车辆工程外文翻译-测试工业制动器衬片摩擦特性【中文3700字】【PDF+中文WORD】

车辆工程外文翻译-测试工业制动器衬片摩擦特性【中文3700字】【PDF+中文WORD】,中文3700字,PDF+中文WORD,车辆,工程,外文,翻译,测试,工业,制动器,摩擦,特性,中文,3700,PDF,WORD 测试工业制动器衬片摩擦特性 摘要 关键词：鼓式制动器，摩擦，测试，摩擦系数 在目前的一项研究了新的制动设备测试鼓制动器摩擦衬片工业制动器与滚筒直径为30的制动带直间的联系。在安装程序进行的测试，制动经过一系列的循环中，鼓是从运行状态降低到减慢到停止状态。在每个周期的达到一定数量的耗散能量的过程中实现一个安全停止。这需要在设置中加入一个飞轮，这样在的转速情况下，系统的动能在紧急情况下所消耗的提升系统的能量匹配停下来。两种不同的制动材料进行了比较，这两种材料进行两个系列的试验研究在多个周期系数摩擦力的变化。据观察，对衬片摩擦系数是依赖于鼓度。随着鼓温度的升高第一材料的摩擦系数降低，后者则有相反的行为。 弹簧电释放鼓式制动器在工业环境中使用，如钢米尔斯，控制起重机以及起重机的起重设备的运动。这种起重机通常由电动机提供动力，但尽管提升机电动机通常是为了产生更大的扭矩，减小输出速度提升升降重物的一个可接受的水平，但它仍然可能是由电机升降过程中的电气故障的情况下一个沉重的驱动对象。这种危险的情况被称为块下降。停止电机在块下降，案例应用弹簧，电释放鼓式制动器使用。这些制动器包含重型弹簧推动制动蹄对与电机或传动输出轴旋转的鼓。缩回弹簧，内置电磁已被供电。电磁阀一般是连接在电机的电路，当电源输给电动机，电磁阀也失去权力，允许弹簧将制动蹄对鼓，从而防止电动机转动自如。当块出现下降，鼓式制动器是封闭的，停止起升载荷下降并保持在它的高度。但在试图解决起重机的电气电路的故障，它是将负载安全上重要的。正常的程序是使用手动控制备份电路一会儿打开制动。防止过快的下降速度，刹车片刻后关闭再次，停止加载。这些行动是重复几次，直到负载降低完全。在这个过程中，制动鼓材料分别考验，因为总负荷必须放慢多次在没有起重设备的牵引的帮助。 制动鼓的制动力不仅取决于由弹簧施加的力，而且所使用的材料在制动蹄与制动鼓之间的摩擦特性决定的。在使用过程中的摩擦材料的行为是因为缺乏可导致制动摩擦滑移由于沉重的负荷。然而，摩擦系数（COF）太高会使滚筒轴和可引起高鼓的温度和在滚筒可导致裂缝在鼓面甚至鼓断裂高动态负载。如今，摩擦材料的使用范围很广，但是已知这些材料是高度依赖于它们的组合物和使用条件。通过对小样本进行了一系列的测试，他们发现的摩擦性能和耐磨性的材料相同的材料在改变负载，滑动速度，和温度。在另一篇研究表明也鼓材料C一对制动摩擦学性能的影响由于在特定的热容量和热导率的变化。因此，当新的制动材料的开发，仍有必要进行实验测试来表征在与滚筒的材料组合的材料。除此之外，它是已知的，压力分布是不均匀的传播由于鼓和制动蹄和动态效果的几何偏差在制动表面。这意味着，对摩擦材料不能用于对全制动性能做出可靠的预测，小规模的试验结果外推。因此，在大多数情况下的全面测试，得到的制动性能准确的信息的唯一选择。全面的测试设置鼓式制动器的设置原则，在以往的研究中，建立了量化的摩擦行为在连续制动。在这种情况下，局部摩擦强度的假想摩擦段改变制动过程。这一过程称为热不稳定（TEI）原因，超过临界速度，在摩擦谐波变化的稳态制度。Tei可以通过有限元分析，准确的预测。然而，在的情况下，块下降和程序安全地降低负载后，短暂的政权是感兴趣的区域，因为没有达到稳态政权。为此，一个新的安装程序是用来模拟一个更好的方法块下降现状。 在新安装的制动器进行了一系列的周期中，鼓是从服务速度慢下来休息。当然有一个现实的情况，应该有同等数量的能源消耗在一个周期为一个真正的安全停止。要获得此，惯性系统的质量矩是这样一种方式，在服务速度系统的动能将匹配的最大的能量被消耗在紧急情况下选择。 在下面的文章中，首先，测试设置的详细信息一起提交获得摩擦系数计算方法。以后的两种不同的制动材料试验数据将被讨论。 测试设置的描述 正面设置的剖视图示意图显示在图1和2。总的观点是建立在fig.3.the设置了包括应用和电气安全制动释放M 30型弹簧，其鼓（1）是由一个直流复合驱动（在100千瓦5000 rpm）电机（17）。制动力由弹簧施加（4）推动制动蹄对鼓（2）。李宁不同摩擦材料（3）可以被安装在制动蹄在刹车试验他们的行为。制动压力可以通过螺栓调节弹簧压缩（5）和可变化之间的0和16.6 N / cm2.the后者对应于最大制动力矩约10 kNm一COF之间的鼓和摩擦0.6.to打开制动电阀（6）供电牵引部分（7）的左侧和压缩弹簧。 图1原理前视图的鼓式制动 图2示意剖面视图的鼓式制动器设置 为了获得一个系统，包含足够的动能来模拟真实的块的下降情况，驱动轮（8）是用来增加系统的惯性。鼓（1）和驱动轮（8）是由主轴进行（10）。驱动轮连接主轴使用两个锁紧组件（9）。主轴是由两个自调心球轴承支承（11）是由一个弹性爪型联轴器连接到直流电动机（12）。 滚筒和驱动轮具有相同的直径30或760毫米。对不同的设置，旋转部件在表1中给出的惯性矩。滚筒，驱动轮，与主轴贡献最大的系统的惯性矩的部分。由于颚耦合，直流电动机的转子旋转和6公斤•M2惯性安装其他旋转部件必须加以考虑。这给设置一个总内TIA 95.1公斤•平方米在422 kJ的总动能在900转的服务速度的时刻。因为制动蹄的面积是0.28平方米，在每个制动周期的平均能量密度大约是1500 kJ / m2.in以前的研究severin5制动与25鼓散热168 kJ在每个制动周期从900转的服务速度开始被使用，提供约1100 kJ / m2.hence本研究建立的能量密度是可以申请一个更高的能量密度为材料，从相同的服务速度出发。 在制动周期，滚筒和驱动轮提出服务速度，而刹车是开放的。一旦达到900 rpm的速度，电机的功率开关合闸。当最后鼓来休息，制动打开再次和周期重复的。 在测试过程中，转速的测量采用全站仪安装在电动机和滚筒的表面温度持续使用sp我- TEC 2005d红外传感器测量（见（18）图）。控制系统的所有信号的测量，通过计算机进行与德克萨斯仪器bnc-2110数据采集卡和LabVIEW编程。速度，表面温度和负荷传感器的力被记录在五个样本的频率/二。 为了制动转矩测量，制动器是安装在两个倾斜的表面（13）和（14），可以看出在fig.1.these两支撑在支撑面垂直于两个建筑线A和B的鼓在逆时针方向旋转的方式制作，在支持反应力（14）可以是负的。针对这种力的部分（15）存在时，其接触面平行于接触表面（14）。一个传感器（16）与一个容量为20 kN安装500毫米的滚筒旋转的中心在制动过程中制动。将尝试与滚筒转动。传感器将防止这种情况发生，将应用一个力FL（N）。由于传感器是刚性的，实际的旋转是非常小的刹车在倾斜的表面的位置（13）不会发生明显变化。因此，在支撑反作用力在连接线A和B在fig.1.this对齐方式的反应力向量通过中心E的滚筒的旋转和反力，不利于在力矩平衡这一点。计算摩擦系数的摩擦系数可从所施加的制动力矩MB计算，这可以从测得的传感器FL表达在鼓的中心的力矩平衡力的计算（图1）： MB = FL•0.500°FG•E（NM）（1） （N）的FG制动重力和E（M）的质量中心到滚筒的旋转中心的偏心。制动器的引力常数，因为制动器的实际转动很小，偏心率可以也被认为是恒定的。当制动是开放的，没有施加制动力矩，但因其制动质量偏心，还有应用于传感器的力。在这种情况下（MB = 0）公式1成 在佛罗里达州是一个测量值。通过这种方式为3136 nm的FG•E值被发现约1吨。随着制动的质量，得到一个估计的偏心距0.31米。在计算产品的成品用。偏心率的估计值是只提到一个例子。 从制动力矩计算公式1，MB，COFμ可以在下面的部分解释计算。如图如图4所示，制动压力P（n／m2）乘以系数，在制动蹄表面综合等于制动力矩MB： 从两个制动鞋是现在式结果因子2可以简化方程3。 因此，B制动蹄的宽度（0.300米），R制动鼓的半径（0.380米），P平均制动压力测试中（8.1 N /平方厘米= 8.1•104 N/m2）和α一制动蹄角的一半35°或0.611 RAD）。与上述数值方程成为一个制动循环过程在每个循环制动，滚筒和驱动轮被带到900转。这花了大约90秒。一旦鼓是在所要求的速度，数据采集开始2秒后制动器关闭。滚筒停两秒钟后，数据采集中断和中断后再次打开，循环重新开始。为了控制数据流和避免过量的数据记录，数据记录被中断时，鼓了服务速度。均鼓温度为摩擦衬片几乎是一样的。此外，它可以从图6，COF显示随温度略有增加观察：COF开始在一个值为36的平均鼓温度0.44°C和增加材料2观察到的是一个价值约0.47.the相反的行为（图7）。这里的COF下降随着鼓温度：在开始的COF = 0.47和平均鼓温度27.2°C，而COF = 0.35的50次循环后。 图3鼓式制动器设置 图4示意图的闸瓦压力 图5测量信号在一个制动循环 长期的测试系列 在长期的试验，证实了这两种材料的温度依赖的动态。材料1的长系列试验结果表明。又可以看出，COF的增加鼓温度增加。值得注意的是，在25个周期短的中断发生时，鼓温度下降到约8°C. TEM - perature下降也清晰可见，在这个周期中COF路径一滴。 材料2的一系列试验结果表明该COF明确的减少与增加鼓温度。即使对于李宁材料在鼓温度和摩擦系数的最重要的变化发生在第一个30制动周期，一个小的变化出现在随后的周期中，导致材料1轻微的COF的增加（0.49在250个周期）和2（COF材料略有减少0.31在250个周期）。 结论 创造工业制动器衬片真实的测试条件下，一种新的测试设置直径尺寸制动的开发。从测量信号的制动衬片的摩擦系数可以计算。 在两个不同的鼓式刹车片进行的试验表明，第一材料有COF，鼓温度升高，而第二个材料显示了相反的行为。因为在COF的安全制动一个太大的减少会导致不安全的工作条件，第一材料应安全制动应用的首选材料。 TECHNICAL A RTICLE Testing the Friction Characteristics of Industrial Drum Brake Linings J. Van Wittenberghe, W. Ost, and P. De Baets Department of Mechanical Construction and Production at Ghent University, Ghent, Belgium 49 Experimental Techniques 36 (2012) 43– 49 © 2010, Society for Experimental Mechanics Keywords Drum Brake, Friction, Testing, Coefficient of Friction, Temperature Correspondence J. Van Wittenberghe, Department of Mechanical Construction and Production at Ghent University, Ghent, Belgium Email: Jeroen.VanWittenberghe@UGent.be Received: December 7, 2009; accepted: August 30, 2010 doi:10.1111/j.1747-1567.2010.00675.x  Abstract In the present study a new brake setup was developed to test drum brake linings on an industrial brake with drum diameter of 3011 . During the tests performed on the setup, the brake undergoes a series of cycles in which the drum is slowed down from service speed to standstill. In each cycle the same amount of energy is dissipated as during a realistic safety stop. This was obtained by adding a flywheel in the setup so that the system’s kinetic energy at service speed matches the energy of the hoisting system dissipated during an emergency stop. Two different brake lining materials were characterized. Both materials were subjected to two test series to study the changes in coefficient of friction over a number of cycles. It was observed that the coefficient of friction of both linings was dependent on the drum temperature. The coefficient of friction of the first material decreased with increasing drum temperature, while the latter had the opposite behaviour. Introduction Spring applied, electrically released drum brakes are used in industrial environments, such as steel mills, to control the movement of travelling cranes as well as the hoisting apparatus of the crane. Such cranes are typically powered by an electromotor, but although the hoist motors are normally geared to produce greater torque and reduce the output speeds to an acceptable level for lifting and lowering heavy objects, it remains nevertheless possible for the motor to be driven by a heavy object in case of an electrical failure during lifting. This dangerous situation is referred to as ‘‘block drop.’’ To stop the motor in case of block drop, spring applied, electrically released drum brakes are used. These brakes contain heavy springs which push the brake shoes against a drum that rotates with the motor or the transmission output shaft. To retract the springs, a built-in electric solenoid has to be powered. The solenoid is generally wired in the motor’s electrical circuit, so when power is lost to the motor, the solenoid also loses power allowing the springs to thrust the brake shoes against the drum and hence preventing the motor to turn freely. When block drop appears, the drum brake is closed, stopping  the lifted load to fall down and keeping it at its height. But before trying to solve the failure of the electrical power circuit of the crane, it is important to put the load safely on the ground. Normal procedure is then to use a backup circuit with manual control to open the brake for a moment. To prevent a too fast rate of descent, the brake is closed after a moment, stopping the load again. These actions are repeated several times until the load is lowered completely. During this procedure, the drum brake material is severally put to the test because total load has to be slowed down repeatedly without the help of hoisting apparatus traction. The drum brake’s braking power depends not only on the force applied by the springs, but is also determined by the frictional properties between the material used in the braking shoes and the drum of the brake. The behaviour of this friction material during its service life has to be known because a lack of friction can cause the brake to slip due to heavy loads. Nevertheless, a coefficient of friction (COF) that is too high can overload the drum axle and can cause high drum temperatures and high dynamic loads on the drum which can lead to cracks at the drum surface J. Van Wittenberghe, W. Ost, and P. De Baets Friction of Drum Brake Linings or even drum fracture. Nowadays, a wide range of friction materials is available, but as is known from Zhang and Wang1 the behaviour of those material is highly dependent on their composition and service conditions. Through a series of tests on small-scale samples, they found the friction performances and wear resistance of the same material to be changing with load, sliding speed, and temperature. In another study2 they showed that also the drum material can have an impact on the tribological behaviour of the brake because of changes in specific heat capacity and thermal conductivity. Hence when new brake materials are developed, it is still necessary to perform experimental tests to characterize the lining material in combination with the drum material. In addition to this it is known that the pressure distribution is not evenly spread across the surface of brakes due to both geometrical deviations of drum and brake shoes and dynamic effects. This means that extrapolations of results of small scale tests on friction material cannot always be used to make reliable predictions on the behaviour of the full-scale brake. Hence in most cases full-scale tests are the only option to get accurate information about the performance of the brake. Full Scale Test Setup Principles of the drum brake setup During previous studies, setups were developed mainly to quantify the frictional behaviour during continuous braking.3 In that case, the local fric- tion intensity of an imaginary friction lining segment changes during braking. This process is called ther- moelastic instability (TEI) and causes, over a critical speed, a steady-state regime with harmonic changes in friction. The TEI can be predicted accurately by finite element analyses.4 However, in the case of block drop and the procedure of safely lowering the load afterwards, the transient regime is the region of interest because the steady-state regime is not reached. For this purpose, a new setup was designed to simulate the block drop situation in a better way. In the new setup the brake undergoes a series of cycles in which the drum is slowed down from service speed to rest. Of course to have a realistic situation, there should be an equal amount of energy dissipated during one cycle as in a real safety stop. To obtain this, the system’s mass moment of inertia was chosen in such a way that the kinetic energy of the system at service speed would match the maximum energy to be dissipated during an emergency stop. In the following paragraphs, firstly, the test setup details are presented together with a calculating method to obtain the COF. Later the test data of the two different brake lining materials will be discussed. Test setup description Schematic drawings of both the frontal and the section view of the setup are shown in Figs. 1 and 2. A view of the total setup is given in Fig. 3. The setup consists of a spring applied and electrically released Igranic safety brake type M 3011 , whose drum (1) is driven by an electrical DC compound 100 kW (at 5000 rpm) motor (17). The braking force is applied by the spring (4) that pushes the brake shoes (2) against the drum. Different friction lining materials (3) can be mounted in the brake shoes to test their behaviour during brak- ing. The braking pressure can be set by adjusting the spring compression with the bolt (5) and can be varied between 0 and 16.6 N/cm2. The latter corresponds to a maximum braking torque of approximately 10 kNm for a COF between the drum and the friction lining of 0.6. To open the brake the solenoid (6) is powered pulling part (7) to the left and compressing the spring. To obtain a system that contains enough kinetic energy to simulate a realistic block drop situation, a drive wheel (8) is added to increase the inertia of the system. Drum (1) and drive wheel (8) are carried by the main axle (10). The drive wheel is connected to the main axle using two locking assemblies (9). The main axle is supported by two self-aligning ball bearings (11) and is connected to the DC motor by a flexible jaw coupling (12). Drum and drive wheel have the same diameter of 3011 or 760 mm. The moments of inertia of the differ- ent rotating parts of the setup are given in Table 1. Drum, drive wheel, and main axle are the parts that contribute the most to the moment of inertia of the system. Since a jaw coupling is used, the rotor of the DC motor rotates with the other rotating parts of the setup and its inertia of 6 kg·m2 has to be taken into account. This gives the setup a total moment of iner- tia of 95.1 kg·m2 resulting in a total kinetic energy of 422 kJ at the service speed of 900 rpm. Because the total brake shoe area is 0.28 m2, the mean energy density during each braking cycle is approximately 1500 kJ/m2. In a previous study by Severin5 a brake with a 2511 drum dissipating 168 kJ during each brak- ing cycle starting from a service speed of 900 rpm was used, giving an energy density of approximately 1100 kJ/m2. Hence the setup of this study is able to apply a much higher energy density into the material starting from the same service speed. During a braking cycle, the drum and the drive wheel are brought up to service speed, while the brake is open. Once the speed of 900 rpm is reached, 0.500m 5 4 6 7 2 3 e FG MB a b 1 15 FL 13 14 Figure 1 Schematic front view of the drum brake setup 1 8 12 11 9 to the motor 10 11 16 Figure 2 Schematic section view of the drum brake setup 17 1 18 8 13 from happening and will apply a force FL (N). Because the loadcell is rigid, the actual rotation is very small and the position of the brake on the inclined surfaces (13) will not change significantly. Hence the reaction forces in the supports stay aligned with the connection lines a and b in Fig. 1. This means the vector of the reaction forces goes through the centre of rotation of the drum and the reaction forces do not contribute to the torque equilibrium around this point. Figure 3 Drum brake setup Table 1 Properties of the rotating parts of the setup Inertia (kg  Calculating the coefficient of friction The COF can be calculated from the applied braking torque MB, which can be calculated from the force measured by the loadcell FL by expressing the torque equilibrium around the centre of the drum (Fig. 1): MB = FL·0.500 − FG·e (Nm) (1) with FG (N) the gravitational force of the brake and e Part Mass (kg) m2) Material (m) the eccentricity of the centre of mass to the centre Drum 320 28.8 Cast iron Drive wheel 700 56.7 Structural steel Main axle 60 3.6 42CrMo4 alloy steel Coupling 9 0.01 Steel + elastomer spider Two locking assemblies 5 0.02 Steel the power of the motor is switched off and the brake is closed. When finally the drum has come to rest, the brake is opened again and the cycle repeated. During the tests, the rotational speed was measured using a tachometer mounted on the motor and the surface temperature of the drum was continuously measured using an SP i-tec 2005D infrared sensor (see (18) in Fig. 3). The control of the system and measuring of all signals are carried out by a computer with a Texas Instruments BNC-2110 data acquisition card and a Labview programme. Speed, surface temperature and force in the loadcell were recorded at a frequency of five samples/second. In order to measure the brake torque, the brake is mounted on two inclined surfaces (13) and (14), as can be seen in Fig. 1. These two supports are manufactured in the way that the supporting surfaces are perpendicular to the two construction lines a and b. As the drum rotates in the counter clockwise of rotation of the drum. The gravitational force of the brake is constant and because the actual rotation of the brake is very small, the eccentricity can also be considered constant. When the brake is open, no braking moment is applied, but due to the eccentric centre of mass of the brake, there is still a force applied on the loadcell. For this case (MB = 0) Eq. 1 becomes FL·0.500 = FG·e (Nm) (2) where FL is a measured value. By this way a value for FG·e of 3136 Nm was found. With the mass of the brake of approximately 1 tonne, an estimated eccentricity of 0.31 m was obtained. In the calculations only the product FG·e is used. The estimated value of the eccentricity is only mentioned as an illustration. From the braking torque MB, calculated from Eq. 1, the COF μ can be calculated as explained in the following section. As is schematically shown in Fig. 4, the braking pressure p (N/m2) multiplied by the COF, integrated over the surface of the brake shoes equals the braking torque MB: r α direction, the reaction force on the support (14) can become negative. To counter this force the part (15) MB = 2·b· r·μp·rdθ (Nm) (3) −α is present, whose contact surface is parallel to the contact surface of (14). A loadcell (16) with a capacity of 20 kN is mounted 500 mm below the centre of rotation of the drum. During braking the brake will try to rotate with the drum. The loadcell will prevent this The factor 2 in Eq. 3 results from the two brake shoes that are present. Equation 3 can be simplified to MB = 4·μ·b·r2·p·α(Nm) (4) mp p a -a r braking Speed Torque Temperature 5000 100 4000 80 Speed [rpm] Torque [Nm] Temperature [°C] 3000 60 2000 1000 0  40 20 0 -1 0 1 2 3 4 Time [s] Figure 5 Measured signals during one braking cycle Figure 4 Schematic view of the pressure in the brake shoe Hence μ MB [—] (5) = 4·b·r2·p·α with b the width of the brake shoes (0.300 m), r the radius of the brake drum (0.380 m), p the mean braking pressure during the tests (8.1 N/cm2 = 8.1·104 N/m2) and α the half angle of one brake shoe (35◦ or 0.611 rad). With the above values Eq. 5 becomes MB(Nm) μ = 8574(Nm) [—] (6) Course of a braking cycle During each braking cycle, the drum and the drive wheel were brought up to 900 rpm. This took about 90 s. Once the drum was at the required speed, data acquisition started and 2 s later the brake was closed. Two seconds after the drum stopped, data acquisition was interrupted and the break opened again, after which the cycle restarted. In order to control the dataflow and avoid recording excess data, data logging was interrupted when the drum was brought up to service speed. In Fig. 5, the course of a braking cycle is shown. For this cycle the braking time is 2.2 s, in which the braking speed is brought from 900 rpm to rest. The course of the braking torque is somehow different from what one could expect from small- scale material tests. Common frictional behaviour of braking materials includes a difference in static and dynamic COF, from which we could expect the braking torque to have a peak when the brake is closed and remain constant until the drum is brought to a halt. In Fig. 5, however, it can be observed that the braking torque increases linearly for about 1.4 s after which the torque reaches a more or less stable value. This linear increase is caused by electromagnetic effects in the solenoid ((6) in Fig. 1) of the brake. When the current over the solenoid is removed, the force of the spring ((4) in Fig. 1) is not immediately applied on the braking shoes. Due to the solenoid’s self-induction, the original magnetic field only decreases gradually and hence, the braking torque is applied over a certain period of time instead of instantaneously. In this cycle the maximum braking torque is 4045 Nm, from which a COF of μ = 0.47 can be calculated according to Eq. 6. The drum temperature increases here from 27◦C before the braking to a maximum of 47◦C during the braking. Experimental Tests In following sections the results of the test series performed on two different composite brake linings with a different composition is presented. Both materials were subjected to two test series on the new setup. First, a short test series was conducted, where the objective was to test until the mean surface temperature of the drum saturated. The short test series was stopped after 50 cycles. Second, a long test series was conducted, consisting of 250 successive cycles to study the integrity of the lining material when subjected to a high number of braking cycles. The conducted tests are summarized in Table 2. The noted numbers for the materials and tests will be used according to this table in the rest of this article. Test series 1 and 3 are the short test series, 2 and 4 are the long series. Temperature [°C] Coefficient of Friction [-] Material 1 Material 2 100 Test Test Test Test 80 Series 1 Series 2 Series 3 Series 4 0.60 0.50 0.40 Number of cycles 50 250 50 250 60 0.30 Final speed (rpm) 900 900 900 900 Environment 22.5 20.4 21.0 20.8 40 temperature at start (◦C) 20 Drum temperature 31.2 22.8 27.2 21.1 Min. Temperature Max Temperature Coefficient of Friction Mean Temperature 0.20 0.10 Table 2 Summary of the tests short test series 120 at start (◦C) Mean drum temperature at end (◦C)  63.6 64.8 69.9 64.9 0 0 10 20 30 40 50 Number of Cycles 0.00 Coefficient of friction last cycle 0.47 0.49 0.35 0.31 Figure 6 Coefficient of friction and temperatures during test series 1 (short) on material 1 Short test series The results for the short test series of materials 1 and 2 are shown in Figs. 6 and 7. For both materials, the COF together with the minimum, maximum, and mean temperatures are plotted as a function of the cycle number. For both materials it can be seen that the mean temperature saturates at about 65◦C after approximately 30 cycles. At this point the minimum and maximum temperatures are also saturated, with a minimum drum temperature of about 50◦C for both materials. The maximum drum temperatures are different for both materials, as can be seen in 120 100 Temperature [°C] 80 60 40 20 0  Min. Temperature Max. Temperature Coefficient of Friction Mean Temperature 0 10 20 30 40 50 Number of Cycles 0.60 Coefficient of Friction [-] 0.50 0.40 0.30 0.20 0.10 0.00 Fig. 6, the maximum drum temperature with lining material 1 can reach peak values of about 118◦C, while only 104◦C for lining material 2 (Fig. 7). This difference is caused by the difference in COF between the two materials. The COF of material 1 is higher than that of material 2, which means the braking time will be shorter for material 1. Consequently, the same amount of kinetic energy has to be transferred from the drum to the friction lining in a shorter time, resulting in higher peak temperatures. However, because the actual braking time (about 2.5 s) is short in comparison to the total cycle time of about 96 s, the minimum and mean drum temperatures for both friction linings are practically the same. Additionally, it can be observed from Fig. 6 that the COF shows a slight increase with increasing temperature: the COF started at a value of 0.44 for a mean drum temperature of 36.0◦C and increased to a value of about 0.47. The opposite behaviour was observed for material 2 (Fig. 7). Here the COF decreased with increasing drum temperature: at start COF = 0.47 and the mean drum temperature was 27.2◦C, while the