康熙,陳光雄,朱琪,董丙杰
摩擦磨損與潤(rùn)滑
軸盤制動(dòng)對(duì)高速列車車輪多邊形磨耗的影響
康熙1,2,陳光雄1,2,朱琪1,2,董丙杰1,2
(1.西南交通大學(xué) 機(jī)械工程學(xué)院,成都 610031;2.摩擦學(xué)研究所,成都 610031)
研究高速列車軸盤制動(dòng)引起車輪多邊形磨耗的形成機(jī)理,并提出相應(yīng)的抑制措施?;谀Σ磷约ふ駝?dòng)引起車輪多邊形磨耗的觀點(diǎn),建立高速列車拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)的有限元模型。采用復(fù)特征值法,分析制動(dòng)工況下制動(dòng)盤和制動(dòng)片摩擦激勵(lì)的振動(dòng)。根據(jù)等效阻尼比判斷摩擦自激振動(dòng)的不穩(wěn)定性,等效阻尼比越小,則不穩(wěn)定振動(dòng)發(fā)生趨勢(shì)越強(qiáng)。當(dāng)?shù)刃ё枘岜刃∮讪C0.001時(shí),不穩(wěn)定振動(dòng)的振幅會(huì)克服系統(tǒng)阻尼逐漸增大。為了考慮模型中非線性因素的影響,采用瞬時(shí)動(dòng)態(tài)仿真,獲得制動(dòng)時(shí)輪軌間的法向接觸力,通過(guò)功率譜密度分析,獲得輪軌振動(dòng)主頻。此外,分析軸盤制動(dòng)系統(tǒng)安裝位置和3種類型的制動(dòng)片對(duì)車輪多邊形磨耗的影響。軸盤制動(dòng)系統(tǒng)摩擦制動(dòng)容易激勵(lì)出637 Hz左右的不穩(wěn)定振動(dòng),由于復(fù)特征值分析與瞬時(shí)動(dòng)態(tài)分析求解方法不同,因此該不穩(wěn)定振動(dòng)頻率的計(jì)算結(jié)果存在6%左右的相對(duì)誤差。軸盤制動(dòng)系統(tǒng)的安裝位置對(duì)于不穩(wěn)定振動(dòng)的發(fā)生趨勢(shì)具有重要影響,考慮到軸盤制動(dòng)系統(tǒng)實(shí)際安裝空間,當(dāng)制動(dòng)壓力角為–10°~10°時(shí),637 Hz左右的振動(dòng)對(duì)應(yīng)的等效阻尼比隨壓力角的增大而減小。采用多個(gè)蜂窩狀制動(dòng)單元組成的制動(dòng)片,在制動(dòng)時(shí)可引起602 Hz左右的不穩(wěn)定振動(dòng)。當(dāng)制動(dòng)片表面存在復(fù)合溝槽結(jié)構(gòu)時(shí),在550~650 Hz內(nèi),沒(méi)有等效阻尼比小于–0.001的不穩(wěn)定振動(dòng)。當(dāng)高速列車運(yùn)行速度為300 km/h時(shí),軸盤摩擦制動(dòng)引起的637 Hz左右的不穩(wěn)定振動(dòng)可通過(guò)輪對(duì)傳導(dǎo)至輪軌系統(tǒng)中,引起輪軌摩擦功周期性波動(dòng),從而導(dǎo)致拖車車輪發(fā)生22~23階多邊形磨耗。在滿足制動(dòng)系統(tǒng)安裝要求的條件下,適當(dāng)增大壓力角,可減輕由軸盤制動(dòng)引起的車輪多邊形磨耗。采用多個(gè)蜂窩狀制動(dòng)單元組成的制動(dòng)片,容易導(dǎo)致拖車車輪發(fā)生20~21階多邊形磨耗。在制動(dòng)片表面添加復(fù)合溝槽結(jié)構(gòu),可抑制由軸盤制動(dòng)引起的19~23階車輪多邊形磨耗。
軸盤制動(dòng)系統(tǒng);車輪多邊形磨耗;摩擦自激振動(dòng);數(shù)值仿真;高速列車;表面溝槽
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近年來(lái),我國(guó)高速鐵路建設(shè)蓬勃發(fā)展,但與此同時(shí),多樣、復(fù)雜化的列車運(yùn)行環(huán)境導(dǎo)致高鐵列車車輪踏面損傷[1]現(xiàn)象越來(lái)越嚴(yán)重。車輪多邊形磨耗,也被稱為車輪非圓磨耗或橢圓形磨耗,指新造或鏇修后的車輪使用一段時(shí)間后,車輪與鋼軌滾動(dòng)接觸的踏面周向出現(xiàn)不均勻磨耗的現(xiàn)象。高鐵車輪多邊形磨耗引起的強(qiáng)迫振動(dòng)不僅會(huì)影響乘坐舒適性,而且當(dāng)振動(dòng)頻率接近列車零部件的固有頻率時(shí),會(huì)產(chǎn)生共振,降低車軸、軸承等零部件的使用壽命,這對(duì)于高速行駛的列車而言,具有很大的安全隱患[2]。在車輪多邊形治理方面,運(yùn)營(yíng)公司普遍采用鏇修的方法來(lái)消除車輪多邊形磨耗,但該方法不僅會(huì)縮短車輪壽命,增加運(yùn)營(yíng)成本,而且鏇修后的車輪在投入使用一段時(shí)間后,仍會(huì)出現(xiàn)多邊形磨耗現(xiàn)象。只有充分認(rèn)識(shí)車輪多邊形磨耗的形成機(jī)理及其影響因素,才能從根本上解決問(wèn)題。
關(guān)于車輪多邊形磨耗的形成機(jī)理,國(guó)內(nèi)外學(xué)者進(jìn)行了詳細(xì)的研究。由于車輪多邊形問(wèn)題涉及了車輛系統(tǒng)動(dòng)力學(xué)、輪軌接觸力學(xué)、振動(dòng)力學(xué)、材料科學(xué)、摩擦學(xué)等多學(xué)科的內(nèi)容,目前國(guó)內(nèi)外學(xué)者對(duì)于車輪多邊形磨耗的形成機(jī)理仍未達(dá)成共識(shí)[3]。Bommundt[4]基于車輪初始不圓順與車輪轉(zhuǎn)動(dòng)慣量共同作用導(dǎo)致車輪多邊形磨耗的觀點(diǎn),采用攝動(dòng)技術(shù)進(jìn)行了研究。結(jié)果表明,車輛速度越快,車輪非圓浪涌低次諧波形成越快。Meywerk[5]采用輪對(duì)-鋼軌彈性模型,對(duì)車輪廓面不圓順的發(fā)展進(jìn)行了研究,發(fā)現(xiàn)輪對(duì)左右車輪廓面的多邊形相位差越大,車輪多邊形形成越快,輪對(duì)的第一、二階彎曲振動(dòng)模態(tài)對(duì)車輪多邊形磨耗有重要影響。Meinders等[6]建立了輪對(duì)-鋼軌彈性系統(tǒng)接觸模型,使用車輪磨損反饋圈,將初始模型推廣為長(zhǎng)期磨損模型,研究了高速車輛車輪初始不圓順和輪對(duì)失衡對(duì)車輪多邊形磨耗的影響,發(fā)現(xiàn)車輪初始不平順對(duì)車輪多邊形磨耗的發(fā)展具有重要影響。初始二階和二階以上的高階車輪多邊形相對(duì)穩(wěn)定,靜態(tài)、動(dòng)態(tài)不平衡對(duì)車輪多邊形磨耗的發(fā)展影響較小。Johansson[7]對(duì)瑞典鐵路所使用的99種運(yùn)營(yíng)里程超過(guò)100 000 km的車輪進(jìn)行了車輪不圓度實(shí)測(cè),發(fā)現(xiàn)客車、貨車、通勤車、地鐵車輛均出現(xiàn)了不同程度的車輪多邊形磨耗,其中高速客車最為嚴(yán)重。在加工過(guò)程中,三角卡盤固定不當(dāng)會(huì)導(dǎo)致部分地鐵車輪形成初始多邊形。Jin等[8]通過(guò)試驗(yàn)與理論計(jì)算結(jié)合的方法,對(duì)地鐵車輛車輪多邊形磨耗的形成機(jī)理進(jìn)行了研究,發(fā)現(xiàn)72 Hz左右的輪對(duì)一階彎曲共振導(dǎo)致了地鐵車輪9階多邊形磨耗。Ma等[9]通過(guò)理論分析和動(dòng)力學(xué)仿真對(duì)地鐵車輪多邊形磨耗的產(chǎn)生機(jī)理進(jìn)行了研究,提出車輪滾動(dòng)多周的振動(dòng)導(dǎo)致地鐵車輪9階多邊形磨耗的觀點(diǎn)。Tao等[10]通過(guò)試驗(yàn)與數(shù)值模型分析對(duì)地鐵車輛車輪多邊形磨耗機(jī)理進(jìn)行了研究,研究發(fā)現(xiàn),P2共振是導(dǎo)致車輪形成5~8階多邊形的原因,通過(guò)改善制動(dòng)閘片與車輪踏面的匹配關(guān)系、降低牽引力、改善鋼軌焊接處不平順、降低扣件剛度等可抑制車輪多邊形磨耗。趙曉男等[11-12]建立了輪對(duì)-軌道系統(tǒng)有限元模型,對(duì)多邊形車輪形成機(jī)理進(jìn)行了研究,認(rèn)為當(dāng)高速線路發(fā)生制動(dòng)滑動(dòng)時(shí),輪軌間的蠕滑力趨于飽和,可能會(huì)導(dǎo)致輪軌系統(tǒng)發(fā)生摩擦自激振動(dòng),進(jìn)而引起車輪多邊形磨耗。Wu等[13]采用有限元方法研究了制動(dòng)引起的不穩(wěn)定振動(dòng)對(duì)輪軌磨耗的影響,研究表明,列車制動(dòng)可能引起車輪踏面產(chǎn)生波浪型磨耗。Ma等[14]采用車輛-軌道多體動(dòng)力學(xué),仿真研究了高速列車高階車輪多邊形磨耗的形成和發(fā)展,并分析了季節(jié)和車輪在車輛縱向分布位置的影響。研究發(fā)現(xiàn),同一轉(zhuǎn)向架輪對(duì)間鋼軌的三階彎曲模態(tài)可導(dǎo)致高階車輪多邊形磨耗的形成,相鄰轉(zhuǎn)向架輪對(duì)間鋼軌的彎曲共振會(huì)加速該磨耗的進(jìn)一步發(fā)展。此外,多雨和高溫導(dǎo)致了車輪多邊形磨耗在夏季比其他季節(jié)更嚴(yán)重。列車在包含曲線區(qū)段的高速線路上往返運(yùn)營(yíng)導(dǎo)致了第一位和第四位的車輪多邊形磨耗更為嚴(yán)重。
在先前研究中,軸盤制動(dòng)系統(tǒng)有限元模型被用于預(yù)測(cè)可導(dǎo)致制動(dòng)尖叫噪聲的摩擦自激振動(dòng)。通過(guò)對(duì)比仿真結(jié)果與現(xiàn)場(chǎng)測(cè)試[15],分析、驗(yàn)證了該模型的準(zhǔn)確性和可靠性。本文以哈爾濱-大連客運(yùn)專線上運(yùn)行的CRH3型動(dòng)車組拖車為原型,建立了直線區(qū)段高速鐵路拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)的有限元模型,對(duì)制動(dòng)工況下軸盤制動(dòng)系統(tǒng)中因摩擦激勵(lì)的振動(dòng)與車輪多邊形磨耗之間的關(guān)系進(jìn)行了研究。研究發(fā)現(xiàn),制動(dòng)盤-制動(dòng)片摩擦自激振動(dòng)可通過(guò)輪對(duì)傳遞至輪軌系統(tǒng)中,從而引起輪軌摩擦功周期性變化,導(dǎo)致車輪多邊形磨耗。
直線區(qū)段高速列車拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)的接觸模型如圖1所示。圖1中,SVL、SVR分別為輪對(duì)左右兩端的垂向懸掛力;L、R和L、R分別為左右輪軌間的法向接觸力和蠕滑力;L、R分別為左右輪軌間的接觸角;RL、RL和RV、RV分別為扣件對(duì)鋼軌橫向和垂向的支撐剛度、阻尼;F、F分別為地基對(duì)軌道板的垂向支撐剛度、阻尼,部分參數(shù)的具體數(shù)值[13]見(jiàn)表1。
圖1 拖車輪對(duì)-軌道系統(tǒng)的接觸模型
表1 拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)具體參數(shù)
Tab.1 Parameters of finite element model of a trailer wheelset-track system with an axle-mounted disc brake system
在Abaqus中建立的直線線路上高速列車拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)的有限元模型如圖2所示。該模型主要由拖車輪對(duì)、制動(dòng)盤、制動(dòng)片、閘片托、制動(dòng)支架、鋼軌和軌道板組成,網(wǎng)格單元類型為C3D8I。在制動(dòng)系統(tǒng)中,起連接作用的銷部件采用鉸鏈單元模擬。制動(dòng)盤與制動(dòng)片之間相互接觸,制動(dòng)片與閘片托之間采用綁定約束。閘片托與制動(dòng)支架前端通過(guò)鉸鏈單元連接,僅保留兩部件之間的相對(duì)轉(zhuǎn)動(dòng)自由度。制動(dòng)支架中間孔與機(jī)架之間通過(guò)鉸鏈單元連接,除轉(zhuǎn)動(dòng)以外的其他自由度均被約束,制動(dòng)力作用在支架末端的銷孔上,制動(dòng)時(shí),通過(guò)杠桿原理,將力放大后傳遞到制動(dòng)片上。車輪和鋼軌、制動(dòng)盤和制動(dòng)片之間的接觸關(guān)系均為摩擦關(guān)系??奂偷来驳闹蝿偠?、阻尼由彈簧、阻尼單元模擬,有限元模型中的材料參數(shù)[13]見(jiàn)表2。
圖2 輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)的有限元模型
表2 有限元模型的材料參數(shù)
Tab.2 Material parameters of finite element model
采用Abaqus軟件對(duì)制動(dòng)工況下高速列車拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)的穩(wěn)定性進(jìn)行復(fù)特征值分析[16-17],在摩擦耦合的作用下,系統(tǒng)的運(yùn)動(dòng)方程為[18-20]:
式中:和分別為運(yùn)動(dòng)方程式(1)的特征值和特征向量。利用QZ法對(duì)式(2)進(jìn)行求解,可得通解為:
高速列車在正常運(yùn)行時(shí),車輪與鋼軌間存在較小的滾動(dòng)摩擦力。列車制動(dòng)時(shí),一般情況下車輪與鋼軌之間不發(fā)生滑動(dòng),輪軌間仍為滾動(dòng)摩擦,而制動(dòng)片與制動(dòng)盤之間存在較大摩擦力。為了排除車輛在制動(dòng)工況下輪軌間摩擦力的影響,在有限元模型中設(shè)置拖車輪對(duì)和軸盤制動(dòng)系統(tǒng)的平移速度為300 km/h,車輪半徑為0.46 m,輪對(duì)的轉(zhuǎn)動(dòng)速度為181.159 rad/s,使輪軌間縱向蠕滑率為0。通過(guò)復(fù)特征值分析,計(jì)算了制動(dòng)時(shí)拖車輪對(duì)-軸盤制動(dòng)-軌道系統(tǒng)可能發(fā)生的摩擦自激振動(dòng)在頻域上的分布情況,如圖3所示。637 Hz左右的不穩(wěn)定振動(dòng)對(duì)應(yīng)的負(fù)等效阻尼比最小為–0.051 55,故可認(rèn)為列車在制動(dòng)時(shí)這個(gè)不穩(wěn)定振動(dòng)模態(tài)最容易被激發(fā),即制動(dòng)系統(tǒng)最容易產(chǎn)生頻率為637 Hz左右的不穩(wěn)定振動(dòng)。該不穩(wěn)定振動(dòng)模態(tài)如圖4所示,鋼軌無(wú)明顯變化,變形主要發(fā)生在輪對(duì)上,其中軸盤沿車軸方向的變形最為嚴(yán)重,說(shuō)明制動(dòng)引起的不穩(wěn)定振動(dòng)主要發(fā)生在軸盤制動(dòng)系統(tǒng)中。該振動(dòng)可通過(guò)安裝在車軸上的制動(dòng)盤傳導(dǎo)至車輪上,從而引起輪軌系統(tǒng)發(fā)生不穩(wěn)定振動(dòng),導(dǎo)致車輪多邊形磨耗。
圖3 不穩(wěn)定振動(dòng)的分布
圖4 不穩(wěn)定振動(dòng)模態(tài)
輪軌表面的損傷問(wèn)題通常被認(rèn)為是由輪軌間摩擦功的變化導(dǎo)致的,文獻(xiàn)[22]提出的磨損公式為:
式中:為單位時(shí)間內(nèi)的磨損量;為磨損常數(shù);為摩擦功;為長(zhǎng)時(shí)摩擦功,是恒定常數(shù)。根據(jù)式(5)可知,單位時(shí)間內(nèi)輪軌的磨損量主要由摩擦功決定,它的計(jì)算公式為:
式中:為輪軌間法向接觸力;Δ為輪軌間相對(duì)滑移速度;、可通過(guò)牽引-滑移率試驗(yàn)曲線獲得,均為常數(shù)。由式(6)可知,在恒定縱向蠕滑力和相對(duì)滑移速度的假設(shè)下,輪軌間法向接觸力的周期性波動(dòng)會(huì)導(dǎo)致摩擦功以相同的頻率發(fā)生周期性變化,進(jìn)而導(dǎo)致車輪踏面發(fā)生非均勻磨耗。因此,為了進(jìn)一步分析制動(dòng)時(shí)制動(dòng)片與制動(dòng)盤間的摩擦力誘發(fā)的自激振動(dòng)在輪軌系統(tǒng)中的傳導(dǎo)情況,對(duì)拖車輪對(duì)在直線線路上的軸盤制動(dòng)過(guò)程進(jìn)行瞬時(shí)動(dòng)態(tài)分析,得到0.2 s內(nèi)左右輪軌間法向接觸力的變化情況,如圖5所示。通過(guò)功率譜密度分析,得出左右輪軌間法向接觸力的振動(dòng)主頻在600 Hz左右,如圖6所示。將瞬時(shí)動(dòng)態(tài)分析獲得的輪軌振動(dòng)主頻與復(fù)特征值分析獲得的最易發(fā)生的不穩(wěn)定振動(dòng)頻率進(jìn)行對(duì)比,發(fā)現(xiàn)這兩種方法預(yù)測(cè)的結(jié)果比較接近。由于復(fù)特征值分析未考慮非線性因素,而瞬時(shí)動(dòng)態(tài)分析是通過(guò)顯式時(shí)間積分求解系統(tǒng)的動(dòng)態(tài)響應(yīng),因此這兩種方法的計(jì)算結(jié)果存在6%左右的相對(duì)誤差。
圖5 輪軌間法向接觸力
圖6 輪軌間法向接觸力PSD分析結(jié)果
一般情況下,直線區(qū)段上高速列車的運(yùn)行速度取300 km/h,拖車車輪的名義滾動(dòng)圓半徑為0.46 m,則車輪多邊形磨耗階數(shù)的計(jì)算公式為:
當(dāng)頻率為637 Hz左右時(shí),通過(guò)計(jì)算可以得到對(duì)應(yīng)的車輪多邊形磨耗階數(shù),即該不穩(wěn)定振動(dòng)頻率可能引起車輪22~23階多邊形磨耗。統(tǒng)計(jì)結(jié)果[23]顯示,當(dāng)高速列車的運(yùn)行速度為300 km/h時(shí),車輪多邊形引起的振動(dòng)的主要頻率為550~650 Hz,對(duì)應(yīng)車輪踏面上19~23階多邊形磨耗,與仿真結(jié)果中預(yù)測(cè)的不穩(wěn)定振動(dòng)頻率近乎一致。故可認(rèn)為本文的車輪多邊形磨耗預(yù)測(cè)結(jié)果與實(shí)際情況較為接近,即預(yù)測(cè)模型具有較高的可靠性。
文獻(xiàn)[24]指出,制動(dòng)壓力角對(duì)制動(dòng)系統(tǒng)的不穩(wěn)定振動(dòng)發(fā)生趨勢(shì)有重要影響。壓力角是指制動(dòng)片壓力線與水平線之間的夾角,規(guī)定順時(shí)針為正。為了研究制動(dòng)系統(tǒng)安裝位置對(duì)高速列車拖車車輪多邊形磨耗的影響,根據(jù)制動(dòng)系統(tǒng)實(shí)際安裝空間,分別取壓力角為10°、5°、–5°、–10°。不同制動(dòng)壓力角對(duì)應(yīng)系統(tǒng)的不穩(wěn)定振動(dòng)模態(tài)分布如圖7所示。從圖7中可以看出,最可能發(fā)生的不穩(wěn)定振動(dòng)頻率均在637 Hz左右,對(duì)應(yīng)的負(fù)等效阻尼比分別為–0.029 19、–0.042 89、–0.056 62、–0.062 23,即負(fù)等效阻尼比隨著壓力角的增大而減小。當(dāng)=10°時(shí),637 Hz左右的不穩(wěn)定振動(dòng)發(fā)生趨勢(shì)最弱,在–10°~10°的范圍內(nèi)。壓力角為10°時(shí),制動(dòng)引起車輪22~24階多邊形磨耗的可能性最低。
圖7 制動(dòng)系統(tǒng)安裝位置對(duì)不穩(wěn)定振動(dòng)分布的影響
為了研究不同類型的制動(dòng)片對(duì)高速列車拖車車輪多邊形磨耗的影響,分別研究了采用A、B、C 3種類型制動(dòng)片的軸盤制動(dòng)系統(tǒng)在制動(dòng)時(shí)的不穩(wěn)定振動(dòng)分布情況,制動(dòng)片和閘片托的有限元模型如圖8所示。其中,B型制動(dòng)片由多個(gè)蜂窩狀制動(dòng)單元組成,C型制動(dòng)片由A型制動(dòng)片改造(在表面添加復(fù)合溝槽結(jié)構(gòu))而成。實(shí)驗(yàn)研究[25]表明,溝槽結(jié)構(gòu)會(huì)擾亂摩擦系統(tǒng)連續(xù)的自激振動(dòng),對(duì)制動(dòng)尖叫噪聲有良好的抑制作用。
當(dāng)車輛以300 km/h的速度運(yùn)行時(shí),采用3種不同制動(dòng)片對(duì)應(yīng)的系統(tǒng)在制動(dòng)工況下的不穩(wěn)定振動(dòng)頻率分布如圖9所示。從圖9中可以看出,采用B型制動(dòng)片系統(tǒng)的不穩(wěn)定振動(dòng)頻率(602 Hz)對(duì)應(yīng)的負(fù)等效阻尼比最小,發(fā)生的可能性最高。通過(guò)公式(7)可以計(jì)算出,該不穩(wěn)定振動(dòng)頻率對(duì)應(yīng)的車輪多邊形磨耗階數(shù)為20~21階,即采用B型制動(dòng)片的軸盤制動(dòng)系統(tǒng)在制動(dòng)時(shí)可導(dǎo)致拖車車輪20~21階多邊形磨耗。采用C型制動(dòng)片的系統(tǒng)制動(dòng)時(shí),在高速列車車輪多邊形主要振動(dòng)頻率范圍(550~650 Hz)內(nèi),沒(méi)有等效阻尼比小于–0.001的不穩(wěn)定振動(dòng),即制動(dòng)系統(tǒng)采用C型制動(dòng)片可抑制由軸盤制動(dòng)引起的高速列車車輪19~23階多邊形磨耗。
圖8 3種類型制動(dòng)片和閘片托的有限元模型
圖9 不同類型的制動(dòng)片對(duì)不穩(wěn)定振動(dòng)分布的影響
1)軸盤制動(dòng)容易引起頻率為637 Hz左右的不穩(wěn)定振動(dòng)。當(dāng)高速列車運(yùn)行速度為300 km/h時(shí),可能導(dǎo)致拖車車輪發(fā)生22~23階多邊形磨耗。
2)軸盤制動(dòng)系統(tǒng)的安裝位置對(duì)高速列車拖車車輪多邊形磨耗有重要影響。在–10°~10°內(nèi),適當(dāng)增大制動(dòng)壓力角,可減緩由軸盤制動(dòng)引起的車輪多邊形磨耗的發(fā)生。
3)由多個(gè)蜂窩狀制動(dòng)單元組成的B型制動(dòng)片在制動(dòng)時(shí)容易引起602 Hz左右的不穩(wěn)定振動(dòng),導(dǎo)致拖車車輪發(fā)生20~21階多邊形磨耗。制動(dòng)系統(tǒng)采用帶有復(fù)合溝槽結(jié)構(gòu)的C型制動(dòng)片,可抑制由摩擦自激振動(dòng)引起的19~23階車輪多邊形磨耗。
本文主要采有限元仿真研究了軸盤制動(dòng)對(duì)車輪多邊形磨耗的影響。在下一階段的工作中,將采用縮尺輪軌實(shí)驗(yàn)對(duì)結(jié)論進(jìn)一步驗(yàn)證。
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Influence of Axle-Mounted Disc Brake on Polygonal Wear of High-Speed Train Wheels
1,2,1,2,1,2,1,2
(1. School of Mechanical Engineering, Southwest Jiaotong University, Chengdu 610031, China; 2. Tribology Research Institute, Chengdu 610031, China)
In order to study the formation mechanism of wheel polygonal wear on high-speed trains and propose corres-ponding inhibition measures. Based on the viewpoint that the friction-induced self-excited vibration causes wheel polygonal wear, a finite element model including high-speed train trailer wheelset, axle-mounted disc brake system and track was established. The vibrations excited by the friction between discs and pads under brake conditions were analyzed by the complex eigenvalue method. According to the effective damping ratio to judge the instability of excited vibrations, the smaller the effective damping ratio, the stronger the occurrence trend of unstable vibration. When the effective damping ratio was less than –0.001, the amplitude of unstable vibrations increased gradually overcoming the system damping. The wheel-rail normal contact forces during braking were obtained by the transient dynamic simulation, which took into account the influence of nonlinear factors. And then, the main frequencies of wheel-rail vibrations were calculated by the power spectral density (PSD) analysis of the wheel-rail normal contact forces. In addition, it was investigated that the effect of the position of the axle-mounted disc brake system and three types of brake pads on wheel polygonal wear. Friction braking of axle-mounted disc brake system was easy to excite unstable vibration about 637 Hz. Because of the difference between complex eigenvalue analysis and instantaneous dynamic analysis, there was a relative error of about 6% in the calculation results of the unstable vibration frequency. The installation position of the axle-mounted disc brake system had an important influence on the occurrence trend of unstable vibration. Considering the actual installation space of the axle-mounted disc brake system, when the brake pressure angle was in the range of –10° to 10°, the equivalent damping ratio corresponding to the vibration about 637 Hz decreased with the increase of the pressure angle. Brake pads composed of several honeycomb brake units could cause unstable vibration of about 602 Hz during braking. When there was a composite groove structure on the surface of the brake pad, there was no unstable vibration with equivalent damping ratio less than –0.001 in the frequency range of 550~650 Hz. Results show that when the high-speed train runs at 300 km/h, the unstable vibration of 637 Hz caused by axle-disc friction braking can be transmitted to the wheel-rail system through the wheelset, which causes the periodic fluctuation of wheel-rail friction work and leads to 22~23 order polygonal wear of trailer wheels. Under the condition of meeting the installation requirements of braking system, increasing the pressure angle properly can reduce the wheel polygon wear caused by axle-disc braking. Using the brake pads consisted of multiple honeycomb brake units can easily cause trailer wheels to occur 20~21 order polygonal wear, and 19~23 order wheel polygonal wear caused by axle-mounted disc brake is greatly suppressed by using the brake pads with groove-textured surface.
axle-mounted disc brake system; wheel polygonal wear; friction-induced self-excited vibration; numerical simulation; high-speed trains; surface groove
2021-03-19;
2021-08-30
KANG Xi (1996—), Male, Doctoral candidate, Research focus: frictional vibration and noise.
陳光雄(1962—),男,博士,教授,主要研究方向?yàn)槟Σ琳駝?dòng)與噪聲。
CHEN Guang-xiong(1962—), Male, Doctor, Professor, Research focus: frictional vibration and noise.
康熙, 陳光雄, 朱琪, 等.軸盤制動(dòng)對(duì)高速列車車輪多邊形磨耗的影響[J]. 表面技術(shù), 2022, 51(3): 43-50.
U271.91;TH17
A
1001-3660(2022)03-0043-08
10.16490/j.cnki.issn.1001-3660.2022.03.003
2021-03-19;
2021-08-30
國(guó)家自然科學(xué)基金(51775461)
Fund:The National Natural Science Foundation of China (51775461)
康熙(1996—),男,博士研究生,主要研究方向?yàn)槟Σ琳駝?dòng)與噪聲。
KANG Xi, CHEN Guang-xiong, ZHU Qi, et al. Influence of Axle-Mounted Disc Brake on Polygonal Wear of High-Speed Train Wheels[J]. Surface Technology, 2022, 51(3): 43-50.