外文翻譯--自動手動變速器的變速控制

1、<p><b>　　中文3177字</b></p><p>　　畢業(yè)設(shè)計（論文）外文文獻(xiàn)翻譯</p><p>　　IEEE/ASME TRANSACTIONS ON MECHATRONICS, VOL. 11, NO. 1, FEBRUARY 2006 </p><p>　　Gearshift Control for Automate

2、d Manual Transmissions</p><p>　　Luigi Glielmo, Member, IEEE, Luigi Iannelli, Member, IEEE, Vladimiro Vacca,</p><p>　　and Francesco Vasca, Member, IEEE</p><p>　　Abstract—A gearshift

3、control strategy for modern automated manual transmissions (AMTs) with dry clutches is proposed. The controller is designed through a hierarchical approach by discriminating among five different AMT operating phases: eng

4、aged, slipping-opening, synchronization, go-to-slipping, and slipping-closing. The control schemes consist of decoupled and cascaded feedback loops based on measurements of engine speed, clutch speed, and throwout bearin

5、g position, and on estimation of the tr</p><p>　　Index Terms—Automated manual transmissions (AMTs), automotive control, clutch engagement control, dry clutch, gearshift. </p><p>　　I. INTRODUCTIO

6、N</p><p>　　CARS with modern transmission systems exhibit high fuel economy, low exhaust emission, and excellent driveability. Recent reports on the future automotive market forecast that in 2010 the producti

7、on of manual transmissions will have fallen below 50% while the modern automatic transmissions will have reached 25% of production [1], [2]. Among other responces, the automated manual transmissions (AMTs) represent a pr

8、omising solution since they can be considered as an inexpensive add-on solution for </p><p>　　One of the most critical operations in AMTs is represented by the gearshift and more specifically by the clutch e

9、ngagement. In automotive drivelines, the goal of the clutch is to smoothly connect two rotating masses, the flywheel and the transmission shaft, that rotate at different speeds, in order to allow the transfer of the torq

10、ue generated by the engine to the wheels through the driveline. The automation of the clutch engagement must satisfy different and conflicting objectives: It should ob</p><p>　　In order to achieve the object

11、ives of the clutch engagement automation, several control approaches which deal with the vehicle startup operating conditions have been proposed: quantitative feedback theory [5], model predictive control strategy [6], f

12、uzzy control [7], decoupling control [4], and optimal control [8], further in [9], the authors propose a particular engagement technique. Problems and solutions related to the clutch engagement during the gearshift phase

13、 have been also considered in </p><p>　　In spite of the extensive literature on AMT control, some problems still need further investigation: the role of speed feedback loops in the clutch engagement control,

14、 the definition of a controller architecture which can be exploited both during vehicle startup and gearshift, the robustness of the solution with respect to clutch aging, and uncertainties in the clutch characteristic.

15、This paper tries to provide a contribution in this direction by proposing a new controller for gearshift and clut</p><p>　　The paper is organized as follows. In Section II, models of driveline, dry clutch, a

16、nd closed-loop electrohydraulic actuator are considered and tuned on experimental data. In Section III, five different operating phases of the AMT are considered: engaged, slipping-opening, synchronization, go-to-slippin

17、g, and slipping-closing. The controllers, designed through a hierarchical approach with decoupled and cascaded feedback loops based on measurements of clutch speed, engine speed, and throwout bear</p><p>　　I

18、I. MODELING</p><p>　　A. Driveline </p><p>　　A driveline model suitable for parameter identification and for the clutch engagement control design can be obtained assuming the clutch speed wc is e

19、qual to the mainshaft speed wm, and considering the mainshaft rigid (Fig. 1). Thus, when the engine flywheel and the clutch disk are in slipping operating conditions, the driveline model can be written as </p><

20、;p><b>　　(1) </b></p><p><b>　　(2)</b></p><p><b>　　(3)</b></p><p>　　Fig.1.Drivelinescheme.</p><p><b>　　(4)</b></p>

21、;<p>　　where J's are inertias; ω's speeds;T's torques; the throwout bearing position; and the subscripts e,c,t, and w indicate engine, clutch, transmission, and wheels, respectively. Moreover, ig is th

22、e gear ratio,id is the differential ratio, Jeq(ig,id)=Jm+(1/im²)(Js1+Js2+(Jt/ id²)), Js1 and Js2 are the inertias of the two disks connected to the synchronizer, Jm is the mainshaft inertia, is the load torque

23、, θ’s are angles, k’s are elastic stiffness coefficients, and β’s are friction coefficients. Whe</p><p>　　A further reduction of the model can be obtained by also considering the driveshaft to be rigid. By a

24、ssuming and by reporting the vehicle inertia to the mainshaft, one obtains the following model: </p><p><b>　　(5) </b></p><p><b>　　(6)</b></p><p>　　Where ,Th

25、e corresponding engaged model can be obtained by adding (5) and (6) with . </p><p>　　The model (1)–(4) provides a good compromise between description of the driveline dynamics and model complexity. The param

26、eters of (1)–(4) have been tuned from experimental data carried out on a FIAT STILO 2.4 gasoline car with , , and the set of gear ratios from the first gear to the fifth gear, respectively. The signals have been acqui

27、red with a sampling frequency of 100 Hz during tests in which a series of upshifts were carried out with different acceleration pedal positions and vehicle sp</p><p><b>　　(7) </b></p><

28、p>　　where is obtained through the so-called dirty derivative, i.e., a filtered incremental ratio on the measured engine speed. The model parameters have been identified by using the least square method and the corresp

29、onding results have been found as follows: , ,, and . Note that in the sum , the dependence of on the gear ratio is negligible. It can be also verified that, as typical for the type of driveline and car under investiga

30、tion, the first resonance in the frequency response of the model ap</p><p>　　B. Dry Clutch </p><p>　　From a physical point of view, the dry clutch consists of two disks (the clutch disk connecte

31、d to the mainshaft and the flywheel disk connected to the engine) covered with a high friction material and a mechanism which presses the disks against each other (clutch closed or engaged) or keeps them apart (clutch op

32、en or disengaged). During an engagement phase, the clutch disk is moved towards the flywheel disk until the friction due to their contact allows the torque transmission. The throwout bear</p><p>　　Moreover,

33、the clutch characteristic is also influenced by the dependence of the friction coefficient on both temperature [14] and slip speed. In particular, the negative variations of the friction coefficient with slip speed can i

34、nduce torsional self-excited vibrations of the driveline [15], [16]. </p><p>　　The nominal nonlinear characteristic has been identified from the experimental tests described above. By representing the estim

35、ated values of the clutch torque obtained through (7) as a function of the corresponding values of the signal , the set of points reported in Fig. 2 (top diagram) is obtained. The absolute value of the clutch torque has

36、been modeled by using the interpolation curve reported in bottom diagram of Fig. 2 (curve b), and its variations (curves a and c) which model different </p><p>　　C. Clutch Actuator </p><p>　　In

37、AMTs, the electrohydraulic actuator is mainly composed of a hydraulic piston connected to a system of springs that keep the clutch closed when the piston does not apply any force (see Fig. 3). The piston is controlled by

38、 a three-port electrovalve that regulates the oil flow through the hydraulic circuit and then determines the force on the mechanical actuator. In standard commercial </p><p>　　applications, the electrohydrau

39、lic actuator is used with a feedback control on the throwout bearing position. Usually a further inner control loop on the current is implemented. By identifying the parameters of the detailed actuator model proposed in

40、[17], we now show that, for the goal of this paper, the actuator with a position feedback loop can be satisfactorily approximated by a first-order linear system. </p><p>　　Fig. 2. Set of points ( ) obtained

41、from experimental data (top diagram) and clutch characteristics (bottom diagram) for positive slip speed with different wear. Curve a: new clutch. Curve b: medium wear. Curve c: high wear.</p><p>　　Fig. 3. H

42、ydraulic actuator scheme corresponding to the clutch engaged.</p><p>　　The actuator model consists of a set of equations describing the dynamics of the electrovalve spool, the servocylinder piston, and the p

43、ressure variation in the mechanical actuator chamber. The motion of the valve spool is described by a force balance equation </p><p><b>　　(8)</b></p><p>　　where is the spool mass,

44、is the spool position, is the force caused by the solenoid current , is the friction coefficient, Kv is the effective spring constant, and is the mechanical spring force due to the preload of the spring. The oil orce

45、Foil is the Bernoulli’s force defined as [18] </p><p><b>　　(9) </b></p><p>　　where is the discharging coefficient, is the velocity coefficient, is the control port width, is the

46、jet angle described by a polynomial function of the spool position , and the variation of pressure in the three-port electrovalve is defined as </p><p>　　where is the line pressure, is the tank pressure, a

47、nd is the oil pressure in the actuator chamber. The variation of the oil pressure can be described by </p><p><b>　　(10)</b></p><p>　　where E is the fluid bulk modulus, is the pist

48、on cross sectional area, is the actuator position (or, equivalently, the throwout bearing position), is the minimum actuator chamber volume (corresponding to =0), is the oil flow in the servocylinder computed as show

49、n by the equation at the bottom of the page(now I have move them here,as follows:</p><p>　　where d is the underlap to the port, Clk is the leakage coefficient, and ρ is the oil density. The last equation of

50、the hydraulic actuator model describes the motion of the servocylinder piston as </p><p><b>　　(11) </b></p><p>　　where Mp and Mc are respectively the piston and the clutch mass, bp i

51、s the friction coefficient, and Fspring is the nonlinear spring force of the diaphragm spring modeled through a static characteristic. For further details on the model, see [17]. </p><p>　　自動手動變速器的變速控制</p

52、><p>　　Luigi Glielmo, Member, IEEE, Luigi Iannelli, Member, IEEE, Vladimiro Vacca,</p><p>　　and Francesco Vasca, Member, IEEE</p><p>　　摘要——提出了一個適用于干式離合器的現(xiàn)代自動手動變速器的變速控制略。該控制器利用分級方法，通過五個

53、不同的AMT操作階段：接合、滑移、同步、滑動、分離而設(shè)計。該控制方案由解耦和級聯(lián)反饋回路組成，它是基于發(fā)動機轉(zhuǎn)速、離合器速度的測量值及計算主、從動盤間間隙并估計傳送扭矩。動力傳動系統(tǒng)、干式離合器和控制執(zhí)行器模型用以中型汽車實驗數(shù)據(jù)的估計，并用實驗數(shù)據(jù)通過仿真檢查該控制器的有效性。</p><p>　　關(guān)鍵詞——自動手動變速器、自動控制、離合器接合控制、干式離合器、變速器</p><p>&

54、lt;b>　　第1章 引言</b></p><p>　　具有現(xiàn)代變速器系統(tǒng)的汽車具有高燃油經(jīng)濟性，低廢氣排放和卓越的駕駛操縱性。關(guān)于未來汽車市場的最新研究報告顯示：在2010年，手動變速器的產(chǎn)量將減少50%，而自動變速器產(chǎn)量將達(dá)到變速器總產(chǎn)量的25% [1] [2]。 在其他的研究中, 自動手動變速器（AMTs）由于被看作是一種廉價的附加解決方案為傳統(tǒng)的手動變速系統(tǒng)改進(jìn)提供了一項可行性方案。此

55、外, AMTs 也被廣泛應(yīng)用于賽車和作為配置單元應(yīng)用于混合動力車。</p><p>　　自動手動變速器的最關(guān)鍵問題之一是換擋和更具體的離合器的接合。在汽車動力傳動系統(tǒng)中，離合器的功用是柔和地接合兩個旋轉(zhuǎn)質(zhì)量塊，飛輪和傳動軸，即在它們以不同的速度旋轉(zhuǎn)的情況下，允許將發(fā)動機產(chǎn)生的扭矩通過動力傳動系統(tǒng)傳遞給車輪。離合器接合自動化必須滿足不同和相互沖突的目標(biāo)：它應(yīng)該至少獲得與司機手動操作所能達(dá)到的相同性能（極短的換擋時間

56、和舒適性）并提高排放性能和磨損性能。在離合器接合到鎖住期間，發(fā)動機和離合器的轉(zhuǎn)速對舒適性和摩擦損失起著重要作用[3] [4]。</p><p>　　為了實現(xiàn)離合器接合自動化的目標(biāo)，幾個涉及汽車啟動操作條件的控制方法被提了出來：定量反饋理論[5]，模型預(yù)測控制策略[6]，模糊控制[7]，去耦控制[4]和優(yōu)化控制[8]，更進(jìn)一步地[9]，作者提出了一種特別的接合技術(shù)。換擋階段離合器接合的相關(guān)問題和解決方案也在本文中被

57、考慮。在文獻(xiàn)[10]，提出了一個計算加速和減速過程中所需的發(fā)動機轉(zhuǎn)速的分析性程序。在文獻(xiàn)[11]，一種基于模型的反演方法被用于那些沒有同步裝置的自動手動變速器的換擋控制設(shè)計。在文獻(xiàn)[12]，應(yīng)用了一種考慮駕駛員的意圖和可變負(fù)載的神經(jīng)模糊方法。 In [11], a model-based backstepping methodology is used to design the gearshift control in AMTs wi

58、thout the synchronizer.</p><p>　　盡管有大量關(guān)于AMT控制的文獻(xiàn)，有些問題還需要進(jìn)一步調(diào)查：在離合器接合控制中速度反饋回路的作用，定義一個在汽車啟動和換擋過程都能被應(yīng)用的控制架構(gòu)，關(guān)于離合器老化的耐用性方案，和離合器特性的不確定性。本文試圖提出一個用于自動手動變速器換擋和離合器接合的新的控制器進(jìn)而在這一方面有所貢獻(xiàn)。</p><p>　　本文的結(jié)構(gòu)如下：

59、在第二章，介紹動力傳動系統(tǒng)模型，干式離合器和閉環(huán)電動液壓執(zhí)行器和實驗數(shù)據(jù)調(diào)整。在第三章，分析AMT的五個不同操作階段：接合，滑移，同步，滑動，分離。在第四章，提出一種由分層方法設(shè)計的控制器，該控制由解耦和級聯(lián)反饋回路組成，它是基于對離合器轉(zhuǎn)速、發(fā)動機轉(zhuǎn)速的測量值和計算軸承位置。該控制的AMT是在Matlab仿真環(huán)境下模擬的，其中對應(yīng)于當(dāng)前AMT的階段和相應(yīng)的控制器的動態(tài)仿真模塊由一個狀態(tài)流的有限狀態(tài)機選中。在第五章，通過仿真結(jié)果表明該方

60、法的有效性。在第六章，綜合本文結(jié)果，得出結(jié)論。</p><p><b>　　第2章 建模</b></p><p>　　2.1 動力傳動系統(tǒng)</p><p>　　假設(shè)離合器速度與主軸速度相等，并且假定主軸為剛性，則可以獲得一個適用于參數(shù)辨識和離合器接合控制設(shè)計的動力傳動系統(tǒng)模型（如圖1所示）。因此，當(dāng)發(fā)動機飛輪和離合器摩擦盤處于滑移狀態(tài)時，該傳動

61、系統(tǒng)模型可寫為：</p><p><b>　　(1) </b></p><p><b>　　(2)</b></p><p><b>　　(3)</b></p><p><b>　　(4)</b></p><p>　　其中J是轉(zhuǎn)動慣量

62、；ω表示角速度；T表示轉(zhuǎn)矩；是離合器主、從動盤間間隙；下標(biāo)e，c，t，和w分別表示發(fā)動機，離合器，變速器和車輪。此外，是齒輪傳動比，是差速器傳動比，，和是連接到同步器的兩個摩擦片的轉(zhuǎn)動慣量， 是主軸轉(zhuǎn)動慣量，表示負(fù)載轉(zhuǎn)矩，表示轉(zhuǎn)角， k表示彈性剛度系數(shù)，β表示摩擦系數(shù)。當(dāng)離合器接合時，發(fā)動機轉(zhuǎn)速和離合器盤速度相等。假定，將(1) 式與 (2) 式相加，可以得到相應(yīng)的接合模型。欲了解更多詳情，請參閱[13]。其中一個更復(fù)雜的模型也已得到，

63、該模型還考慮到了主軸的靈活性。 </p><p>　　圖1 動力傳動系統(tǒng)結(jié)構(gòu)</p><p>　　考慮主軸為剛性，可以進(jìn)一步獲得一個簡化模型。假設(shè) ，并計算汽車主軸轉(zhuǎn)動慣量，得到以下模型：</p><p><b>　　(5) </b></p><p><b>　　(6)</b></p>

64、<p>　　其中。假定，將(5) 式與 (6) 式相加，可以得到相應(yīng)的接合模型。</p><p>　　模型（1）—（4）很好地協(xié)調(diào)了傳動系統(tǒng)動力學(xué)描述和模型復(fù)雜性之間的問題。模型（1）—（4）的參數(shù)已經(jīng)從菲亞特斯蒂洛2.4升汽油車的實驗數(shù)據(jù)進(jìn)行了調(diào)整，該車型的參數(shù)為，，從第一檔到第五檔的齒輪比分別為。 已經(jīng)以100赫茲的采樣頻率獲得，在測試期間進(jìn)行了一系列不同加速踏板位置和車輛速度的加減。離合器轉(zhuǎn)矩

65、由（1）推導(dǎo)估計得到： </p><p><b>　　(7) </b></p><p>　　其中是通過求導(dǎo)獲得的，即對測量發(fā)動機轉(zhuǎn)速經(jīng)過篩選的增量比。該模型參數(shù)已經(jīng)利用最小二乘法進(jìn)行了確定，并且發(fā)現(xiàn)如下的相應(yīng)結(jié)果： Jc+Jeq=0.004 kgm²，，及。注意到的的和， Jeq對齒輪比的依賴可以忽略?？梢宰C實的是，作為動力傳動系統(tǒng)和調(diào)研車型的典型，該模型的

66、一階共振頻率響應(yīng)出現(xiàn)在幾赫茲的范圍內(nèi)。</p><p><b>　　2.2 干式離合器</b></p><p>　　從物理的角度來看，干式離合器由兩個摩擦盤（連接到主軸的從動盤和連接到發(fā)動機的飛輪摩擦盤）組成，它具有高摩擦材料和一個使它們相互壓緊（離合器接合）或保持它們分開（離合器脫離）的機構(gòu)。在接合階段，離合器從動盤向飛輪摩擦盤移動，直到接觸摩擦至允許扭矩傳輸。主、

67、從動盤間間隙決定了兩個摩擦盤之間的壓力，進(jìn)而決定了滑移階段的扭矩傳遞。與主、從動盤間間隙相關(guān)的非線性特性與通過離合器傳遞的轉(zhuǎn)矩不易模擬。離合器磨損大大影響了其特性和扭矩傳遞。離合器的特性還因摩擦系數(shù)對溫度 [14] 和滑動速度的依賴性而受到影響。特別地，摩擦系數(shù)與滑動速度的不良變化會引起傳動系統(tǒng)扭轉(zhuǎn)自激振動。[15] [16]。</p><p>　　標(biāo)稱非線性特性Tc(xc)已由上述實驗測試得到確定。將通過（7）

68、獲得的離合器扭矩的估計值代表信號的相應(yīng)值，得到圖2（上圖）的點集。離合器扭矩的絕對值已通過使用插值曲線得到，見圖2下圖曲線b，以及模擬不同離合器磨損成時的變化（曲線a和c），并已用于測試控制器的魯棒性。注意，磨損狀況改變了軸承的位置，例如，當(dāng)兩個摩擦盤接觸時，對同一值，傳輸?shù)呐ぞ赜闪汩_始達(dá)到不同的值。</p><p>　　圖2 由實驗數(shù)據(jù)獲得的點集（）（上圖）和對應(yīng)于不同磨損時離合器的絕對滑行速度特性（下圖）。

69、曲線a：新的離合器。曲線b：中等磨損。曲線c：高磨損。</p><p>　　2.3 離合器制動器</p><p>　　在自動手控變速器中, 電控液壓傳動制動器主要由一個液壓活塞連接到彈簧組組成，以此保證當(dāng)液壓活塞不受力時保持離合器閉合（見圖3）。該活塞由一三端口電磁閥控制，它通過液壓回路調(diào)節(jié)油流并確定作用于機械制動器上的力。在標(biāo)準(zhǔn)的商業(yè)應(yīng)用中，電控液壓執(zhí)行器和反饋控制一起被用于主、從動盤間

70、間隙的控制。當(dāng)前，通常使用進(jìn)一步內(nèi)部控制回路。通過對文獻(xiàn)[17]中提到的精細(xì)的制動器模型參數(shù)的比較，我們現(xiàn)在表明，為了本文的目標(biāo)，一階線性系統(tǒng)可以完美的接近帶有位置反饋回路的制動器。</p><p>　　圖3 對應(yīng)于離合器接合的液壓傳動裝置方案</p><p>　　執(zhí)行器模型由一組描述電磁閥閥芯動態(tài)，伺服活塞和機械傳動裝置室壓力變化的方程組組成。閥芯動力是通過一個力平衡方程描述的： &l

71、t;/p><p><b>　　(8) </b></p><p>　　其中是閥芯質(zhì)量， 是閥芯閥芯位移， 螺線管電流產(chǎn)生的力，是摩擦系數(shù)，是有效彈簧常數(shù)， 是由彈簧預(yù)緊力產(chǎn)生的機械彈簧力。油壓力 伯努利力，定義為[18] </p><p><b>　　(9) </b></p><p>　　其中是放電系數(shù)，

72、是速度系數(shù)，是控制端口寬度，是由一個閥芯位置多項式函數(shù)描述的噴射角度，三端口電磁閥的壓力變化定義為 </p><p>　　其中是線壓力，是罐內(nèi)壓力，是制動器室油壓。油壓的變化可以表示為：</p><p><b>　　(10)</b></p><p>　　其中是液體體積模量，是活塞的橫截面積，是執(zhí)行器的位移（相當(dāng)于主、從動盤間間隙），是最小器室容

73、積（對應(yīng)于），是伺服油流，其計算方程如本頁底部方程組所示（因篇幅改動，摘抄如下：</p><p><b>　　，</b></p><p>　　其中d是the underlap to the port，Clk是泄露系數(shù)，ρ是油的密度。)液壓執(zhí)行器模型的最后一個方程描述了伺服缸活塞的運動，如下所示： </p><p><b>　　(11)