1、某轿车动力装置参数的匹配设计英文翻译毕业论文某轿车动力装置参数的匹配设计英文翻译毕业论文AbstractThe availability of pressure information of a hydraulic actuator makes it possible to improve the quality of vehicle power transmission via precise feedback control and to realize on-board fault diagnosis. However, the high cost of a pressure senso
2、r has not allowed its widespread deployment despite such apparent advantages. This paper presents an observer-based algorithm to estimate the pressure output of a hydraulic actuator in a vehicle power transmission control system. The proposed algorithm builds on more readily available slip velocity
3、and the models of a hydraulic actuator and a mechanical subsystem. The former is obtained empirically via system identification due to the complexity of the hydraulic actuator, while the latter is derived physically. The resulting robust observer is guaranteed to be stable against possible parametri
4、c variations and torque estimation errors. The hardware in-the-loop studies demonstrate the viability of the proposed algorithm in the field of advanced vehicle power transmission control and fault diagnosis. C 2002 Elsevier Science Ltd. All rights reserved. Received 1 March2001; accepted 27 August
5、2001Keywords: Robust observer; System identification; Hydraulic actuator; Vehicle power transmission control system; Hardware-in-the-loop simulation1.IntroductionAutomatic and continuously variable transmission systems have been expanding their presence in passenger vehicles in recent years,which ha
6、s naturally prompted active research on vehicle power transmission control systems. The main topics include shift control algorithm (Shin, Hahn, Yi, & Lee, 2000a; Zheng, Srinivasan, & Rizzoni, 1999; Shin, Hahn, & Lee, 2000b), feedback control of torque converter clutch slip systems (Jauch, 1999; Hib
7、ino, Osawa, Yamada, Kono, & Tanaka, 1996; Hahn & Lee, 2000), new hydraulic circuits for control performance enhancement (Jung, Cho, & Lee, 2000), etc. Despite the extensive research effort on control algorithms, it appears that the pressure information of a hydraulic actuator has not been fully util
8、ized in vehicle power transmission control, largely due to the high cost of a pressure sensor. As a result, most practical controllers have been largely built upon the mechanical subsystem only, while neglecting the dynamics of a hydraulic actuator.Instead of directly measuring the pressure output o
9、f a hydraulic actuator, this paper proposes an indirect alternative to estimate the pressure output: an observer- based approach. The main thrust of this paper is that the pressure output of a hydraulic actuator is observable with the slip velocity measurement of the mechanical subsystem in a vehicl
10、e power transmission control system. In addition to the readily available slip velocity measurement, the observer design requires the models of hydraulic and mechanical subsystems whose accuracy directly impacts the observer performance. The mechanical subsystem is physically modeled with relative e
11、ase. The complexity of the hydraulic actuator dynamics does not allow a physical model amenable to observer design. Instead, system identification yields a simplified empirical model of the hydraulic actuator. Actual observer design focuses on guaranteeing robust stability against parametric variati
12、ons and torque estimation errors that are bound to occur in a mechanical power transmission control system. Hardware- in-the-loop simulation studies are conducted to examine the performance of the robust observer-based estimator for the hydraulic actuator pressure, which shows the viability of the p
13、roposed approach. The outcome is a robust pressure estimator that relies on the readily available slip velocity measurement only and thus has a potential to be widely employed in vehicle power transmission control and fault diagnosis.This paper is organized as follows. Section 2 derives a physical m
14、odel for the mechanical subsystem. Simplified empirical models are developed for a hydraulic actuator in Section 3. Section 4 deals with the observer design. The performance of the designed observer is examined in Section 5.2. Overview of a vehicle power transmission control systemA vehicle power tr
15、ansmission control system typically consists of two subsystems, a mechanical subsystem and a hydraulic actuator. The input and output of the system under consideration are the voltage signal to the hydraulic actuator and the slip velocity between the friction elements in the mechanical subsystem, re
16、spectively. The hydraulic actuator drives the friction elements and generates the slip velocity of the mechanical subsystem according to its pressure output. Fig. 1 shows a vehicle power transmission control system considered in this paper, a torque converter clutch slip control system. The mechanic
17、al subsystem consists of an engine, a torque converter, an automatic transmission with planetary gear sets, and wheels with a final reduction gear. The torque converter clutch generates friction torque acting upon the engine according to the hydraulic actuator pressure, which in turn determines the
18、slip velocity between the engine and the turbine of the torque converter at a desired target value.In order to derive a physical model of the mechanical subsystem, the power transmission at each stage is examined. At the very first stage, the engine torque is transmitted to the impeller and is balan
19、ced by the reaction torque of the impeller and the friction torque from the torque converter clutch. The torque converter amplifies and transmits the impeller torque to the turbine. The turbine torque drives the automatic transmission system together with the friction torque of the torque converter
20、clutch, while the driving load torque of the vehicle provides additional resistive force. Denoting the slip velocity between the engine and the turbine as the output of interest (y) results in Hahn & Lee (2000)where Ie is the equivalent rotational inertia of the engine, Iv the equivalent rotational
21、inertia of the vehicle, We the angular velocity of the engine, Wt the angular velocity the turbine, Te the engine torque, Tp the impeller torque, Tt the turbine torque, Tc the friction torque of the torque converter clutch, Tl the driving load torque, c the equivalent damping constant of the torque
22、converter clutch, rt the gear ratio of the automatic transmission, rf the gear ratio of the final reduction, m the friction coefficient of the torque converter clutch, Ro the outer radius of the torque converter clutch, Ri the inner radius of the torque converter clut chand Pc the pressure output of
23、 the hydraulic actuator, It is worth noting that the quantities inEq. (1) are subject to errors; the damping constant is not exactly known; absence of torque sensors in a commercial vehicle entails torque estimation errors; the equivalent rotational inertia of the vehicle varies as the number of pas
24、sengers changes, and only a rough bound on the friction coefficient is available.3. Identification of a hydraulic actuator3.1. Motivation for empirical modelingFig. 2 shows the hydraulic actuator considered in this paper. It basically consists of three elements: a PWM type solenoid valve, a pressure
25、 modulator valve and a pressure control valve. The first and second regulating valves, not shown in Fig.2, regulates the main pressure of the entire hydraulic circuit. The pressure modulator valve further decreases the output pressure of the first regulating valve (channel #7) to a lower pressure le
26、vel.The output pressure of the pressure modulator valve at channel #9 is always regulated around 4.0 bar in the steady-state by means of the feedback chamber #10 (Hahn, 1999). The voltage signal from the transmission control unit (TCU) drives the PWM-type solenoid valve so that the pressure at chann
27、el #11 assumes values between 0 bar and 4 bar. The pressure at channel #11 acts on the spool of the pressure control valve, which in turn generates engage/disengage pressures for the friction element in the mechanical subsystem. Since the engage/disengage pressures are applied to the same friction e
28、lement from the opposite sides, the difference between engage and disengage pressures may be regarded as the output of the hydraulic actuator. A nonlinear mathematical model of the hydraulic actuator has been obtained in Hahn (1999) using the Newtons second law of motion. Although the nonlinear mode
29、l in Hahn (1999) matches the experimental results to a certain extent, it has some drawbacks when applied to the observer design problem considered in this paper:1. The model order is too high (E10).2. The governing differential equations are too stiff to numerically solve in real-time.3. There exis
30、t numerous unknown parameters that need to be estimated or tuned in order to obtain reasonable match between experimental results and model predictions.A possible alternative is to capture the dynamics essential to the design of a nonlinear observer and to obtain a lower-order control-oriented empir
31、ical model. In this paper, two empirical models are proposed based on the system identification (Ljung, 1999).1. Nonlinear.2. Mostly smooth, although its behavior is qualitatively different at a couple of points, i.e. nearly discontinuous.3. Not one-to-one when the duty cycle is around 60%. 4. Not e
32、ven symmetric with respect to origin; 60% duty cycle approximately corresponds to 0 bar pressure output.These observations raise an immediate concern that the brute-force application of system identification (Ljung, 1999) may utterly fail to give a high-fidelity model if a conventional model structu
33、re, e.g., ARX (auto-regressive with exogenou s input) is adopted. In the steady-state, an ARX model approximates the nonlinear mapping in Fig. 3 by a straight line passing through the origin, by virtue of its linearity. A better approach would be to shift the origin to a point around which the inputoutput mapping is approximately sy
