彈性點(diǎn)。起初，剛性軸承（概念(b)，圖2所示）是測(cè)試的，在測(cè)試中進(jìn)行不回火的外圈。接著，將柔性軸承（概念(c)，圖2 所示）與所述外圈的回火進(jìn)行了測(cè)試，測(cè)得的溫度顯示相關(guān)的環(huán)境溫度。在測(cè)試過(guò)程中的扭矩值來(lái)自電動(dòng)機(jī)的電流。在圖7中使用的縮寫在表2中說(shuō)明。 表2 在圖7中使用的縮寫 曲線[IT1]和[ OT1]說(shuō)明第一個(gè)試驗(yàn)軸承的內(nèi)圈和外圈的溫度，如概念(b)所示。軸向載荷達(dá)1,000 N，5克的軸承轉(zhuǎn)速高達(dá)19000轉(zhuǎn)，每2小時(shí)增加2000轉(zhuǎn)。 圖7 多點(diǎn)（4P)軸承的行為操作 3.可移動(dòng)的彈性圓柱滾子軸承 最高轉(zhuǎn)速主軸單元通常是基于角接觸球軸承的彈性裝置設(shè)計(jì)的。這種主軸由一個(gè)固定和移動(dòng)軸承單元，以補(bǔ)償熱運(yùn)動(dòng)的主軸伸長(zhǎng)率來(lái)設(shè)計(jì)的。主軸在溫度梯度的情況下，軸承套圈的外殼的軸向運(yùn)動(dòng)可減少甚至避免造成主軸故障。 在這種情況下，圓柱滾子軸承可稱為一個(gè)“理想”的移動(dòng)軸承。軸向相對(duì)的內(nèi)、外環(huán)是由一個(gè)螺旋滾筒的旋轉(zhuǎn)來(lái)運(yùn)動(dòng)的。然而，由于徑向干擾溫度及離心力作用于軸承組件而造成可達(dá)到的旋轉(zhuǎn)速度。因此，對(duì)高速圓柱滾子軸承的方法是減少基于功率損失而造成內(nèi)部產(chǎn)生的熱量和增加在線接觸赫茲壓力。 4.對(duì)軸承的故障安全特性功率的優(yōu)化 除了軸承設(shè)計(jì)的開發(fā)，更多的研究工作集中在傳統(tǒng)主軸軸承的故障安全性能的最優(yōu)化上。主軸故障往往是由潤(rùn)滑不足造成的，特別是潤(rùn)滑脂的潤(rùn)滑，潤(rùn)滑劑的供給不足，這些都可能會(huì)導(dǎo)致軸承的保持架破損或過(guò)熱。 5.總結(jié) 根據(jù)所設(shè)計(jì)的主軸角接觸球軸承以及圓柱滾子軸承的現(xiàn)有技術(shù)的狀態(tài)，被廣泛用于高速主軸的應(yīng)用程序中。然而，這兩種類型的軸承的旋轉(zhuǎn)速度是有限的，特別是通過(guò)物理作用如熱和離心載荷。在本文中，一些方法都瞄準(zhǔn)在軸承上的提高穩(wěn)定性和速度性能。 然而，由于不充分的滑動(dòng)軸承襯套，可動(dòng)軸承可能也會(huì)失敗。因此，適應(yīng)于高速運(yùn)轉(zhuǎn)的圓柱滾子軸承的開發(fā)工作尤為重要。軸承比傳統(tǒng)的圓柱滾子軸承表現(xiàn)出更高的合規(guī)性。這個(gè)屬性是通過(guò)提供狹窄水道或外圈和內(nèi)圈與地面形成凹槽來(lái)實(shí)現(xiàn)的。在實(shí)際測(cè)試中，這些軸承比傳統(tǒng)類型達(dá)到更高的轉(zhuǎn)速。最后，介紹了涂層下故障安全特性的主軸軸承潤(rùn)滑不足的情況的優(yōu)化。 6.參考文獻(xiàn) [1] Weck, M., Spachtholz, G., 2003, 3- and 4-Contact Point Spindle Bearings-a new Approach for High Speed Spindle Systems, Annals of the CIRP, 52/1: 311-316. [2] Moller, B., 2006, Hochgeschwindigkeits-Spindellager, Proceedings Gestaltung von Spindel-Lager-Systemen“, WZL Forum (Publisher), Aachen. [3] Harris, T.A., 2001, Rolling Bearing Analysis, 4th Edition, John Wiley & Sons, Inc, New York. [4] Cao, Y., Altintas, Y., 2004, A General Method for the Modeling of Spindle-Bearing Systems, Journal of Mechanical Design, Vol. 126: 1089 -1104. [5] Yangang, W. et al., 2004, FE-Analysis of a novel Roller Form – a deep End Cavity Roller for Roller Type Bearings, Journal of Materials Processing Technology 145: 233-241. Developments for High Performance Machine Tool Spindles C. Brecher1 (2), G. Spachtholz1, F. Paepenmüller1 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 – 4.0x106 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 (varied  elastically preloaded back-to-back arrangement HSK A 63/80 bearing diameter: 70 mm max. rotational speed: 30,000 rpm 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 Annals of the CIRP Vol. 56/1/2007 -395- doi:10.1016/j.cirp.2007.05.092 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. 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/mm2). ax. load 500 N ET 3 4 1 1: in. ring1 2: out. ring4 3: in. ring1, ET 4: out. ring4, ET 2 2 3 4 axial preload 1 Hertzian pressure [kN/mm2] excess temperature [K] 2,5 15 2 12 1,5 9 1 6 0,5 3 -399- 0 0 0 5 10 15 20 25 30 rotational speed [1,000 rpm] (a) Multipoint(3P)- bearing  (b) Multipoint(4P)- bearing  (c) Multipoint(4P)- bearing, internal spring preload 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 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 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/mm2. 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. 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 25 Ax. Load 370 N 3, 4 spring preload 2a 2b 1: in.ring1, ET 2: in.ring2, ET 2a: in.ring2 2b: in.ring2, ET, 20° 3: out.ring3, ET 4: out.ring4, ET contact angles [°] 20 15 10 5 0 425 spring preload [N] 1 405 385 2 365 345 325 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 0 5 10 15 20 25 30 rotational speed [1,000 rpm] Figure 4: Contact angles in a multipoint(4P)-bearing with internal spring preload of 370 N. 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/mm2. 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/mm2. 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/mm2 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. 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 Hertzian pressure [kN/mm2] 2 1,6 1,2 0,8 ET 15 ax. load 370 N 1 1: in.ring1 2: out.ring4 3: in.ring1, ET 4: out.ring4, ET 5: in.ring2, ET 6: in.ring2, ET, 20° 5 3 4 12 9 excess temperature [K] excess temperature (rel. to Tambient) [K] 2 6 6 load amounts to 1,000 N. The bearing is grease- lubricated with 5 g of KlüberSpeed 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. 0,4 0 60 3 50 0 0 5 10 15 20 25 30 40 rotational speed [1,000 rpm] 1,2 it1, ot1 ot2 t2 t3 t1 15t 19t 21t 25t 27t 29t 7t 11t ax. load: 1000 N it2 it3 ot3 23t friction torque [Nm] 1 0,8 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  30 20 10 0 0 4 8 12 16 20 24 28 32 36 time [h]  0,6 0,4 0,2 0 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 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 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 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). 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 (a) Narrow raceway (b) Ceramic rollers (c) Elastic rings 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 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 μm and are air-oil lubricated (oil viscosity 32 mm2/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 70°C. 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