大學生方程式純電動賽車（總體設計）

大學生方程式純電動賽車（總體設計）,畢業設計,設計 Transmission control for power-shift agricultural tractors:Designand end-of-line automatic tuningMara Tanellia,Giulio Panzania,Sergio M.Savaresia,Carlo PirolabaDipartimento di Elettronica e Informazione,Politecnico di Milano,Piazza L.da Vinci,32,20133 Milano,ItalybSAME Deutz-Fahr Group,Viale F.Cassani,15,24047 Treviglio(Bergamo),Italya r t i c l ei n f oArticle history:Received 24 May 2010Accepted 14 November 2010Available online 8 December 2010Keywords:Power-shift transmissionAgricultural tractorsAutomotive systemsEnd-of-line tuninga b s t r a c tThis paper addresses the analysis and design of the transmission control system for a high-powerpower-shift agricultural tractor.Specifically,all the criticalities involved with the correct managementof both single clutch and double clutch gear shifts are investigated,and a control system capable ofproviding good shifting performance in all operating conditions is proposed.Further,to comply withcomponents tolerances and spreads in the production line,an automatic procedure for the end-of-linetuning of the transmission control system is proposed to objectively classify the quality of the gear shiftand automatically optimize it.The suitability of the proposed approach is thoroughly tested on an instru-mented vehicle.?2010 Elsevier Ltd.All rights reserved.1.Introduction and motivationAgricultural vehicles have to cope with working conditionswhich are more complex and demanding than those experiencedby other ground vehicles,10.In fact,agricultural vehicles areessentially designed to work at low speed while providing largetraction forces.Moreover,their ease of moving on uneven soilmakes them suitable also for heavy trailers transportation.To en-sure the maximum flexibility of use at each speed and to exploitthe maximum engine power available in all working conditions,nowadays agricultural vehicles are often equipped with a so-calledpower-shift transmission.This kind of transmission has a largenumber of gears available(typically from 9 to 30)and it allowsto perform a gearshift with no(or at least with a minimum)lossof power from the engine to the driving wheels.Usually,a power-shift transmission is characterized by thepresence of two or more(depending from the number of gearsand the overall mechanical architecture of the gearbox)wetclutches connected to an hydraulic circuit,whose pressure can beregulated by a proportional solenoid valve.Considering the largenumber of gears available and the fact that to achieve an optimalgear shift it is necessary to correctly manage several control vari-ables,this kind of transmission needs to be properly controlled.The design of such a control system is not a trivial task.In thescientific literature,some works dealing with power-shift or dualclutch transmissions control for ground vehicles are available,see e.g.,38,15,but very little has been done on specific solutionsfor agricultural tractors.This is mainly due to the fact that agricul-tural vehicles have very specific performance specifications due tothe very broad range of working conditions and variability of thevehicle load,which make the gear shift optimal performance defi-nition different from that of ground vehicles.As a matter of fact,the main constraints are the repeatability of the manoeuvre andthe comfort of the driver on all working grounds,which vary fromasphalt roads to rough off-road terrains.Also the load distributionin tractors is much different than for other vehicles,due to the factthat it might be due to either front or rear additional loads due tothe various working instruments that need to be employed fordifferen tasks.Finally,note also that the variation of the operatingconditions is most often non measurable via on-board sensors,andthus asks for robust and easily tunable gear shift controllers.Thesefacts make the problem of ensuring an optimal and repeatable gearshift on an agricultural tractor a very challenging task.To design an effective transmission control system,first of allthe most significant variables which influence the gear shift qualitymust be identified,see e.g.,2,16.Further,the gear shift controlsystem has to optimally manage the trade-off among the followingconflicting requirements:(i)yield comfortable gear shifts;(ii)guarantee that no loss of power to the driving wheels occursduring gear shifts;(iii)cause a minimum wear and tear of mechanical componentsover the life of the vehicle transmission.0957-4158/$-see front matter?2010 Elsevier Ltd.All rights reserved.doi:10.1016/j.mechatronics.2010.11.006Corresponding author.Tel.:+39 02 2399 3621;fax:+39 02 2399 3412.E-mail address:tanellielet.polimi.it(M.Tanelli).Mechatronics 21(2011)285297Contents lists available at ScienceDirectMechatronicsjournal homepage: the industrial context,once the control designphase is accomplished and the control system is implemented intofinal products,an end-of-line tuning phase is usually scheduled todeal with constructive tolerances and production spreads whichcause the final system to be different from the prototype one usedfor control validation and testing.Hence,this phase is tailored tooptimize the controller parameters so as to guarantee that theexpected gear shifting performance is achieved on all vehicles.Usually,this phase is carried out by human testers,who tune thecontroller parameters based on personal driving preferences andexperience.Thus,is it clear that end-of-line tuning is a crucialand difficult phase to deal with.As a matter of fact,since no objec-tive indexes to evaluate the gear shift performance and comfort ex-ist,a gear shift can be qualified as comfortable by one operator,butnot by another one:this means that the final tuning can lead tovery different gear shift behaviors on different vehicles of the sametype.Note that,as the vehicle handling qualities,of which the gearshift characteristics are a significant component,is often consid-ered as a trademark of the single manufacturer,the ability of deliv-ering vehicles with identical manoeuvre features can be a key toachieve customers satisfaction and to promote customers loyaltyto the brand.Moreover,another significant advantage of the pro-posed approach is that of reducing the industrial costs associatedwith end-of-line tuning by reducing the number of gear shiftsneeded to tune each vehicle and by making the process automatic,thus not requiring highly experienced operators to perform it.It is worth noting that the approach presented in this work,even though tailored to a specific application,has a validity whichgoes beyond the considered problem,as the aforementioneddesign steps constitute a working paradigm which can be appliedin many different production contexts.As a matter of fact,this pa-per is one of the first contributions which aims at formalizing theend-of-line tuning of industrial applications endowed with controlsystems,proposing a systematic approach to the considered prob-lem.In this respect,the results in 13,16 offer other applications ofthe proposed methodology and address the problem of quantifyingof the driving style and safety via measured data,and of designingand objectively tuning the motion inversion control of an agricul-tural tractor,respectively.Although being different problems with respect to that consid-ered herein,both these works share(all or part of)the systematicapproach presented in this work,which is constituted by the fol-lowing steps:?an evaluation of the characteristic features which define thequality of the considered system;?an experimental sensitivity analysis to single out the relationbetween the features to be optimized and the measurablevariables;?the definition of the cost functions;?the design of the control algorithm and of the procedures for itsend-of-line tuning grounded on the cost functions optimization.This methodology makes the results in the present paper ofgeneral interest for all those applications in which a control systemmust be designed and tuned while dealing with the dispersioncoming from production spreads and tolerances which make theunderlying plant(i.e.,the final vehicle)different from that usedfor design purposes.The resulting research area requires tools bothof control theory and optimization,combined with the specificapplication-domain knowledge.The presented results are based on a joint research workbetween the Politecnico di Milano and the R&D Department ofthe SAME Deutz-Fahr Group(SAME,Lamborghini,Deutz-Fahr,Hrlimann,Adim Diesel and Deutz AG).The work has been focusedon a power-shift transmission designed for high-power(200 HP)agricultural tractors(see Fig.1).The first effort has been devoted to define appropriate costfunctions which allow an objective evaluation of gear shift comfortand quality.Then,an accurate analysis of all relevant gear shiftdynamics has led to design a simple but effective transmissioncontrol strategy.Further,to obtain the best possible gear shift per-formance on every production vehicle,an automatic tuning phaseis proposed which guarantees satisfactory and repeatable gear shiftperformance.The structure of the paper is as follows.Section 2 provides adescription of the power-shift transmission system,both from anhydraulic and a mechanical viewpoint.Section 3 is focused on pre-senting the performance indexes which have been selected tojudge the gear shift quality.In Section 4,the proposed controlstrategy for power-shift gear shifts is described,both for the caseof single clutch and double clutch gearshifts,together with the re-sults of an experimental sensitivity analysis of the performance in-dexes with respect to the controller parameters.Finally,Section 5is devoted to describe the end-of-line self-tuning procedure and topresent related experimental results.2.System descriptionThe overall mechanical layout of the considered power-shifttransmission is depicted in Fig.2.As can be seen,the power flowsfrom the engine(on the left of Fig.2)towards the driving wheelsthrough two different gearboxes:the High-Mean-Low(HML)group,composed of three different gears and the 123 group,whichalso comprises three different gear ratios.The transmission is com-pleted with two other components,namely the motion inverterand the mode selector.The motion inverter(see 11,16)is an elec-tro-hydraulic system,constituted by two clutches,which allows toperform an automatic motion inversion,i.e.,it takes the vehiclefrom a,say,forward speed to a reverse speed with the driver sim-ply acting on a lever.The mode selector allows to choose amongthree different working modes:creep,work and transport,whichcan be varied only when the tractor is at standstill.In what follows,we concentrate on the control of the gear shift,and consider thetwo gearboxes only,assuming that no motion inversion is occur-ring(note,in passing,that during a motion inversion the drivercannot command a gear shift),and that a fixed mode has beenengaged.As the two gearboxes are in series,nine transmission ratios be-tween the engine and the driving wheels are available(disregard-ing the final differential,whose ratio is fixed).Conceptually,although mechanically different,the two gearboxes can be treatedequally for control design purposes.Each gear is associated with aFig.1.The tractor employed in this work.286M.Tanelli et al./Mechatronics 21(2011)285297wet clutch:to select a particular gear the corresponding clutchmust be completely engaged,so that the torque coming from theengine can be completely transferred via the clutch itself.Thewet clutches handled in this work are multi-plate wet:in orderto be engaged(and hence to select the associated gear)the surfacesof the plates must be in close contact and the normal force they ex-change must be large enough to develop a friction force whichguarantees that no relative slip occurs between them.Fig.3 shows a schematic view of the physical relationship be-tween normal force and clutch oil pressure.For gear shift analysis,three different zones must be considered.Starting from zero(i.e.,from atmospheric pressure),an increase in the pressure bringsno changes in the normal force between the plates,which remainsnegligible.When the so-called kiss-point pressure is reached(see Fig.3),this distance that separates the plates has been covered and thesurfaces come into contact.From here over,the normal force in-creases proportionally to the pressure.During this phase,thefriction force between the surfaces allows to transfer a certainamount of incoming torque through the clutch but,as there is anon-zero relative slip between the plates,the gear ratio isindefinite.Once the engage pressure is reached(refer again to Fig.3),thenormal force is large enough to annihilate any relative slip be-tween the plates of the clutch and a precise gear ratio can bedefined.Fig.4 shows a schematic view of the hydraulic scheme,whichallows to understand how the pressure in a clutch can be con-trolled.There are six ONOFF directional valves which connecteach clutch to the master hydraulic pressure,regulated by a pro-portional solenoid valve.Such hydraulic architecture yields the fol-lowing clutch pressure behavior:when a directional valve isswitched ON the pressure in the clutch is equal to the master pres-sure.Being this a one-way valve,note that the pressure can onlyincrease,even if the master pressure decreases.Conversely,whenthe directional valve is switched OFF the pressure in the clutch iszero.As such,the available control variables are the following:(1)the master hydraulic pressure.Note that,as no pressure sen-sors are available,the real control variable is the currentdriving the proportional valve.Such a variable can be linkedto the output pressure via a static map.In what follows,wewill regard the pressure as control variable,keeping in mindthat the aforementioned conversion from current to pres-sure has to be performed;(2)the onoff status of each directional valve.To execute a gear shift with a power-shift transmission,theoutgoing clutch must be brought to zero pressure,whereas theincoming clutch must be brought to maximum pressure.Note thata non-power-shift gear shift would disengage the outgoing clutchand then engage the incoming one.In so doing,there is a timeinterval in which the vehicle is in a neutral state,and no enginetorque can reach the driving wheels.In agricultural vehicles theneutral state must be avoided,as the large load forces would causethe vehicle to stop.Thus,it is of utmost importance to ensure acontinuous torque transfer to the driving wheels during the gearshift,which is the main characteristic of a power-shift gear shift.To conclude the system description,Table 1 shows the nineavailable gears together with the associated engaged clutches.Ascan be seen,usually a gear shift requires to change only one clutch(i.e.,the one belonging to the HML gearbox).We refer to such gearshifts as single clutch gear shifts.However,when the 34 and 67gear shifts are considered,two clutches must be changed(oneHML gearbox 123 gearboxMotion InverterMode selector Fig.2.Schematic view of the power-shift transmission.Kiss-pointEngagePressure bar Normal force NFig.3.Oil pressure in the clutch as a function of the normal force.H M L1 2 3Fig.4.Simplified hydraulic scheme of the considered transmission.Table 1Available gears and respective engaged clutches.GearLMH1231?2?3?4?5?6?7?8?9?M.Tanelli et al./Mechatronics 21(2011)285297287belonging to the HML and one to the 123 gearbox),making the de-sign of the gear shift controller more complex,as will be shownsubsequently.We refer to such gear shifts as double clutch gearshifts.3.Gear shift quality assessmentAs discussed in Section 1,defining an objective gear shift qualityassessment yields the following advantages:?it provides a unique and objective indication of gear shift per-formance,helpful to compare different vehicles and/or differentcontrol algorithms.?it makes the end-of-line tuning phase easier and cheaper byrelying on the automatic optimization of appropriate perfor-mance indexes.The crucial issue to deal with in defining the most suitable costfunctions is that of determining meaningful relationships betweenmeasured signals and gear shift comfort and quality.Several stud-ies have been carried out in the Automotive context,showing goodresults in evaluating comfort via acceleration measurements,seee.g.,9,14.For the type of vehicle considered in this work,it is easyto understand that this kind of signal is not suitable,as soil irreg-ularities cause measurement noise which shadows the actual gearshift contributions to vehicle accelerations.Moreover,accelerome-ters are not standard sensors to have on-board of agricultural trac-tors.Thus,we concentrated on investigating the relationshipsbetween gear shift quality and vehicle speed,whose measurementis commonly available via wheel encoders.As discussed in 16,this signal can be exploited to provide satisfactory comfortevaluation.To understand the rationale behind the quality index design,Fig.5 shows the time histories of the vehicle speed in three differ-ent gear shifts,whose performance was judged by an expert driver:the first one(Fig.5a)has been classified as good,whereas the lasttwo(Fig.5b and c)as medium and bad,respectively.The speedbehavior in the three considered gear shifts is as follows:the speedalways starts from a constant value and increases(up-shifts havebeen performed in all cases)until it reaches a higher final value,which depends only on the final gear ratio as the engine speed iskept fixed and constant during the gear shift.What really makesthe gear shifts different is the smoothness with which the speed in-creases.Note,in fact,that while in the good gear shift in Fig.5a thespeed increases with a smooth ramp,in the medium quality gearshift the speed increase is only piecewise linear(see dotted ovalbox in Fig.5b)and shows a significant initial undershoot.Finally,the bad gear shift is characterized by a quite irregular speed behav-ior and large oscillations(see dotted oval box in Fig.5c).Based on these considerations,the performance index has beendefined asJ Varvmt?vreft?;1wherevm(t)is the measured wheel speed andvref(t)is a referencesignal to be designed,which describes the speed behavior in anoptimal gear shift.The measured vehicle speedvm(t)is computed asvmt 14X4i1xitri;2wherexi(t),i=1,.,4 are the wheel rotational speeds measuredvia the wheel encoders and ri,i=1,.,4 are the wheel radii.The reference signalvref(t)has been designed as composed ofthree different parts(see Fig.6).The first is defined by the constantspeed value at the beginning of the manoeuvre,i.e.,vref1vmtreq;3where treqis the time instant at which the gear shift is requested bythe driver.The reference speed in the last part of the manoeuvre isalso constant and computed asvrefendxengtreqrsInc:;4wherexeng(treq)is the engine speed at the beginning of the gearshift(recall that the engine speed is fixed and constant during thegear shift),r is the average wheel radius andsincis the transmissionratio of the incoming gear(also known when the gear shift is issuedby the driver).The reference speed evolution in time between these limitingspeed levels,which definesvref2,is chosen as linear,yieldingvref2t vref1vrefend?vref1t2?t1t?t1;5where the time instant t1is defined ast1:jvmt1?vref1j P 0?fjvmt?vref1j t1g:6Namely,t1is the last time instant at which the measured speed islower than the initial reference speedvref1,while the time instantt2is defined ast2:jvmt1?vrefendj 0;8t tDO,HMLbut the two incoming clutchesmay not engage simultaneously.With the proposed controlapproach,the fact that the engagement phase occurs in a correctway is evaluated by means of the cost functions,therefore withouta direct tuning of the engagement instant.The obtained results are reported in Fig.11,which shows thevalues of J1and J2,respectively,as functions of Overlap and Delay-HML.For the sake of conciseness,the analysis for the KP pressurevalue is not shown,as the obtained results are similar to those dis-cussed for the single clutch