李佳琪,倪計(jì)民,石秀勇,徐曉川,劉 越,李冬冬,陳振斌
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結(jié)構(gòu)參數(shù)對(duì)增壓器浮環(huán)軸承潤(rùn)滑特性和環(huán)速比的影響
李佳琪1,倪計(jì)民1※,石秀勇1,徐曉川1,劉 越1,李冬冬1,陳振斌2
(1. 同濟(jì)大學(xué)汽車學(xué)院,上海 201804; 2. 海南大學(xué)機(jī)電工程學(xué)院,儋州 571737)
基于以往對(duì)增壓器的浮環(huán)軸承潤(rùn)滑分析中大都忽略浮環(huán)的環(huán)速比影響,或?qū)?rùn)滑性能和環(huán)速比獨(dú)立分析。該文采用數(shù)值分析方法研究了增壓器浮環(huán)軸承的潤(rùn)滑特性和環(huán)速比,分析中考慮了轉(zhuǎn)軸、浮環(huán)、軸承座之間的傳熱因素,基于Reynolds方程和浮環(huán)平衡方程,建立了浮環(huán)軸承潤(rùn)滑模型,對(duì)比分析了浮環(huán)內(nèi)、外層間隙,內(nèi)、外圓半徑4個(gè)結(jié)構(gòu)參數(shù)對(duì)浮環(huán)軸承潤(rùn)滑特性和環(huán)速比的影響。結(jié)果表明,實(shí)際設(shè)計(jì)浮環(huán)時(shí),需綜合考慮結(jié)構(gòu)參數(shù)對(duì)浮環(huán)潤(rùn)滑特性和環(huán)速比的影響及影響程度;浮環(huán)內(nèi)層間隙增加,環(huán)速比降低,與內(nèi)層間隙0.02 mm時(shí)相比,轉(zhuǎn)速60 000 r/min時(shí),內(nèi)層間隙0.04 mm時(shí)的環(huán)速比減幅達(dá)23%,內(nèi)層間隙增加,內(nèi)、外膜溫度減小,摩擦功耗略有增加,內(nèi)層間隙0.03 mm時(shí),浮環(huán)具有較理想的潤(rùn)滑性能和環(huán)速比;外層間隙0.06 mm的環(huán)速比均比外層間隙0.04 mm的環(huán)速比增加30%以上,外層間隙增加,外膜溫度減小,且轉(zhuǎn)速越高,外膜溫度減幅越大;浮環(huán)內(nèi)圓半徑越小,環(huán)速比越小,內(nèi)、外膜溫度和摩擦功耗越小,浮環(huán)潤(rùn)滑性能越好;浮環(huán)外圓半徑增加,環(huán)速比降低,但內(nèi)膜溫度、外膜溫度、總摩擦功耗和總端泄流量變化幅度均在5%以內(nèi),外圓半徑對(duì)浮環(huán)潤(rùn)滑性能影響不顯著;浮環(huán)實(shí)際設(shè)計(jì)時(shí),調(diào)整內(nèi)圓半徑比調(diào)整外圓半徑對(duì)改善浮環(huán)潤(rùn)滑性能更有效。
軸承;模型;溫度;浮環(huán)軸承;潤(rùn)滑;結(jié)構(gòu)參數(shù);環(huán)速比;傳熱
增壓器浮環(huán)軸承由兩層油膜組成,中間由浮環(huán)隔開,兩層油膜分別為轉(zhuǎn)軸與浮環(huán)之間的內(nèi)層油膜(內(nèi)膜)和浮環(huán)與軸承座之間的外層油膜(外膜),浮環(huán)軸承具有低功耗、潤(rùn)滑性能好等特點(diǎn)而被廣泛應(yīng)用。
Clarke等[1]對(duì)重載荷情況下的浮環(huán)軸承進(jìn)行潤(rùn)滑分析;Koeneke等[2]研究了高速軸承慣性力對(duì)油膜破裂的影響;Mokhtar等[3]研究了內(nèi)、外層外徑間隙對(duì)最小油膜厚度的影響;Andres等[4-7]在某一給定載荷條件下,考慮了內(nèi)膜、浮環(huán)、外膜之間的傳熱,對(duì)浮環(huán)功耗、溫度和油膜厚度進(jìn)行了預(yù)測(cè);Li等[8-10]對(duì)轉(zhuǎn)子和浮環(huán)軸承系統(tǒng)動(dòng)力學(xué)進(jìn)行了研究;Guo等[11-14]研究了徑推聯(lián)合動(dòng)靜壓的靜態(tài)特性和動(dòng)態(tài)特性;Zhang等[15]研究了氣體軸承的穩(wěn)定性;Dong等[16]研究了浮環(huán)軸承在發(fā)動(dòng)機(jī)中的應(yīng)用;Deligant等[17]采用CFD軟件對(duì)增壓器軸承性能進(jìn)行了預(yù)測(cè);Liang 等[18-20]對(duì)半浮動(dòng)軸承、浮環(huán)表面織構(gòu)和浮環(huán)失圓等影響因素進(jìn)行了分析;康召輝等[21-22]考慮了浮環(huán)渦動(dòng)對(duì)潤(rùn)滑性能的研究;師占群等[23-25]考慮了貧油條件對(duì)浮環(huán)軸承潤(rùn)滑性能的影響;寧峰平等[26]研究了過(guò)盈量對(duì)軸承預(yù)緊力的影響;同時(shí),已有一些學(xué)者也開展了浮環(huán)軸承結(jié)構(gòu)參數(shù)、浮環(huán)傳熱的影響研究[27-29]。
上述研究者往往將浮環(huán)的環(huán)速比脫離浮環(huán)軸承潤(rùn)滑性能分析之外,分別對(duì)環(huán)速比和潤(rùn)滑特性單獨(dú)分析或不考慮環(huán)速比的影響,這顯然與浮環(huán)軸承的實(shí)際工作狀況有所差異。浮環(huán)實(shí)際工作中,轉(zhuǎn)軸向浮環(huán)內(nèi)膜傳熱,浮環(huán)內(nèi)膜通過(guò)端泄帶走部分轉(zhuǎn)軸傳導(dǎo)內(nèi)膜和內(nèi)膜自身摩擦產(chǎn)生的熱量,另一部分熱量傳遞到浮環(huán);浮環(huán)將內(nèi)膜吸收的熱量傳遞到外膜,外膜通過(guò)端泄帶走部分從浮環(huán)吸收的熱量和外膜自身摩擦產(chǎn)生的熱量,其余熱量傳遞到軸承座。因此,轉(zhuǎn)軸、浮環(huán)、軸承座之間的傳熱會(huì)影響內(nèi)、外膜的潤(rùn)滑性能以及作用于浮環(huán)上的摩擦力,繼而引起環(huán)速比的變化。此外,過(guò)小的環(huán)速比不利于浮環(huán)軸承的穩(wěn)定運(yùn)轉(zhuǎn),且環(huán)速比的變化也會(huì)對(duì)潤(rùn)滑性能產(chǎn)生影響。因此,浮環(huán)的環(huán)速比和潤(rùn)滑性能互相作用、互相影響。在研究結(jié)構(gòu)參數(shù)對(duì)浮環(huán)軸承潤(rùn)滑性能影響時(shí),還應(yīng)同時(shí)考慮其對(duì)環(huán)速比的影響。
本文以一農(nóng)用柴油機(jī)增壓器浮環(huán)軸承為研究對(duì)象,以浮環(huán)軸承潤(rùn)滑模型、轉(zhuǎn)軸、浮環(huán)、軸承座熱量傳遞模型和浮環(huán)平衡模型為基礎(chǔ),采用有限差分法求解雷諾方程,采用熱變形方程求解轉(zhuǎn)軸、浮環(huán)和軸承座熱變形量,分析探究結(jié)構(gòu)參數(shù)對(duì)浮環(huán)潤(rùn)滑特性和環(huán)速比的影響。
1.1 浮環(huán)軸承動(dòng)壓潤(rùn)滑模型
圖1所示為浮環(huán)軸承結(jié)構(gòu)示意圖。
a. 徑向截面
a. Radial cross section
b. 軸向截面
b. Axial cross section
注:J為轉(zhuǎn)軸角速度,rad·s-1;R為浮環(huán)角速度,rad·s-1;i、o為內(nèi)、外膜油膜厚度,m;i、o為浮環(huán)內(nèi)、外圓寬度,mm;o為軸承座坐標(biāo)系,'''為浮環(huán)軸承坐標(biāo)系。
Note:Jis journal angular velocity, rad·s-1;Ris ring angular velocity, rad·s-1;i,oare the inner and outer film thicknesses,m;i,oare the inner and outer circle widths;ois the coordinate system of bearing block,''' is the coordinate system of floating ring.
圖1 浮環(huán)軸承結(jié)構(gòu)示意圖
Fig.1 Schematic view of structure of floating ring bearing
1.1.1 Reynolds 方程
采用Reynolds方程表征內(nèi)膜、浮環(huán)、外膜系統(tǒng)中內(nèi)層油膜和外層油膜的壓力分布,忽略體積力及慣性力的影響,由式(1)、式(2)表示。
式中i、o分別為浮環(huán)內(nèi)圈半徑和外圈半徑,mm;J為轉(zhuǎn)軸半徑,mm;i、o分別為內(nèi)膜油膜厚度和外膜油膜厚度,mm;i、o分別為內(nèi)膜油膜壓力和外膜油膜壓力,Pa;J、R分別為轉(zhuǎn)軸和浮環(huán)角速度,rad/s;i、o分別為內(nèi)膜和外膜油膜黏度,Pa·s;為油膜角坐標(biāo),rad。
1.1.2 浮環(huán)油膜厚度方程
浮環(huán)油膜厚度方程如式(3)和式(4)所示[4]。
i=i0+iTr?Tj(3)
o=o0?oTr+Tc(4)
不計(jì)軸承表面變形的油膜厚度為
浮環(huán)受熱變形引起的膜厚變化量為
轉(zhuǎn)軸受熱變形引起的膜厚變化量為
軸承座受熱變形引起的膜厚變化量為
式中i0、o0為不計(jì)浮環(huán)表面變形的內(nèi)外油膜厚度,mm;iTr、oTr分別為浮環(huán)表面熱變形引起的內(nèi)外油膜厚度變化量,mm,其中浮環(huán)熱變形使內(nèi)膜間隙變大,使外膜間隙變??;Tj為轉(zhuǎn)軸熱變形引起的內(nèi)膜油膜厚度變化量,mm,轉(zhuǎn)軸熱變形使內(nèi)膜間隙變?。籘c為軸承座熱變形引起的外膜油膜厚度變化量,mm;i0,o0分別為內(nèi)、外層油膜半徑間隙,mm;i、o分別為內(nèi)外油膜偏心率;i、o分別為內(nèi)外油膜偏位角,rad;R為浮環(huán)熱膨脹系數(shù);R為浮環(huán)溫度,K;ref為參考溫度,K;J為轉(zhuǎn)軸熱膨脹系數(shù);J為轉(zhuǎn)軸表面溫度,K;C為軸承座熱膨脹系數(shù);C為軸承座溫度,K;C為軸承座半徑,mm。
1.1.3 潤(rùn)滑油的黏溫關(guān)系
黏溫關(guān)系采用Vogel模型,這里使用CD30級(jí)潤(rùn)滑油,其黏溫關(guān)系表達(dá)式為[30]
式中為潤(rùn)滑油黏度,Pa·s;為溫度,K。
1.1.4 浮環(huán)摩擦功耗
浮環(huán)軸承摩擦功耗如式(12)-式(14)所示[4]。
式中i、o為內(nèi)外油膜摩擦功耗,W;total為浮環(huán)摩擦功耗,W;為浮環(huán)軸承內(nèi)圈寬度、外圈寬度,mm。
1.1.5 潤(rùn)滑油端泄流量
浮環(huán)軸承潤(rùn)滑油端泄流量如式(15)-式(20)所示[30]。
式中i1、i2分別為內(nèi)層潤(rùn)滑油前端、后端流量,L/s;o1、o2分別為外層潤(rùn)滑油前端、后端流量,L/s;i、o分別為內(nèi)、外層潤(rùn)滑油流量,L/s。
1.2 轉(zhuǎn)軸-浮環(huán)-軸承座傳熱模型
增壓器轉(zhuǎn)軸-浮環(huán)系統(tǒng)運(yùn)轉(zhuǎn)中,熱流由轉(zhuǎn)軸流向軸承座。其中,內(nèi)膜、浮環(huán)和外膜均保持熱量平衡。
1.2.1 浮環(huán)內(nèi)膜熱平衡方程
浮環(huán)內(nèi)膜熱量由轉(zhuǎn)軸向內(nèi)膜傳熱和內(nèi)膜摩擦兩部分組成。內(nèi)膜通過(guò)端泄帶走部分熱量,另一部分熱量則傳遞到浮環(huán),其熱平衡方程為
式中Journal-i為轉(zhuǎn)軸對(duì)浮環(huán)內(nèi)膜傳熱量,J;Journal-i=JJi(J?i),其中J為轉(zhuǎn)軸與內(nèi)膜傳熱區(qū)域面積,m2,Ji為內(nèi)膜與轉(zhuǎn)軸的對(duì)流換熱系數(shù);i為內(nèi)膜溫度,K;i=in+?i,in為進(jìn)油溫度,K;?i為內(nèi)膜溫升,K;i為內(nèi)膜摩擦產(chǎn)生熱量,J;i=i,i-out為內(nèi)膜端泄熱量散失,J,i-out=cρQi?i,c為潤(rùn)滑油比熱容,J/(kg·K);為潤(rùn)滑油密度,kg/m3;i-ring為內(nèi)膜對(duì)浮環(huán)傳熱量,J,i-ring=RiR(i?R),其中R為油膜與浮環(huán)傳熱區(qū)域面積,m2,iR為內(nèi)膜與浮環(huán)的對(duì)流換熱系數(shù)。
1.2.2 浮環(huán)熱平衡方程
浮環(huán)將從內(nèi)膜吸收的熱量傳遞到外膜,并且浮環(huán)從內(nèi)膜吸收的熱量等于浮環(huán)傳遞到外膜的熱量,其方程為
式中ring-o為浮環(huán)向外膜傳遞的熱量,J;ring-o=RRo(R?o),J,Ro為浮環(huán)與外膜的對(duì)流換熱系數(shù),o為外膜溫度,K,o=in+?o,其中?o為外膜溫升,K。
1.2.3 浮環(huán)外膜熱平衡方程
浮環(huán)外膜熱量由浮環(huán)向外膜傳熱和外膜自身摩擦做功兩部分組成。同樣,外膜通過(guò)端泄帶走部分熱量,另一部分熱量則傳遞到軸承座散失,其熱平衡方程為
式中o為外膜摩擦生熱量,J,o=o;o-out為外膜端泄熱量散失,J,o-out=co?o;o-casing為外膜傳遞到軸承座熱量散失,J,o-casing=CoC(o?C),C為外膜與軸承座傳熱區(qū)域面積,m2;oC為外膜與軸承座對(duì)流換熱系數(shù)。
1.3 浮環(huán)平衡模型
浮環(huán)軸承穩(wěn)定工作時(shí),浮環(huán)受到內(nèi)膜承載力與外膜承載力,且內(nèi)膜承載力與外膜承載力大小相等,方向相反。內(nèi)膜作用在浮環(huán)上的方向?yàn)榇怪毕蛳拢饽ぷ饔迷诟…h(huán)的方向?yàn)榇怪毕蛏?。力平衡方程為[4]
式中i、o分別為內(nèi)、外膜承載力,N;ix、ox分別為內(nèi)、外膜承載力在′,坐標(biāo)軸方向上的分量,N;iy、oy分別為內(nèi)、外膜承載力在′,坐標(biāo)軸方向上的分量,N。
1.3.2 浮環(huán)力矩平衡方程
浮環(huán)高速運(yùn)轉(zhuǎn)時(shí),內(nèi)、外油膜在浮環(huán)上產(chǎn)生的摩擦力力矩相等。力矩平衡方程為
1.4 數(shù)值分析方法
數(shù)值分析流程圖如圖2所示。
5.從作文內(nèi)容來(lái)看,31%的同學(xué)能做到中心明確,內(nèi)容具體,其余同學(xué)有不同層次問(wèn)題;從結(jié)構(gòu)來(lái)看,35%的同學(xué)能做到條理清楚,結(jié)構(gòu)完整,其余同學(xué)有不同層次問(wèn)題;從語(yǔ)言來(lái)看,31%的同學(xué)語(yǔ)言通順,用語(yǔ)準(zhǔn)確,其余同學(xué)有不同層次問(wèn)題。
為驗(yàn)證論文模型正確性,與文獻(xiàn)[6]試驗(yàn)結(jié)果進(jìn)行比較。首先,本文模型與文獻(xiàn)[4-6]中異同處在于:文獻(xiàn)[4-6]中的計(jì)算模型沒(méi)有考慮內(nèi)膜與轉(zhuǎn)軸之間的傳熱和外膜與軸承座之間的傳熱,以轉(zhuǎn)軸溫度近似等于內(nèi)膜溫度,軸承座溫度近似等于外膜溫度作為處理方式,只考慮了內(nèi)膜、浮環(huán)、外膜之間的傳熱狀況,而本文計(jì)算模型考慮了轉(zhuǎn)軸、浮環(huán)、軸承座之間的傳熱因素,并且,在本文進(jìn)一步分析中,考慮了浮環(huán)軸承潤(rùn)滑性能和環(huán)速比的相互影響關(guān)系,研究浮環(huán)結(jié)構(gòu)參數(shù)對(duì)其潤(rùn)滑性能和環(huán)速比的綜合影響。
在算例中,浮環(huán)軸承參數(shù)采用試驗(yàn)中軸承性能參數(shù)[4-6],并采用本文的模型進(jìn)行分析,獲取不同轉(zhuǎn)速下的環(huán)速比與文獻(xiàn)[6]相應(yīng)試驗(yàn)值的對(duì)比情況,見(jiàn)圖3。由圖3可見(jiàn),本文采取的潤(rùn)滑與傳熱相結(jié)合的模型算法結(jié)果與試驗(yàn)值吻合較好,誤差在15%內(nèi),且變化趨勢(shì)也較為一致,證明本文模型的可靠性。
表1為分析浮環(huán)軸承和潤(rùn)滑油主要參數(shù)。
表1 浮環(huán)軸承和潤(rùn)滑油主要計(jì)算參數(shù)
3.1 內(nèi)層間隙
圖4為內(nèi)層間隙對(duì)浮環(huán)軸承環(huán)速比與潤(rùn)滑特性的影響。由圖4a可知,隨內(nèi)層間隙增加,內(nèi)膜膜厚增加,導(dǎo)致內(nèi)膜力矩減小,故需減小環(huán)速比使內(nèi)膜力矩與外膜力矩重新達(dá)到平衡,因而環(huán)速比減小,且轉(zhuǎn)速越小,環(huán)速比下降趨勢(shì)越明顯。與內(nèi)層間隙0.02 mm時(shí)相比,轉(zhuǎn)速60 000 r/min時(shí),內(nèi)層間隙0.04 mm時(shí)的環(huán)速比減幅達(dá)23%。由圖4b可知,隨內(nèi)層間隙增加,內(nèi)膜通過(guò)端泄帶走更多熱量,從而使內(nèi)膜溫度減小。
由圖4c可知,外膜溫度也隨內(nèi)層間隙增加而減小,主要原因在于環(huán)速比減小使浮環(huán)轉(zhuǎn)速下降,導(dǎo)致浮環(huán)與液膜之間摩擦生熱明顯減小。由圖4d可知,隨內(nèi)層間隙增加,內(nèi)膜最小膜厚增加,而外膜最小膜厚的變化并不顯著,原因在于最小膜厚主要由間隙結(jié)構(gòu)決定,內(nèi)層間隙增加,浮環(huán)內(nèi)層幾何間隙明顯變大,因而內(nèi)膜最小膜厚增加;而外層幾何間隙沒(méi)有發(fā)生改變。由圖4e可知,總摩擦功耗隨內(nèi)層間隙增加而略有增加。由圖4f可知,隨內(nèi)層間隙增加,總端泄流量增加,但增幅基本不隨轉(zhuǎn)速變化。因此,在實(shí)際設(shè)計(jì)浮環(huán)軸承時(shí),應(yīng)在避免環(huán)速比過(guò)小的前提下,適當(dāng)增加內(nèi)層間隙從而提高浮環(huán)軸承的潤(rùn)滑性能。
在本算例中,盡管內(nèi)層間隙為0.04 mm時(shí)浮環(huán)具有最佳的潤(rùn)滑性能,但內(nèi)層間隙0.04 mm時(shí)過(guò)小的環(huán)速比可能會(huì)導(dǎo)致浮環(huán)實(shí)際工作中失效。綜上分析,要獲得理想的潤(rùn)滑性能和環(huán)速比,內(nèi)層間隙取0.03 mm較為合適。
3.2 外層間隙
圖5為外層間隙對(duì)浮環(huán)軸承環(huán)速比與潤(rùn)滑特性的影響。由圖5a可知,外層間隙0.06 mm的環(huán)速比均比外層間隙0.04 mm的環(huán)速比增加30%以上。由圖5b可知,外層間隙對(duì)內(nèi)膜溫度幾乎沒(méi)有影響。由圖5c可知,盡管環(huán)速比隨外層間隙增加而增加,導(dǎo)致浮環(huán)轉(zhuǎn)速升高,浮環(huán)與外膜摩擦產(chǎn)生更多熱量,但外膜間隙增加的同時(shí)也使液膜端泄帶走更多熱量,其影響大于外膜摩擦生熱影響,故外膜溫度反而減少,且隨轉(zhuǎn)速升高,外膜溫度減幅有擴(kuò)大趨勢(shì)。由圖5d可知,內(nèi)膜最小膜厚基本不隨外層間隙變化而變化,而外膜最小膜厚隨外層間隙增加而增加。由圖5f可知,由于外層間隙增加使外膜端泄流量增加,因而總端泄流量增加。綜上所述,增加外層間隙以提高浮環(huán)的環(huán)速比,同時(shí)還可降低浮環(huán)外膜溫度,但對(duì)內(nèi)膜溫度的影響并不明顯?;谕鈱娱g隙對(duì)浮環(huán)潤(rùn)滑特性和環(huán)速比的分析結(jié)果,在浮環(huán)其他幾何結(jié)構(gòu)參數(shù)不變的情況下,外層間隙取0.06 mm較為合適。
3.3 內(nèi)圓半徑
圖6為內(nèi)圓半徑對(duì)浮環(huán)軸承環(huán)速比和潤(rùn)滑特性的影響。由圖6a可知,隨內(nèi)圓半徑增加,環(huán)速比增加。由圖6b可知,由于內(nèi)圓半徑增加使浮環(huán)內(nèi)圓表面與液膜接觸面積增加,液膜摩擦做功增加,故內(nèi)膜溫度增加。由圖6c可知,外膜溫度隨內(nèi)圓半徑增加而增加,原因在于浮環(huán)轉(zhuǎn)速增加使外膜摩擦生熱增多,導(dǎo)致外膜溫度增加,且轉(zhuǎn)速越高,外膜溫度的增幅有擴(kuò)大趨勢(shì)。由圖6d可知,隨內(nèi)圓半徑增加,內(nèi)膜最小膜厚與外膜最小膜厚均增加,但外膜最小膜厚增幅明顯大于內(nèi)膜最小膜厚增幅,這是因?yàn)閮?nèi)膜膜厚受轉(zhuǎn)軸與浮環(huán)轉(zhuǎn)速綜合影響,而外膜膜厚只受浮環(huán)轉(zhuǎn)速影響。由圖6e可知,隨著內(nèi)圓半徑的增加,總摩擦功耗增加,同樣,總摩擦功耗增幅隨轉(zhuǎn)軸轉(zhuǎn)速增加而增加,說(shuō)明內(nèi)圓半徑對(duì)浮環(huán)軸承潤(rùn)滑性能與轉(zhuǎn)軸轉(zhuǎn)速有直接聯(lián)系。由圖6f可知,總端泄流量隨內(nèi)圓半徑增加而增加。綜上所述,由于內(nèi)、外膜溫度和摩擦功耗均隨內(nèi)圓半徑減小而減小,因此減小浮環(huán)內(nèi)圓半徑可以明顯改善浮環(huán)潤(rùn)滑性能,但浮環(huán)的環(huán)速比也相應(yīng)減小。內(nèi)圓半徑為6 mm時(shí),環(huán)速比均在0.1以下,不利于浮環(huán)的穩(wěn)定運(yùn)轉(zhuǎn)。因此,為確保浮環(huán)穩(wěn)定運(yùn)行,且同時(shí)確保浮環(huán)潤(rùn)滑性能最佳,內(nèi)圓半徑取7 mm較為合適。
3.4 外圓半徑
圖7為外圓半徑對(duì)浮環(huán)軸承環(huán)速比和潤(rùn)滑特性的影響。由圖7a可知,隨外圓半徑增加,環(huán)速比降低。由圖7b可知,內(nèi)膜溫度隨外圓半徑增加而略有增加,原因在于隨著外圓半徑增加,浮環(huán)轉(zhuǎn)速減小,而轉(zhuǎn)軸與浮環(huán)之間的相對(duì)速度增加,內(nèi)膜摩擦生熱加劇,導(dǎo)致內(nèi)膜溫度增加。由圖7c可知,由于外圓半徑增加使浮環(huán)外表面與液膜接觸面積增加,液膜做功增加,外膜溫度略有增加。但是,與內(nèi)圓半徑對(duì)內(nèi)、外膜溫度影響程度相比,外圓半徑對(duì)內(nèi)、外膜溫度的增幅很小。由圖7d可知,隨外圓半徑增加,外膜最小膜厚減小,而內(nèi)膜最小膜厚基本沒(méi)有變化。由圖7e、7f可知,隨外圓半徑增加,總摩擦功耗略有增加,總端泄流量略有減小,外圓半徑對(duì)內(nèi)膜溫度、外膜溫度、總摩擦功耗和總端泄流量變化幅度均在5%以內(nèi),說(shuō)明外圓半徑對(duì)浮環(huán)潤(rùn)滑性能的影響并不顯著。由于外圓半徑為11和12 mm時(shí),環(huán)速比較小,為確保浮環(huán)穩(wěn)定運(yùn)轉(zhuǎn),外圓半徑應(yīng)取10 mm較為合適,此時(shí)浮環(huán)在不影響潤(rùn)滑性能的同時(shí)也能夠保證較大的環(huán)速比,從而保證浮環(huán)軸承的可靠性。此外,浮環(huán)外圓半徑對(duì)浮環(huán)軸承潤(rùn)滑性能的影響程度小于內(nèi)圓半徑對(duì)其的影響。因此,想要提高潤(rùn)滑性能,實(shí)際設(shè)計(jì)浮環(huán)軸承時(shí),調(diào)整浮環(huán)內(nèi)圓半徑比調(diào)整外圓半徑更有效。
1)增壓器浮環(huán)軸承的環(huán)速比和潤(rùn)滑特性存在互相影響、互相作用關(guān)系。實(shí)際設(shè)計(jì)浮環(huán)軸承時(shí),應(yīng)綜合考慮結(jié)構(gòu)參數(shù)對(duì)潤(rùn)滑特性和環(huán)速比的影響及影響程度。
2)與內(nèi)層間隙0.02 mm相比,內(nèi)層間隙0.04 mm時(shí)的環(huán)速比減幅達(dá)23%;內(nèi)層間隙增加,內(nèi)膜溫度和外膜溫度減小,摩擦功耗略有增加,端泄流量增加,浮環(huán)潤(rùn)滑性能提高。要獲得理想的環(huán)速比,內(nèi)層間隙取0.03 mm較為合適。
3)外層間隙為0.06 mm的環(huán)速比均比外層間隙為0.04 mm的環(huán)速比增加30%以上;外膜溫度減小,且轉(zhuǎn)速越高,外膜溫度減幅越大。
4)浮環(huán)軸承內(nèi)圓半徑越小,浮環(huán)潤(rùn)滑性能越好,但浮環(huán)的環(huán)速比也相應(yīng)降低。
5)隨浮環(huán)外圓半徑增加,環(huán)速比降低,但內(nèi)膜溫度、外膜溫度、總摩擦功耗和總端泄流量變化幅度均在5%以內(nèi),潤(rùn)滑性能變化不顯著;為提高潤(rùn)滑性能,實(shí)際設(shè)計(jì)浮環(huán)軸承時(shí),調(diào)整浮環(huán)內(nèi)圓半徑比調(diào)整外圓半徑更有效。
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Effect of structural parameters on lubrication performance of floating ring bearing and ring speed ratio in turbocharger
Li Jiaqi1, Ni Jimin1, Shi Xiuyong1, Xu Xiaochuan1, Liu Yue1, Li Dongdong1,Chen Zhenbin2
(1.,,201804,; 2.,,571737,)
Turbochargers are widely used in internal combustion engines. The floating ring bearing is the most important part in the turbocharger, which is composed of inner film and outer film. The floating ring bearing can reduce the frictional power and oil film temperature. In the past, the lubrication performances of the floating ring bearing were analyzed without considering the heat transfer among the shaft, floating ring and bearing block. Furthermore, some researchers analyzed the lubrication performance of floating ring and the ring speed ratio separately. In reality, there exists heat transfer among the shaft, floating ring and bearing block and this part of heat will directly influence the lubrication performance of the floating ring. Besides, the change of the structural parameters may decrease ring speed ratio, which maybe is unfavorable to the practical operation if it is too small, and the change of the ring speed ratio may also influence the lubrication performance. So there exists a strong coupled relationship between the lubrication performance and ring speed ratio, and they can be studied more comprehensively and systematically with the change of structural parameters. The lubrication performance of floating ring bearing and the floating ring speed ratio in turbocharger were studied by considering the heat transfer among the shaft, floating ring and bearing block. Based on the Reynolds equation and the floating ring balance equation, the lubrication model of floating ring bearing was established. The Reynolds equation was solved by the finite difference method. The thermal deformation of the floating ring bearing was calculated by the thermal deformation equation. Comparative analysis on the influence of structural parameters on lubrication performance of floating ring bearing and floating ring speed ratio was performed, and the parameters included inner film clearance, outer film clearance, inner circle radius and outer circle radius. Results showed the effect and affecting level of structural parameter on lubrication performance and floating ring speed ratio should be considered comprehensively. The ring speed ratio decreased with the increase of inner film clearance. Compared with that inner film clearance is 0.02 mm, the ring speed ratio decreased 23% when inner film clearance is 0.04 mm. The inner film temperature, the outer film temperature decreased and the total friction power loss slightly increased with the increase of inner film clearance. The lubrication performance and ring speed ratio are ideal when inner film clearance is 0.03 mm. Compared with that outer film clearance is 0.04 mm,the ring speed ratio increased more than 30%. The outer film temperature decreased with the increase of outer film clearance. The higher the rotation speed, the greater the decrease amplitude of the outer film temperature. The smaller the inner circle radius, the smaller the floating ring speed ratio, and the smaller the inner film temperature, outer film temperature and total friction power loss. So the lubrication performance of the floating ring bearing will be improved with the decrease of the inner circle radius. The ring speed ratio decreased with the increase of the outer circle radius. But the inner film temperature, the outer film temperature increased, the total friction power loss and total oil leakage flowrate changed less than 5% with the increase of the outer circle radius. The outer circle radius has little effect on the lubrication performance of the floating ring bearing. Compared with the adjustment of the outer circle radius, it is more effective to improve the lubrication performance of floating ring bearing with the adjustment of the inner circle radius. There is a good agreement between the results predicted by the calculation model in this paper and some published experimental data.
bearings; models; temperature; floating ring bearing; lubrication; structural parameters; ring speed ratio; heat transfer
10.11975/j.issn.1002-6819.2017.02.007
TH133.31
A
1002-6819(2017)-02-0048-08
2016-05-09
2016-12-07
國(guó)家自然科學(xué)基金資助(51166002);上海市自然基金資助(16ZR1438500)
李佳琪,博士生,主要從事浮環(huán)軸承潤(rùn)滑機(jī)理研究。上海 同濟(jì)大學(xué)汽車學(xué)院,201804。Email:lijiaqi_1987@126.com
倪計(jì)民,男,教授,博士生導(dǎo)師,主要從事渦輪增壓器潤(rùn)滑系統(tǒng)設(shè)計(jì)研究。上海 同濟(jì)大學(xué)汽車學(xué)院,201804。Email:njmwjyx@hotmail.com
李佳琪,倪計(jì)民,石秀勇,徐曉川,劉 越,李冬冬,陳振斌. 結(jié)構(gòu)參數(shù)對(duì)增壓器浮環(huán)軸承潤(rùn)滑特性和環(huán)速比的影響[J]. 農(nóng)業(yè)工程學(xué)報(bào),2017,33(2):48-55. doi:10.11975/j.issn.1002-6819.2017.02.007 http://www.tcsae.org
Li Jiaqi, Ni Jimin, Shi Xiuyong, Xu Xiaochuan, Liu Yue, Li Dongdong, Chen Zhenbin. Effect of structural parameters on lubrication performance of floating ring bearing and ring speed ratio in turbocharger[J]. Transactions of the Chinese Society of Agricultural Engineering (Transactions of the CSAE), 2017, 33(2): 48-55. (in Chinese with English abstract) doi:10.11975/j.issn.1002-6819.2017.02.007 http://www.tcsae.org
農(nóng)業(yè)工程學(xué)報(bào)2017年2期