Transmission control for power-shift agricultural tractors Design.pdf

Transmission control for power-shift agricultural tractors: Design and end-of-line automatic tuning Mara Tanelli a, Giulio Panzania, Sergio M. Savaresia, Carlo Pirolab aDipartimento di Elettronica e Informazione, Politecnico di Milano, Piazza L. da Vinci, 32, 20133 Milano, Italy bSAME Deutz-Fahr Group, Viale F. Cassani, 15, 24047 Treviglio (Bergamo), Italy a r t i c l ei n f o Article history: Received 24 May 2010 Accepted 14 November 2010 Available online 8 December 2010 Keywords: Power-shift transmission Agricultural tractors Automotive systems End-of-line tuning a b s t r a c t This paper addresses the analysis and design of the transmission control system for a high-power power-shift agricultural tractor. Specifi cally, all the criticalities involved with the correct management of both single clutch and double clutch gear shifts are investigated, and a control system capable of providing good shifting performance in all operating conditions is proposed. Further, to comply with components tolerances and spreads in the production line, an automatic procedure for the end-of-line tuning of the transmission control system is proposed to objectively classify the quality of the gear shift and 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 motivation Agricultural vehicles have to cope with working conditions which are more complex and demanding than those experienced by other ground vehicles, 10. In fact, agricultural vehicles are essentially designed to work at low speed while providing large traction forces. Moreover, their ease of moving on uneven soil makes them suitable also for heavy trailers transportation. To en- sure the maximum fl exibility of use at each speed and to exploit the maximum engine power available in all working conditions, nowadays agricultural vehicles are often equipped with a so-called power-shift transmission. This kind of transmission has a large number of gears available (typically from 9 to 30) and it allows to perform a gearshift with no (or at least with a minimum) loss of power from the engine to the driving wheels. Usually, a power-shift transmission is characterized by the presence of two or more (depending from the number of gears and the overall mechanical architecture of the gearbox) wet clutches connected to an hydraulic circuit, whose pressure can be regulated by a proportional solenoid valve. Considering the large number of gears available and the fact that to achieve an optimal gear 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 the scientifi c literature, some works dealing with power-shift or dual clutch transmissions control for ground vehicles are available, see e.g., 38,15, but very little has been done on specifi c solutions for agricultural tractors. This is mainly due to the fact that agricul- tural vehicles have very specifi c performance specifi cations due to the very broad range of working conditions and variability of the vehicle 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 and the comfort of the driver on all working grounds, which vary from asphalt roads to rough off-road terrains. Also the load distribution in tractors is much different than for other vehicles, due to the fact that it might be due to either front or rear additional loads due to the various working instruments that need to be employed for differen tasks. Finally, note also that the variation of the operating conditions is most often non measurable via on-board sensors, and thus asks for robust and easily tunable gear shift controllers. These facts make the problem of ensuring an optimal and repeatable gear shift on an agricultural tractor a very challenging task. To design an effective transmission control system, fi rst of all the most signifi cant variables which infl uence the gear shift quality must be identifi ed, see e.g., 2,16. Further, the gear shift control system has to optimally manage the trade-off among the following confl icting requirements: (i) yield comfortable gear shifts; (ii) guarantee that no loss of power to the driving wheels occurs during gear shifts; (iii) cause a minimum wear and tear of mechanical components over the life of the vehicle transmission. 0957-4158/$ - see front matter ? 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.mechatronics.2010.11.006 Corresponding author. Tel.: +39 02 2399 3621; fax: +39 02 2399 3412. E-mail address: tanellielet.polimi.it (M. Tanelli). Mechatronics 21 (2011) 285297 Contents lists available at ScienceDirect Mechatronics journal homepage: Moreover, in the industrial context, once the control design phase is accomplished and the control system is implemented into fi nal products, an end-of-line tuning phase is usually scheduled to deal with constructive tolerances and production spreads which cause the fi nal system to be different from the prototype one used for control validation and testing. Hence, this phase is tailored to optimize the controller parameters so as to guarantee that the expected gear shifting performance is achieved on all vehicles. Usually, this phase is carried out by human testers, who tune the controller parameters based on personal driving preferences and experience. Thus, is it clear that end-of-line tuning is a crucial and diffi cult 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 qualifi ed as comfortable by one operator, but not by another one: this means that the fi nal tuning can lead to very different gear shift behaviors on different vehicles of the same type. Note that, as the vehicle handling qualities, of which the gear shift characteristics are a signifi cant 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 to achieve customers satisfaction and to promote customers loyalty to the brand. Moreover, another signifi cant advantage of the pro- posed approach is that of reducing the industrial costs associated with end-of-line tuning by reducing the number of gear shifts needed 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 specifi c application, has a validity which goes beyond the considered problem, as the aforementioned design steps constitute a working paradigm which can be applied in many different production contexts. As a matter of fact, this pa- per is one of the fi rst contributions which aims at formalizing the end-of-line tuning of industrial applications endowed with control systems, proposing a systematic approach to the considered prob- lem. In this respect, the results in 13,16 offer other applications of the proposed methodology and address the problem of quantifying of the driving style and safety via measured data, and of designing and 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 systematic approach presented in this work, which is constituted by the fol- lowing steps: ? an evaluation of the characteristic features which defi ne the quality of the considered system; ? an experimental sensitivity analysis to single out the relation between the features to be optimized and the measurable variables; ? the defi nition of the cost functions; ? the design of the control algorithm and of the procedures for its end-of-line tuning grounded on the cost functions optimization. This methodology makes the results in the present paper of general interest for all those applications in which a control system must be designed and tuned while dealing with the dispersion coming from production spreads and tolerances which make the underlying plant (i.e., the fi nal vehicle) different from that used for design purposes. The resulting research area requires tools both of control theory and optimization, combined with the specifi c application-domain knowledge. The presented results are based on a joint research work between the Politecnico di Milano and the R (2) the onoff status of each directional valve. To execute a gear shift with a power-shift transmission, the outgoing clutch must be brought to zero pressure, whereas the incoming clutch must be brought to maximum pressure. Note that a non-power-shift gear shift would disengage the outgoing clutch and then engage the incoming one. In so doing, there is a time interval in which the vehicle is in a neutral state, and no engine torque can reach the driving wheels. In agricultural vehicles the neutral state must be avoided, as the large load forces would cause the vehicle to stop. Thus, it is of utmost importance to ensure a continuous torque transfer to the driving wheels during the gear shift, which is the main characteristic of a power-shift gear shift. To conclude the system description, Table 1 shows the nine available gears together with the associated engaged clutches. As can 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 gear shifts as single clutch gear shifts. However, when the 34 and 67 gear shifts are considered, two clutches must be changed (one HML gearbox 123 gearbox Motion Inverter Mode selector Fig. 2. Schematic view of the power-shift transmission. Kiss-pointEngagePressure bar Normal force N Fig. 3. Oil pressure in the clutch as a function of the normal force. H M L1 2 3 Fig. 4. Simplifi ed hydraulic scheme of the considered transmission. Table 1 Available gears and respective engaged clutches. GearLMH123 1? 2? 3? 4? 5? 6? 7? 8? 9? M. Tanelli et al./Mechatronics 21 (2011) 285297287 belonging to the HML and one to the 123 gearbox), making the de- sign of the gear shift controller more complex, as will be shown subsequently. We refer to such gear shifts as double clutch gear shifts. 3. Gear shift quality assessment As discussed in Section 1, defi ning an objective gear shift quality assessment yields the following advantages: ? it provides a unique and objective indication of gear shift per- formance, helpful to compare different vehicles and/or different control algorithms. ? it makes the end-of-line tuning phase easier and cheaper by relying on the automatic optimization of appropriate perfor- mance indexes. The crucial issue to deal with in defi ning the most suitable cost functions is that of determining meaningful relationships between measured signals and gear shift comfort and quality. Several stud- ies have been carried out in the Automotive context, showing good results in evaluating comfort via acceleration measurements, see e.g., 9,14. For the type of vehicle considered in this work, it is easy to understand that this kind of signal is not suitable, as soil irreg- ularities cause measurement noise which shadows the actual gear shift 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 relationships between gear shift quality and vehicle speed, whose measurement is commonly available via wheel encoders. As discussed in 16, this signal can be exploited to provide satisfactory comfort evaluation. 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 fi rst one (Fig. 5a) has been classifi ed as good, whereas the last two (Fig. 5b and c) as medium and bad, respectively. The speed behavior in the three considered gear shifts is as follows: the speed always starts from a constant value and increases (up-shifts have been performed in all cases) until it reaches a higher fi nal value, which depends only on the fi nal gear ratio as the engine speed is kept fi xed and constant during the gear shift. What really makes the 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 the speed increases with a smooth ramp, in the medium quality gear shift the speed increase is only piecewise linear (see dotted oval box in Fig. 5b) and shows a signifi cant 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 been defi ned as J Varvmt ?vreft ?; 1 wherevm(t) is the measured wheel speed andvref(t) is a reference signal to be designed, which describes the speed behavior in an optimal gear shift. The measured vehicle speedvm(t) is computed as vmt 1 4 X 4 i1 xitri;2 wherexi(t), i = 1, . , 4 are the wheel rotational speeds measured via the wheel encoders and ri, i = 1, . , 4 are the wheel radii. The reference signalvref(t) has been designed as composed of three different parts (see Fig. 6). The fi rst is defi ned by the constant speed value at the beginning of the manoeuvre, i.e., vref1vmtreq;3 where treqis the time instant at which the gear shift is requested by the driver. The reference speed in the last part of the manoeuvre is also constant and computed as vrefendxengtreqrsInc:;4 wherexeng(treq) is the engine speed at the beginning of the gear shift (recall that the engine speed is fi xed and constant during the gear shift), r is the average wheel radius andsincis the transmission ratio of the incoming gear (also known when the gear shift is issued by the driver). The reference speed evolution in time between these limiting speed levels, which defi nesvref2, is chosen as linear, yielding vref2t vref1 vrefend?vref1 t2? t1 t ? t1;5 where the time instant t1 is defi ned as t1:jvmt1 ?vref1j P 0 ? fjvmt ?vref1j t1g:6 Namely, t1is the last time instant at which the measured speed is lower than the initial reference speedvref1, while the time instant t2 is defi ned as t2:jvmt1 ?vrefendj 0;8t tDO,HMLbut the two incoming clutches may not engage simultaneously. With the proposed control approach, the fact that the engagement phase occurs in a correct way is evaluated by means of the cost functions, therefore without a direct tuning of the engagement instant. The obtained results are reported in Fig. 11, which shows the values of J1and J2, respectively, as functions of Overlap and Delay- HML. For the sake of conciseness, the analysis for the KP pressure value is not shown, as the obtained results are similar to those dis- cussed for the single clutch gear shift. By inspecting Fig. 11, some remarks can be made. ? First of all note that each controller parameter, i.e., Overlap and DelayHML, has a predominant effect over one single index. Namely, Overlap mostly affects the J1index, whereas DelayHML the J2one. This fact allows to decouple the optimization phase, and to consider two successive single variable optimization -1000-500050010001500 0 2 4 6 8 10 12 14 P kiss-point pressure J1 index 1-2 2-3 0100200300400 0 0.5 1 1.5 2 2.5 3 Overlap ms J1 index 1-2 2-3 mbar Fig. 8. Sensitivity analysis on single shifts controller parameters. 012345678 5.5 6 6.5 7 7.5 8 8.5 Time s Speed km/h Not optimized Optimized Fig. 9. Time histories of the measured vehicle speed in a single clutch shift: results with non-optimized (dashed line) and with optimized (solid line) controller parameters. tt Fig. 10. Double clutch gearshift control: master pressure profi le (solid line) and outgoing clutches disengagement (dashed lines). M. Tanelli et al./Mechatronics 21 (2011) 285297291 problems, which can more easily managed in view of the auto- matic end-of-line tuning phase. Specifi cally, as will be described in more detail in Section 5, the optimal value for Overlap will be found by minimizing J1, while DelayHML will be tuned accord- ing to J2 . Specifi cally, to assess the correctness of the sequential minimization of the performance indexes, one has to observe that the function J2(?, DelayHML) has the same shape for all val- ues of Overlap. Therefore, once a value of Overlap has been fi xed by optimizing J1, then the optimization of J2done by varying the value of DelayHML will lead to a fi nal value for J2which is approximately always the same (note also that the cost function always decreases as the value of DelayHML increases indepen- dently of the value of Overlap). Of course, the fi nal value of J2 will not be rigorously the same irrespectively of the value of Overlap with which it is evaluated and which is determined by the minimization of J1 , but the small differences in the fi nal value for J2 are not practically relevant as they yield no signifi - cant changes in the fi nal gear shift performance. ? Note further that the results in Fig. 11b seem to suggest that a large value of DelayHML would bring good quality gear shifts. Nonetheless, it is worth pointing out that, as DelayHML increases, the clutches are left slipping for an increasing amount of time. Thus, this parameter should be kept at the lowest pos- sible value so as to prevent an excessive wear of the clutches. Section 5 will better discuss how to effectively deal with this issue. ? Finally, it is worth comparing the results obtained in the single and double clutch gear shif