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关键词：: x0031 解放中型货车后轮制动器设计

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解放牌中型货车后轮鼓式制动器设计

摘要

鼓式制动也叫块式制动，现在鼓式制动器的主流是内张式，它的制动蹄位于制动轮内侧，刹车时制动块向外张开，摩擦制动轮的内侧，达到刹车的目的。

制动系统在汽车中有着极为重要的作用，如果失效将会造成灾严重的后果。制动系统的主要部件就是制动器，在现代汽车上仍然广泛使用的是具有较高制动效能的蹄—鼓式制动器。本设计就摩擦式鼓式制动器进行了相关的设计和计算。在设计过程中，以实际产品为基础，根据我国目前进行制动器新产品开发的一般程序，并结合理论设计的要求，首先根据给定车型的整车参数和技术要求，确定制动器的结构形式、制动器主要参数及其选择，然后计算制动器的最大制动力矩、同步附着系数、制动力与制动力分配系数、制动器的结构参数与摩擦系数等，并在此基础上进行制动器主要零部件的结构设计。最后，完成装配图和零件图的绘制。

关键词：鼓式制动器；制动力；最大制动力矩；结构参数；摩擦系数

The design of jiefang medium-sized truck rear wheel drum brake

Abstract

Drum brake, also known as block-type brake, drum brakes, now within the mainstream style sheets, and its brake shoes located inside the brake wheel, brake brake blocks out when open, the inside wheel friction brake, to achieve the purpose of the brakes.

In the vehicle braking system has a very important role, failure will result in disaster if serious consequences. The main parts of the braking system is the brake, in the modern car is still widely used in high performance brake shoe - brake drum. The design of the friction drum brakes were related to the design and calculation. In the design process, based on the actual product, according to our current brake factory general new product development process, and theoretical design requirements, the first model of the vehicle according to the given parameter and the technical requirements, determine the brake structure and, brake main parameters and their choice, and then calculate the maximum braking torque of brake, the synchronous adhesion coefficient and brake force and brake force distribution coefficient, the structural parameters of the brake and friction coefficient, deformation shoe, brake effectiveness factor, braking deceleration, wear characteristics, brake temperature, etc., and in this brake on the basis of the structural design of major components. Finally, assembly drawings and parts to complete mapping.

KeyWords:drumbrake; braking force; maximum braking torque; Structure parameters; the coefficient of friction

1 绪论1

1.1汽车制动器发展的概况1

1.2研究制动器系统的意义2

1.3制动系应满足的要求2

1.4本设计要完成的内容2

2 鼓式制动器的结构形式与选择3

2.1鼓式制动器的结构形式4

2.1.1领从蹄式制动器4

2.1.2双领从蹄式制动器4

2.1.3双向双领从蹄式制动器4

2.1.4单项增力式制动器5

2.1.5双向增力式制动器5

3 制动器的主要参数及其选择6

3.1制动力与制动力分配系数6

3.2同步附着系数的计算10

3.3制动器最大制动力矩11

3.4制动器的结构参数与摩擦系数12

4 制动器的主要零件的结构计算15

4.1制动鼓15

4.2制动蹄15

4.3制动底板15

4.4支承16

4.5制动轮缸16

4.6摩擦材料16

4.7制动器间隙的调整方法及相应机构16

4.8液压驱动机构的设计与计算17

4.9制动器的校核17

5 结论19

致谢20

参考文献21

毕业设计（论文）知识产权声明22

毕业设计（论文）独创性声明23

附录124

附录225

A0-制动器总成.jpg

A1-制动鼓.jpg

A1-制动器底板.jpg

A1-制动器踏板.jpg

A2-制动蹄摩擦片总成.jpg

A2-制动主缸.jpg

A3-制动l轮缸总成.jpg

QQ截图20150604104729.png

QQ截图20150604105056.png

内容简介：: 测试工业制动器衬片摩擦特性摘要在目前的研究中一个新的制动设置了测试鼓制动器摩擦衬片工业制动器与滚筒直径为30。在安装程序进行的测试，制动经过一系列的循环中，鼓是从服务速度降低到停滞放缓。在每个周期的相同数量的能量耗散一个现实的安全停止。这是通过添加在安装飞轮使系统的动能在服务速度相匹配的吊装系统消耗紧急停止时获得的能量。两种不同的制动李宁材料进行了表征。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。关键词：鼓式制动器，摩擦，测试，摩擦系数，温度介绍简介应用弹簧，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，但是已知的从张和王这些材料的行为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李宁的材料。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。3在这种情况下，局部摩擦强度的假想摩擦李宁段改变制动过程。这一过程称为热不稳定moelastic（TEI）和原因，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。4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李宁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李宁打开制动电磁阀（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器设置图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。因为总制动蹄的面积是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 = FL0.500FGE（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FGE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E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，COF可以在下面的部分解释计算。如图如图4所示，制动压力P（nm2）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= 8.1104 N/m2）和一制动蹄角的一半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0.44C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27.2C，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8C. TEM - perature下降也清晰可见，在这个周期中COF路径一滴。材料2的长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，第一李宁材料有COF，鼓温度升高，而第二个李宁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。毕业设计(论文)外文资料翻译系别：机电信息系专业：机械设计制造及其自动化班级： B090207 姓名：王玮东学号： B09020724 外文出处： J.VanWittenberghe 附件： 1. 原文； 2. 译文 2013年3月TECHNICAL ARTICLETesting the Friction Characteristics of Industrial Drum BrakeLiningsJ. Van Wittenberghe, W. Ost, and P. De BaetsDepartment of Mechanical Construction and Production at Ghent University, Ghent, BelgiumKeywordsDrum Brake, Friction, Testing, Coefficient ofFriction, TemperatureCorrespondenceJ. Van Wittenberghe,Department of Mechanical Construction andProduction at Ghent University,Ghent, BelgiumEmail: Jeroen.VanWittenbergheUGent.beReceived: December 7, 2009; accepted:August 30, 2010doi:10.1111/j.1747-1567.2010.00675.xAbstractIn the present study a new brake setup was developed to test drum brake liningson an industrial brake with drum diameter of 30?. During the tests performedon the setup, the brake undergoes a series of cycles in which the drum isslowed down from service speed to standstill. In each cycle the same amount ofenergy is dissipated as during a realistic safety stop. This was obtained by addinga flywheel in the setup so that the systems kinetic energy at service speedmatches the energy of the hoisting system dissipated during an emergencystop. Two different brake lining materials were characterized. Both materialswere subjected to two test series to study the changes in coefficient of frictionover a number of cycles. It was observed that the coefficient of friction of bothlinings was dependent on the drum temperature. The coefficient of friction ofthe first material decreased with increasing drum temperature, while the latterhad the opposite behaviour.IntroductionSpring applied, electrically released drum brakes areused in industrial environments, such as steel mills, tocontrol the movement of travelling cranes as well asthe hoisting apparatus of the crane. Such cranes aretypically powered by an electromotor, but althoughthe hoist motors are normally geared to producegreater torque and reduce the output speeds to anacceptable level for lifting and lowering heavy objects,it remains nevertheless possible for the motor to bedriven by a heavy object in case of an electrical failureduring lifting. This dangerous situation is referred toas block drop. To stop the motor in case of blockdrop, spring applied, electrically released drum brakesare used. These brakes contain heavy springs whichpush the brake shoes against a drum that rotateswith the motor or the transmission output shaft. Toretract the springs, a built-in electric solenoid has tobe powered. The solenoid is generally wired in themotors electrical circuit, so when power is lost tothe motor, the solenoid also loses power allowing thesprings to thrust the brake shoes against the drumand hence preventing the motor to turn freely. Whenblock drop appears, the drum brake is closed, stoppingthe lifted load to fall down and keeping it at its height.But before trying to solve the failure of the electricalpower circuit of the crane, it is important to putthe load safely on the ground. Normal procedure isthen to use a backup circuit with manual control toopen the brake for a moment. To prevent a too fastrate of descent, the brake is closed after a moment,stopping the load again. These actions are repeatedseveral times until the load is lowered completely.During this procedure, the drum brake material isseverally put to the test because total load has to beslowed down repeatedly without the help of hoistingapparatus traction.The drum brakes braking power depends not onlyon the force applied by the springs, but is alsodetermined by the frictional properties between thematerial used in the braking shoes and the drumof the brake. The behaviour of this friction materialduring its service life has to be known because a lackof friction can cause the brake to slip due to heavyloads. Nevertheless, a coefficient of friction (COF) thatis too high can overload the drum axle and can causehigh drum temperatures and high dynamic loads onthe drum which can lead to cracks at the drum surfaceExperimental Techniques 36 (2012) 4349 2010, Society for Experimental Mechanics43Friction of Drum Brake LiningsJ. Van Wittenberghe, W. Ost, and P. De Baetsor even drum fracture. Nowadays, a wide range offriction materials is available, but as is known fromZhang and Wang1the behaviour of those material ishighly dependent on their composition and serviceconditions. Through a series of tests on small-scalesamples, they found the friction performances andwear resistance of the same material to be changingwith load, sliding speed, and temperature. In anotherstudy2they showed that also the drum material canhave an impact on the tribological behaviour of thebrake because of changes in specific heat capacityand thermal conductivity. Hence when new brakematerials are developed, it is still necessary to performexperimental tests to characterize the lining materialin combination with the drum material. In additionto this it is known that the pressure distribution is notevenly spread across the surface of brakes due to bothgeometrical deviations of drum and brake shoes anddynamic effects. This means that extrapolations ofresults of small scale tests on friction material cannotalways be used to make reliable predictions on thebehaviour of the full-scale brake. Hence in most casesfull-scale tests are the only option to get accurateinformation about the performance of the brake.Full Scale Test SetupPrinciples of the drum brake setupDuring previous studies, setups were developedmainly to quantify the frictional behaviour duringcontinuous braking.3In that case, the local fric-tion intensity of an imaginary friction lining segmentchanges during braking. This process is called ther-moelastic instability (TEI) and causes, over a criticalspeed, a steady-state regime with harmonic changesin friction. The TEI can be predicted accurately byfinite element analyses.4However, in the case ofblock drop and the procedure of safely lowering theload afterwards, the transient regime is the regionof interest because the steady-state regime is notreached. For this purpose, a new setup was designedto simulate the block drop situation in a better way.In the new setup the brake undergoes a series ofcycles in which the drum is slowed down from servicespeed to rest. Of course to have a realistic situation,there should be an equal amount of energy dissipatedduring one cycle as in a real safety stop. To obtainthis, the systems mass moment of inertia was chosenin such a way that the kinetic energy of the system atservice speed would match the maximum energy tobe dissipated during an emergency stop.In the following paragraphs, firstly, the test setupdetails are presented together with a calculatingmethod to obtain the COF. Later the test data of thetwo different brake lining materials will be discussed.Test setup descriptionSchematicdrawingsofboththefrontalandthesectionview of the setup are shown in Figs. 1 and 2. A viewof the total setup is given in Fig. 3. The setup consistsof a spring applied and electrically released Igranicsafety brake type M 30?, whose drum (1) is drivenby 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 mountedin the brake shoes to test their behaviour during brak-ing. The braking pressure can be set by adjusting thespringcompressionwiththebolt(5) andcanbevariedbetween 0 and 16.6 N/cm2. The latter corresponds toa maximum braking torque of approximately 10 kNmfor a COF between the drum and the friction liningof 0.6. To open the brake the solenoid (6) is poweredpulling part (7) to the left and compressing the spring.To obtain a system that contains enough kineticenergy to simulate a realistic block drop situation,a drive wheel (8) is added to increase the inertia ofthe system. Drum (1) and drive wheel (8) are carriedby the main axle (10). The drive wheel is connectedto the main axle using two locking assemblies (9).The main axle is supported by two self-aligning ballbearings (11) and is connected to the DC motor by aflexible jaw coupling (12).Drum and drive wheel have the same diameter of30?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 thatcontribute the most to the moment of inertia of thesystem. Since a jaw coupling is used, the rotor of theDC motor rotates with the other rotating parts of thesetup and its inertia of 6 kgm2has to be taken intoaccount. This gives the setup a total moment of iner-tia of 95.1 kgm2resulting in a total kinetic energyof 422 kJ at the service speed of 900 rpm. Becausethe total brake shoe area is 0.28 m2, the mean energydensity during each braking cycle is approximately1500 kJ/m2. In a previous study by Severin5a brakewith a 25?drum dissipating 168 kJ during each brak-ing cycle starting from a service speed of 900 rpmwas used, giving an energy density of approximately1100 kJ/m2. Hence the setup of this study is able toapply a much higher energy density into the materialstarting from the same service speed.During a braking cycle, the drum and the drivewheel are brought up to service speed, while thebrake is open. Once the speed of 900 rpm is reached,44Experimental Techniques 36 (2012) 4349 2010, Society for Experimental MechanicsJ. Van Wittenberghe, W. Ost, and P. De BaetsFriction of Drum Brake Linings1FGe324567131415abFL0.500mMBFigure 1 Schematic front view of the drum brake setupto the motor1112111810916Figure 2 Schematic section view of the drum brake setupExperimental Techniques 36 (2012) 4349 2010, Society for Experimental Mechanics45Friction of Drum Brake LiningsJ. Van Wittenberghe, W. Ost, and P. De Baets18117813Figure 3 Drum brake setupTable 1 Properties of the rotating parts of the setupPartMass (kg)Inertia (kgm2)MaterialDrum32028.8Cast ironDrive wheel70056.7Structural steelMain axle603.642CrMo4 alloy steelCoupling90.01Steel + elastomer spiderTwo locking assemblies50.02Steelthe power of the motor is switched off and the brakeis closed. When finally the drum has come to rest, thebrake is opened again and the cycle repeated.During the tests, the rotational speed was measuredusing a tachometer mounted on the motor and thesurface temperature of the drum was continuouslymeasured using an SP i-tec 2005D infrared sensor(see (18) in Fig. 3). The control of the systemand measuring of all signals are carried out by acomputer with a Texas Instruments BNC-2110 dataacquisition card and a Labview programme. Speed,surface temperature and force in the loadcell wererecorded at a frequency of five samples/second.In order to measure the brake torque, the brakeis mounted on two inclined surfaces (13) and (14),as can be seen in Fig. 1. These two supports aremanufactured in the way that the supporting surfacesare perpendicular to the two construction lines aand b. As the drum rotates in the counter clockwisedirection, the reaction force on the support (14) canbecome negative. To counter this force the part (15)is present, whose contact surface is parallel to thecontact surface of (14). A loadcell (16) with a capacityof 20 kN is mounted 500 mm below the centre ofrotationofthedrum.Duringbrakingthebrakewilltryto rotate with the drum. The loadcell will prevent thisfrom happening and will apply a force FL(N). Becausethe loadcell is rigid, the actual rotation is very smalland the position of the brake on the inclined surfaces(13) will not change significantly. Hence the reactionforcesinthesupportsstayalignedwiththeconnectionlines a and b in Fig. 1. This means the vector of thereaction forces goes through the centre of rotation ofthe drum and the reaction forces do not contribute tothe torque equilibrium around this point.Calculating the coefficient of frictionThe COF can be calculated from the applied brakingtorque MB, which can be calculated from the forcemeasured by the loadcell FLby expressing the torqueequilibrium around the centre of the drum (Fig. 1):MB= FL0.500 FGe (Nm)(1)with FG(N) the gravitational force of the brake and e(m) the eccentricity of the centre of mass to the centreof rotation of the drum. The gravitational force of thebrake is constant and because the actual rotation ofthe brake is very small, the eccentricity can also beconsidered constant. When the brake is open, nobraking moment is applied, but due to the eccentriccentre of mass of the brake, there is still a force appliedon the loadcell. For this case (MB= 0) Eq. 1 becomesFL0.500 = FGe (Nm)(2)where FLis a measured value. By this way avalue for FGe of 3136 Nm was found. With themass of the brake of approximately 1 tonne, anestimated eccentricity of 0.31 m was obtained. Inthe calculations only the product FGe is used. Theestimated value of the eccentricity is only mentionedas an illustration.From the braking torque MB, calculated from Eq. 1,the COF can be calculated as explained in thefollowing section.As is schematically shown in Fig. 4, the brakingpressure p (N/m2) multiplied by the COF, integratedover the surface of the brake shoes equals the brakingtorque MB:MB= 2b?rprd (Nm)(3)The factor 2 in Eq. 3 results from the two brakeshoes that are present. Equation 3 can be simplifiedtoMB= 4br2p(Nm)(4)46Experimental Techniques 36 (2012) 4349 2010, Society for Experimental MechanicsJ. Van Wittenberghe, W. Ost, and P. De BaetsFriction of Drum Brake Liningsa-arpmpFigure 4 Schematic view of the pressure in the brake shoeHence =MB4br2p(5)with b the width of the brake shoes (0.300 m),r the radius of the brake drum (0.380 m), p themean braking pressure during the tests (8.1 N/cm2=8.1104N/m2) and the half angle of one brake shoe(35or 0.611 rad).With the above values Eq. 5 becomes =MB(Nm)8574(Nm)(6)Course of a braking cycleDuring each braking cycle, the drum and the drivewheel were brought up to 900 rpm. This took about90 s. Once the drum was at the required speed, dataacquisition started and 2 s later the brake was closed.Two seconds after the drum stopped, data acquisitionwas interrupted and the break opened again, afterwhich the cycle restarted. In order to control thedataflowandavoidrecordingexcessdata,dataloggingwas interrupted when the drum was brought up toservice speed.In Fig. 5, the course of a braking cycle is shown.For this cycle the braking time is 2.2 s, in whichthe braking speed is brought from 900 rpm torest. The course of the braking torque is somehowdifferent from what one could expect from small-scale material tests. Common frictional behaviourof braking materials includes a difference in staticand dynamic COF, from which we could expect thebraking torque to have a peak when the brake is010002000300040005000-101234Time sSpeed rpmTorque Nm020406080100Temperature CSpeedTorqueTemperaturebrakingFigure 5 Measured signals during one braking cycleclosed and remain constant until the drum is broughtto a halt. In Fig. 5, however, it can be observedthat the braking torque increases linearly for about1.4 s after which the torque reaches a more orless stable value. This linear increase is caused byelectromagnetic effects in the solenoid (6) in Fig. 1)of the brake. When the current over the solenoidis removed, the force of the spring (4) in Fig. 1) isnot immediately applied on the braking shoes. Dueto the solenoids self-induction, the original magneticfield only decreases gradually and hence, the brakingtorque is applied over a certain period of timeinstead of instantaneously. In this cycle the maximumbraking torque is 4045 Nm, from which a COF of = 0.47 can be calculated according to Eq. 6. Thedrum temperature increases here from 27C beforethebrakingtoamaximumof47Cduringthebraking.Experimental TestsIn following sections the results of the test seriesperformed on two different composite brake liningswith a different composition is presented. Bothmaterials were subjected to two test series on thenew setup. First, a short test series was conducted,where the objective was to test until the mean surfacetemperature of the drum saturated. The short testseries was stopped after 50 cycles. Second, a long testseries was conducted, consisting of 250 successivecycles to study the integrity of the lining materialwhen subjected to a high number of braking cycles.The conducted tests are summarized in Table 2. Thenoted numbers for the materials and tests will be usedaccording to this table in the rest of this article. Testseries 1 and 3 are the short test series, 2 and 4 are thelong series.Experimental Techniques 36 (2012) 4349 2010, Society for Experimental Mechanics47Friction of Drum Brake LiningsJ. Van Wittenberghe, W. Ost, and P. De BaetsTable 2 Summary of the tests short test seriesMaterial 1Material 2TestSeries 1TestSeries 2TestSeries 3TestSeries 4Number of cycles5025050250Final speed (rpm)900900900900Environmenttemperature atstart (C)22.520.421.020.8Drum temperatureat start (C)31.222.827.221.1Mean drumtemperature atend (C)63.664.869.964.9Coefficient of frictionlast cycle0.470.490.350.31Short test seriesThe results for the short test series of materials 1and 2 are shown in Figs. 6 and 7. For both materials,the COF together with the minimum, maximum, andmean temperatures are plotted as a function of thecycle number. For both materials it can be seen thatthe mean temperature saturates at about 65C afterapproximately 30 cycles. At this point the minimumand maximum temperatures are also saturated, witha minimum drum temperature of about 50C forboth materials. The maximum drum temperaturesare different for both materials, as can be seen inFig. 6, the maximum drum temperature with liningmaterial 1 can reach peak values of about 118C,while only 104C for lining material 2 (Fig. 7). Thisdifference is caused by the difference in COF betweenthe two materials. The COF of material 1 is higherthan that of material 2, which means the brakingtime will be shorter for material 1. Consequently, thesame amount of kinetic energy has to be transferredfrom the drum to the friction lining in a shortertime,resultinginhigherpeaktemperatures.However,because the actual braking time (about 2.5 s) is shortin comparison to the total cycle time of about 96 s,the minimum and mean drum temperatures for bothfriction linings are practically the same.Additionally, it can be observed from Fig. 6 thatthe COF shows a slight increase with increasingtemperature: the COF started at a value of 0.44 fora mean drum temperature of 36.0C and increasedto a value of about 0.47. The opposite behaviourwas observed for material 2 (Fig. 7). Here the COFdecreased with increasing drum temperature: at startCOF = 0.47 and the mean drum temperature was27.2C, while the COF = 0.35 after 50 cycles.02040608010012001020304050Number of CyclesTemperature C0.0000.400.500.60Coefficient of Friction -Min. TemperatureMax TemperatureCoefficient of FrictionMean TemperatureFigure 6 Coefficient of friction and temperatures during test series 1(short) on material 102040608010012001020304050Number of CyclesTemperature C0.0000.400.500.60Coefficient of Friction -Min. TemperatureMax. TemperatureCoefficient of FrictionMean TemperatureFigure 7 Coefficient of friction and temperatures during test series 3(short) on material 2Long test seriesDuring the long-term tests, the temperature depen-dency of both materials was confirmed. In Fig. 8, theresults of the long test series on material 1 are shown.Again it can be seen that the COF increases withincreasing drum temperature. It is noted that at cycle25 a short interruption took place during which thedrum temperature dropped to about 8C. This tem-perature drop is also clearly visible as a drop in thepath of th

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