車輛工程外文翻譯-測試工業(yè)制動器襯片摩擦特性【中文3700字】【PDF+中文WORD】,中文3700字,PDF+中文WORD,車輛,工程,外文,翻譯,測試,工業(yè),制動器,摩擦,特性,中文,3700,PDF,WORD 測試工業(yè)制動器襯片摩擦特性 摘要 關(guān)鍵詞：鼓式制動器，摩擦，測試，摩擦系數(shù) 在目前的一項研究了新的制動設(shè)備測試鼓制動器摩擦襯片工業(yè)制動器與滾筒直徑為30的制動帶直間的聯(lián)系。在安裝程序進行的測試，制動經(jīng)過一系列的循環(huán)中，鼓是從運行狀態(tài)降低到減慢到停止狀態(tài)。在每個周期的達到一定數(shù)量的耗散能量的過程中實現(xiàn)一個安全停止。這需要在設(shè)置中加入一個飛輪，這樣在的轉(zhuǎn)速情況下，系統(tǒng)的動能在緊急情況下所消耗的提升系統(tǒng)的能量匹配停下來。兩種不同的制動材料進行了比較，這兩種材料進行兩個系列的試驗研究在多個周期系數(shù)摩擦力的變化。據(jù)觀察，對襯片摩擦系數(shù)是依賴于鼓度。隨著鼓溫度的升高第一材料的摩擦系數(shù)降低，后者則有相反的行為。 彈簧電釋放鼓式制動器在工業(yè)環(huán)境中使用，如鋼米爾斯，控制起重機以及起重機的起重設(shè)備的運動。這種起重機通常由電動機提供動力，但盡管提升機電動機通常是為了產(chǎn)生更大的扭矩，減小輸出速度提升升降重物的一個可接受的水平，但它仍然可能是由電機升降過程中的電氣故障的情況下一個沉重的驅(qū)動對象。這種危險的情況被稱為塊下降。停止電機在塊下降，案例應(yīng)用彈簧，電釋放鼓式制動器使用。這些制動器包含重型彈簧推動制動蹄對與電機或傳動輸出軸旋轉(zhuǎn)的鼓?？s回彈簧，內(nèi)置電磁已被供電。電磁閥一般是連接在電機的電路，當電源輸給電動機，電磁閥也失去權(quán)力，允許彈簧將制動蹄對鼓，從而防止電動機轉(zhuǎn)動自如。當塊出現(xiàn)下降，鼓式制動器是封閉的，停止起升載荷下降并保持在它的高度。但在試圖解決起重機的電氣電路的故障，它是將負載安全上重要的。正常的程序是使用手動控制備份電路一會兒打開制動。防止過快的下降速度，剎車片刻后關(guān)閉再次，停止加載。這些行動是重復(fù)幾次，直到負載降低完全。在這個過程中，制動鼓材料分別考驗，因為總負荷必須放慢多次在沒有起重設(shè)備的牽引的幫助。 制動鼓的制動力不僅取決于由彈簧施加的力，而且所使用的材料在制動蹄與制動鼓之間的摩擦特性決定的。在使用過程中的摩擦材料的行為是因為缺乏可導(dǎo)致制動摩擦滑移由于沉重的負荷。然而，摩擦系數(shù)（COF）太高會使?jié)L筒軸和可引起高鼓的溫度和在滾筒可導(dǎo)致裂縫在鼓面甚至鼓斷裂高動態(tài)負載。如今，摩擦材料的使用范圍很廣，但是已知這些材料是高度依賴于它們的組合物和使用條件。通過對小樣本進行了一系列的測試，他們發(fā)現(xiàn)的摩擦性能和耐磨性的材料相同的材料在改變負載，滑動速度，和溫度。在另一篇研究表明也鼓材料C一對制動摩擦學(xué)性能的影響由于在特定的熱容量和熱導(dǎo)率的變化。因此，當新的制動材料的開發(fā)，仍有必要進行實驗測試來表征在與滾筒的材料組合的材料。除此之外，它是已知的，壓力分布是不均勻的傳播由于鼓和制動蹄和動態(tài)效果的幾何偏差在制動表面。這意味著，對摩擦材料不能用于對全制動性能做出可靠的預(yù)測，小規(guī)模的試驗結(jié)果外推。因此，在大多數(shù)情況下的全面測試，得到的制動性能準確的信息的唯一選擇。全面的測試設(shè)置鼓式制動器的設(shè)置原則，在以往的研究中，建立了量化的摩擦行為在連續(xù)制動。在這種情況下，局部摩擦強度的假想摩擦段改變制動過程。這一過程稱為熱不穩(wěn)定（TEI）原因，超過臨界速度，在摩擦諧波變化的穩(wěn)態(tài)制度。Tei可以通過有限元分析，準確的預(yù)測。然而，在的情況下，塊下降和程序安全地降低負載后，短暫的政權(quán)是感興趣的區(qū)域，因為沒有達到穩(wěn)態(tài)政權(quán)。為此，一個新的安裝程序是用來模擬一個更好的方法塊下降現(xiàn)狀。 在新安裝的制動器進行了一系列的周期中，鼓是從服務(wù)速度慢下來休息。當然有一個現(xiàn)實的情況，應(yīng)該有同等數(shù)量的能源消耗在一個周期為一個真正的安全停止。要獲得此，慣性系統(tǒng)的質(zhì)量矩是這樣一種方式，在服務(wù)速度系統(tǒng)的動能將匹配的最大的能量被消耗在緊急情況下選擇。 在下面的文章中，首先，測試設(shè)置的詳細信息一起提交獲得摩擦系數(shù)計算方法。以后的兩種不同的制動材料試驗數(shù)據(jù)將被討論。 測試設(shè)置的描述 正面設(shè)置的剖視圖示意圖顯示在圖1和2?？偟挠^點是建立在fig.3.the設(shè)置了包括應(yīng)用和電氣安全制動釋放M 30型彈簧，其鼓（1）是由一個直流復(fù)合驅(qū)動（在100千瓦5000 rpm）電機（17）。制動力由彈簧施加（4）推動制動蹄對鼓（2）。李寧不同摩擦材料（3）可以被安裝在制動蹄在剎車試驗他們的行為。制動壓力可以通過螺栓調(diào)節(jié)彈簧壓縮（5）和可變化之間的0和16.6 N / cm2.the后者對應(yīng)于最大制動力矩約10 kNm一COF之間的鼓和摩擦0.6.to打開制動電閥（6）供電牽引部分（7）的左側(cè)和壓縮彈簧。 圖1原理前視圖的鼓式制動 圖2示意剖面視圖的鼓式制動器設(shè)置 為了獲得一個系統(tǒng)，包含足夠的動能來模擬真實的塊的下降情況，驅(qū)動輪（8）是用來增加系統(tǒng)的慣性。鼓（1）和驅(qū)動輪（8）是由主軸進行（10）。驅(qū)動輪連接主軸使用兩個鎖緊組件（9）。主軸是由兩個自調(diào)心球軸承支承（11）是由一個彈性爪型聯(lián)軸器連接到直流電動機（12）。 滾筒和驅(qū)動輪具有相同的直徑30或760毫米。對不同的設(shè)置，旋轉(zhuǎn)部件在表1中給出的慣性矩。滾筒，驅(qū)動輪，與主軸貢獻最大的系統(tǒng)的慣性矩的部分。由于顎耦合，直流電動機的轉(zhuǎn)子旋轉(zhuǎn)和6公斤?M2慣性安裝其他旋轉(zhuǎn)部件必須加以考慮。這給設(shè)置一個總內(nèi)TIA 95.1公斤?平方米在422 kJ的總動能在900轉(zhuǎn)的服務(wù)速度的時刻。因為制動蹄的面積是0.28平方米，在每個制動周期的平均能量密度大約是1500 kJ / m2.in以前的研究severin5制動與25鼓散熱168 kJ在每個制動周期從900轉(zhuǎn)的服務(wù)速度開始被使用，提供約1100 kJ / m2.hence本研究建立的能量密度是可以申請一個更高的能量密度為材料，從相同的服務(wù)速度出發(fā)。 在制動周期，滾筒和驅(qū)動輪提出服務(wù)速度，而剎車是開放的。一旦達到900 rpm的速度，電機的功率開關(guān)合閘。當最后鼓來休息，制動打開再次和周期重復(fù)的。 在測試過程中，轉(zhuǎn)速的測量采用全站儀安裝在電動機和滾筒的表面溫度持續(xù)使用sp我- TEC 2005d紅外傳感器測量（見（18）圖）?？刂葡到y(tǒng)的所有信號的測量，通過計算機進行與德克薩斯儀器bnc-2110數(shù)據(jù)采集卡和LabVIEW編程。速度，表面溫度和負荷傳感器的力被記錄在五個樣本的頻率/二。 為了制動轉(zhuǎn)矩測量，制動器是安裝在兩個傾斜的表面（13）和（14），可以看出在fig.1.these兩支撐在支撐面垂直于兩個建筑線A和B的鼓在逆時針方向旋轉(zhuǎn)的方式制作，在支持反應(yīng)力（14）可以是負的。針對這種力的部分（15）存在時，其接觸面平行于接觸表面（14）。一個傳感器（16）與一個容量為20 kN安裝500毫米的滾筒旋轉(zhuǎn)的中心在制動過程中制動。將嘗試與滾筒轉(zhuǎn)動。傳感器將防止這種情況發(fā)生，將應(yīng)用一個力FL（N）。由于傳感器是剛性的，實際的旋轉(zhuǎn)是非常小的剎車在傾斜的表面的位置（13）不會發(fā)生明顯變化。因此，在支撐反作用力在連接線A和B在fig.1.this對齊方式的反應(yīng)力向量通過中心E的滾筒的旋轉(zhuǎn)和反力，不利于在力矩平衡這一點。計算摩擦系數(shù)的摩擦系數(shù)可從所施加的制動力矩MB計算，這可以從測得的傳感器FL表達在鼓的中心的力矩平衡力的計算（圖1）： MB = FL?0.500°FG?E（NM）（1） （N）的FG制動重力和E（M）的質(zhì)量中心到滾筒的旋轉(zhuǎn)中心的偏心。制動器的引力常數(shù)，因為制動器的實際轉(zhuǎn)動很小，偏心率可以也被認為是恒定的。當制動是開放的，沒有施加制動力矩，但因其制動質(zhì)量偏心，還有應(yīng)用于傳感器的力。在這種情況下（MB = 0）公式1成 在佛羅里達州是一個測量值。通過這種方式為3136 nm的FG?E值被發(fā)現(xiàn)約1噸。隨著制動的質(zhì)量，得到一個估計的偏心距0.31米。在計算產(chǎn)品的成品用。偏心率的估計值是只提到一個例子。 從制動力矩計算公式1，MB，COFμ可以在下面的部分解釋計算。如圖如圖4所示，制動壓力P（n／m2）乘以系數(shù)，在制動蹄表面綜合等于制動力矩MB： 從兩個制動鞋是現(xiàn)在式結(jié)果因子2可以簡化方程3。 因此，B制動蹄的寬度（0.300米），R制動鼓的半徑（0.380米），P平均制動壓力測試中（8.1 N /平方厘米= 8.1?104 N/m2）和α一制動蹄角的一半35°或0.611 RAD）。與上述數(shù)值方程成為一個制動循環(huán)過程在每個循環(huán)制動，滾筒和驅(qū)動輪被帶到900轉(zhuǎn)。這花了大約90秒。一旦鼓是在所要求的速度，數(shù)據(jù)采集開始2秒后制動器關(guān)閉。滾筒停兩秒鐘后，數(shù)據(jù)采集中斷和中斷后再次打開，循環(huán)重新開始。為了控制數(shù)據(jù)流和避免過量的數(shù)據(jù)記錄，數(shù)據(jù)記錄被中斷時，鼓了服務(wù)速度。均鼓溫度為摩擦襯片幾乎是一樣的。此外，它可以從圖6，COF顯示隨溫度略有增加觀察：COF開始在一個值為36的平均鼓溫度0.44°C和增加材料2觀察到的是一個價值約0.47.the相反的行為（圖7）。這里的COF下降隨著鼓溫度：在開始的COF = 0.47和平均鼓溫度27.2°C，而COF = 0.35的50次循環(huán)后。 圖3鼓式制動器設(shè)置 圖4示意圖的閘瓦壓力 圖5測量信號在一個制動循環(huán) 長期的測試系列 在長期的試驗，證實了這兩種材料的溫度依賴的動態(tài)。材料1的長系列試驗結(jié)果表明。又可以看出，COF的增加鼓溫度增加。值得注意的是，在25個周期短的中斷發(fā)生時，鼓溫度下降到約8°C. TEM - perature下降也清晰可見，在這個周期中COF路徑一滴。 材料2的一系列試驗結(jié)果表明該COF明確的減少與增加鼓溫度。即使對于李寧材料在鼓溫度和摩擦系數(shù)的最重要的變化發(fā)生在第一個30制動周期，一個小的變化出現(xiàn)在隨后的周期中，導(dǎo)致材料1輕微的COF的增加（0.49在250個周期）和2（COF材料略有減少0.31在250個周期）。 結(jié)論 創(chuàng)造工業(yè)制動器襯片真實的測試條件下，一種新的測試設(shè)置直徑尺寸制動的開發(fā)。從測量信號的制動襯片的摩擦系數(shù)可以計算。 在兩個不同的鼓式剎車片進行的試驗表明，第一材料有COF，鼓溫度升高，而第二個材料顯示了相反的行為。因為在COF的安全制動一個太大的減少會導(dǎo)致不安全的工作條件，第一材料應(yīng)安全制動應(yīng)用的首選材料。 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, Coef?cient 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 Coef?cient of friction last cycle 0.47 0.49 0.35 0.31 Figure 6 Coef?cient 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