6DCT250_INGLES.pdf

HOCHSCHULE KONSTANZ TECHNIK, WIRTSCHAFT UND GESTALTUNG UNIVERSITY OF APPLIED SCIENCES

hofer-pdc GmbH

MASTER THESIS Simulation Based Comfort Evaluation for Vehicles with Automated Transmissions

Ahmad Hakim Mohd Sorihan Automotive Systems Engineering Engineering 28.02.2013 Supervisors: Prof. Dr.-Ing Uwe Kosiedowski Dr. Mathias Lutz

Abstract In the last few years, the design variation of automated transmission is becoming more and more diverse. Some examples besides the well-known automatic transmission with torque converters and planetary gears are the Dual Clutch Transmission (DCT) and the Automated Manual Transmission (AMT), to name a few. These transmission variations are further divided according to their realisation concept, such as the Dry Dual Clutch and Wet Dual Clutch Transmission. The very diverse design of a transmission causes different driving experience and influences the driving comfort. This comfort perception is evaluated in a subjective way by the driver. The aim of this master thesis is to reproduce comfort-relevant driving situations in simulation models and to evaluate the driving situations with both proven and newly defined evaluation criteria. The evaluation steps and result obtaining were automated with programming scripts for convenience. The long term aim of this thesis is to provide a knowledge of simulation based comfort evaluation.

In den letzten Jahren wächst bei den automatisierten Getrieben die Vielfalt der Getriebetypen. Neben den bekannten Wandlerautomatgetrieben mit Planetenradsätzen sind das automatisierte Schaltgetriebe (AMT) und das Doppelkupplungsgetriebe (DCT) zu nennen. Zu den verschiedenen Getriebetypen existieren verschiedene Realisierungskonzepte, wie z.B. Getriebe mit nasser und mit trockener Doppelkupplung. Die verschiedenen

Realisierungskonzepte und Getriebetypen verursachen unterschiedliches

Fahrerlebnis. Das Fahrerlebnis wird subjektiv von Fahrer wahrgenommen. Das Ziel dieser Arbeit ist die

verschiedenen

komfortrelevanten

Fahrsituationen

realistisch

in

Simulationsmodellen

nachzubilden und die Situationen mit sowohl bewährten als auch mit neu entwickelten Kriterien zu bewerten. Die Bewertungsschritte wurden durch programmierte Skripte automatisiert. Das langfristige Ziel dieser Arbeit ist das Bereitstellen von Kenntnissen für die simulationsbasierte Komfortbewertung.

Ahmad Hakim Mohd Sorihan

i

Declaration of Confidentiality We, the University of Applied Sciences Konstanz, hereby acknowledge and agree to comply that this master thesis entitled “Simulation Based Comfort Evaluation for Vehicles with Automated Transmissions”

and the all the information contained in this thesis are not to be revealed to a third person or made public without the written approval of hofer-pdc GmbH.

hofer-pdc GmbH

Prof. Dr. –Ing. Uwe Kosiedowski

Stuttgart,

Konstanz,


ii

Declaration of Originality I hereby declare that this master thesis entitled “Simulation Based Comfort Evaluation for Vehicles with Automated Transmissions”

submitted as the final thesis of the master program Automotive Systems Engineering of University of Applied Sciences Konstanz is written on my own and not made use of the work of any other party or students past or present without a cknowledgement, except those indicated by referencing.

____________________ Stuttgart, 28.02.2012 Ahmad Hakim Mohd Sorihan


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Acknowledgement This master thesis for the final thesis of the Master course Automotive Systems Engineering would not have been possible without the generous support and guidance of several individuals who in one way or another contributed their valuable assistance in the preparation and completion of this study. First and foremost, I would like to express my gratitude to my supervisors Prof. Dr. –Ing. Uwe Kosiedowski of HTWG Konstanz and Dr Mathias Lutz of hofer-pdc GmbH for the opportunity as well as the continuous assistance and supervision during my 5 month Master Thesis at hofer-pdc GmbH. I would also like to thank Mr Jens Schäfer and Mr Matteo Mocchi for the assistance in regards to the software AMESim and DIAdem as well for the helpful advises for this thesis. My gratitude also goes to Ms Elke Gamper and Ms Julia Hendrich for the help in reviewing this thesis. Finally, I would like to thank the whole Simulation Department of hofer-pdc GmbH for the comfortable and friendly atmosphere from the start till the end of my master thesis.


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Table of Contents 1

2

3

Introduction .................................................................................................................................... 1 1.1

Motivation............................................................................................................................... 1

1.2

Scope of Work ......................................................................................................................... 2

Simulation Software Used............................................................................................................... 3 2.1

LMS Imagine.Lab AMESim....................................................................................................... 3

2.2

DIAdem ................................................................................................................................... 5

Theoretical Foundations ................................................................................................................. 6 3.1

4

5

Automotive Transmission and Powertrain ............................................................................. 6

3.1.1

Dual Clutch Transmission ................................................................................................ 7

3.1.2

Powertrain .................................................................................................................... 12

3.2

Subjective Evaluation of Driving Situation ............................................................................ 13

3.3

Objectification of Comfort Criteria ....................................................................................... 14

3.3.1

Driving Capability vs. Driving Tasks ............................................................................... 14

3.3.2

Driving Situations and the Respective Comfort Evaluation Criteria ............................. 15

3.3.3

Summary of the Driving Situations ............................................................................... 23

Simulation Model Setup and Parameterisation ............................................................................ 25 4.1

Overview of the Complexity of the Simulation Model ......................................................... 25

4.2

Simulation Components in AMESim ..................................................................................... 26

4.3

Reference Car........................................................................................................................ 27

4.4

Reference Transmission: Getrag Powershift 6DCT250 ......................................................... 29

4.4.1

Dry Dual Clutches .......................................................................................................... 29

4.4.2

Electromechanical Actuator of the Dual Clutches ........................................................ 31

4.4.3

Gears and Gear Actuators ............................................................................................. 34

4.5

Reference Engine: 1.6 Ti-VCT ................................................................................................ 38

4.6

Control System ...................................................................................................................... 40

4.6.1

Launch / Moving Off ..................................................................................................... 42

4.6.2

Upshift ........................................................................................................................... 45

4.6.3

Downshift ...................................................................................................................... 48

Evaluation of Results..................................................................................................................... 54 5.1

Script/Apps for Evaluation of Results ................................................................................... 54

5.2

Evaluation of Simulation Results........................................................................................... 55

5.2.1

Launch/Moving Off ....................................................................................................... 55


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5.2.2

Upshift ........................................................................................................................... 65

5.2.3

Downshift ...................................................................................................................... 76

5.3

Comparison of Simulated Driving Situations with the Real Measurement Data .................. 82

6

Conclusions and Future Improvements ........................................................................................ 85

7

Reference Index ............................................................................................................................ 87

8

Appendix ....................................................................................................................................... 89 8.1

AMESim Submodels Used in Simulation ............................................................................... 89

8.2

Table for Subjective Evaluation of Driving Situations ........................................................... 93

8.3

Simulation Model Basis ......................................................................................................... 95

8.4

App Interfaces ....................................................................................................................... 98

8.5

Python Code Snippets ......................................................................................................... 102


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List of Figures Figure 1: LMS Imagine.Lab AMESim ....................................................................................................... 3 Figure 2: App Designer in AMESim ......................................................................................................... 4 Figure 3: Gear ratio ................................................................................................................................. 6 Figure 4: Schematic of a dual clutch transmission [1] ............................................................................ 7 Figure 5: Dry dual clutch with external torsion damper (left) and friction disk integrated damper (right) [2] ................................................................................................................................................. 8 Figure 6: Electromechanical actuator of a dry dual clutch [4] ................................................................ 9 Figure 7: Concentric (left) and parallel design (right) of a multi disk wet dual clutch [1] ...................... 9 Figure 8: Wet dual clutch [2] ................................................................................................................. 10 Figure 9: Single cone synchroniser unit [5] ........................................................................................... 11 Figure 10: Gear shifter unit [1].............................................................................................................. 11 Figure 11: Powertrain structure of a commercial vehicle [1] ............................................................... 12 Figure 12: Driving capability vs. driving task ......................................................................................... 14 Figure 13: Launch, tL and launch hesitation, t LH .................................................................................... 16 Figure 14: Jerk during gear upshift 1 to 2 ............................................................................................. 18 Figure 15: Upshift from gear 1 to 2 ....................................................................................................... 19 Figure 16: Power on downshift from gear 4 to 3 .................................................................................. 20 Figure 17: Judder during clutch harmonisation .................................................................................... 22 Figure 18: Change of mind (let off) ....................................................................................................... 23 Figure 19: Rotary load in AMESim ........................................................................................................ 26 Figure 20: AMESim model of the b-segment car .................................................................................. 28 Figure 21: Getrag Powershift 6DCT250 Transmission [7] ..................................................................... 29 Figure 22: Cross-section view of Getrag 6DCT250 dry dual clutches [7] .............................................. 30 Figure 23: Dry dual clutch model in AMESim........................................................................................ 31 Figure 24: LuK actuator unit for Ford 6DCT250 dual clutch transmission [8] ....................................... 31 Figure 25: Lever concept (left) and clutch actuator unit (right) [4] ...................................................... 32 Figure 26: Clutch actuator (for launch) ................................................................................................. 32 Figure 27: Clutch actuator for launch and gearshift ............................................................................. 33 Figure 28: Complex model of the clutch actuator ................................................................................ 34 Figure 29: Getrag Powershift 6DCT250 transmission layout [7]........................................................... 35 Figure 30: Transmission model in AMESim........................................................................................... 36 Figure 31: Gear actuator of Getrag Powershift 6DCT250 Transmission [7] ......................................... 36 Figure 32: Gearshift diagram for gear 1 to gear 4 ................................................................................ 37 Figure 33: 1.6 Ti-VCT engine [11] .......................................................................................................... 38 Figure 34: Engine torque characteristic curve ...................................................................................... 39 Figure 35: Engine model in AMESim ..................................................................................................... 39 Figure 36: “Direct” control system ....................................................................................................... 40

Figure 37: Control system with AMESim Sequential Function Chart .................................................... 41 Figure 38: Engine speed controller ....................................................................................................... 42 Figure 39: Desired engine speed curve ................................................................................................. 42 Figure 40: Control stages during launch in flowchart view .................................................................. 43 Figure 41: Launch from creep ............................................................................................................... 44


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Figure 42: Launch from brake ............................................................................................................... 45 Figure 43: Control stages during upshift from gear 1 to gear 2 in flowchart view ............................... 46 Figure 44: Engine torque controller ...................................................................................................... 47 Figure 45: Upshift from gear 1 to gear 2 ............................................................................................... 47 Figure 46: Gearshift diagram areas....................................................................................................... 48 Figure 47: Control stages during power on downshift for gear 4 to gear 3 in flowchart view ............. 50 Figure 48: power on downshift for gear 4 to gear 3 ............................................................................. 51 Figure 49: Control stages during power on downshift for gear 2 to gear 1 in flowchart view ............. 52 Figure 50: Power off downshift for gear 2 to gear 1............................................................................. 53 Figure 51: Form window for base App (right) and scripted App for evaluation (left) .......................... 54 Figure 52: Launch from creep with simple clutch actuator model (see chapter 4.4.2) ........................ 56 Figure 53: Launch from creep with the complex clutch actuator model (see chapter 4.4.2) .............. 58 Figure 54: Launch from brake with simple clutch actuator model ....................................................... 60 Figure 55: Launch from brake with complex clutch actuator model .................................................... 61 Figure 56: Launch on hill with simple actuator model.......................................................................... 63 Figure 57: Comparison of upshift of gear 1 to gear 2 between the simple and complex actuator model .................................................................................................................................................... 65 Figure 58: Upshift from gear 1 to gear 2 for accelerator pedal position 40 %, 70 %, 100 % ................ 67 Figure 59: Jerk of upshift from gear 1 to gear 2 for accelerator pedal position 40 %, 70 %, 100 % ..... 68 Figure 60: Upshift from gear 2 to gear 3 for accelerator pedal position 40 %, 70 %, 100 % ................ 70 Figure 61: Jerk of upshift from gear 1 to gear 2 for accelerator pedal position 40 %, 70 %, 100 % ..... 71 Figure 62: Upshift from gear 3 to gear 4 for accelerator pedal position 40 %, 70 %, 100 % ................ 73 Figure 63: Jerk of upshift from gear 3 to gear 4 for accelerator pedal position 40 %, 70 %, 100 % ..... 74 Figure 64: Power on downshift from gear 4 to gear 3 .......................................................................... 77 Figure 65: Jerk of power on downshift from gear 3 to gear 4 for different accelerator pedal position change................................................................................................................................................... 78 Figure 66: Power off downshift for gear 4 to gear 3............................................................................. 80 Figure 67: Jerk during power off downshift from gear 4 to gear 3 and gear 2 to gear 1 ..................... 80 Figure 68: Launch comparison between measured data and simulation ............................................. 82 Figure 69: Upshift gear 1 to gear 2 comparison between measured data and simulation .................. 83 Figure 70: Comparison of acceleration between measured and simulation ........................................ 84 Figure 71: Simulation model basis for launch ....................................................................................... 95 Figure 72: Simulation model basis for upshift ...................................................................................... 96 Figure 73: Simulation model basis for downshift ................................................................................. 97 Figure 74: App interface for launch with simple actuator model ......................................................... 98 Figure 75: App interface for launch with complex actuator model ...................................................... 98 Figure 76: App interface for upshift (shift time) ................................................................................... 99 Figure 77: App interface for upshift (jerk) ............................................................................................ 99 Figure 78: App interface for power on downshift (shift time) ............................................................ 100 Figure 79: App interface for power on downshift (jerk) ..................................................................... 100 Figure 80: App interface for power off downshift (shift time) ........................................................... 101 Figure 81: App interface for power off downshift (jerk)..................................................................... 101 Figure 82: Code snippet for basic plotting app class .......................................................................... 102 Figure 83: Code snippet for basic LED display of calculated values.................................................... 103


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List of Tables Table 1: Subjective evaluation widely used by Automakers [1] ........................................................... 13 Table 2: Summary of the driving situations .......................................................................................... 24 Table 3: Overview of the simulation model complexity ....................................................................... 25 Table 4: Overview of the variable parameters and evaluation criteria for the simulation .................. 26 Table 5: Technical data of the reference b-segment car ...................................................................... 27 Table 6: Technical data of Getrag Powershift 6DCT250 transmission [7] ............................................ 29 Table 7: Gear ratio of Getrag Powershift 6DCT250 Transmission ........................................................ 35 Table 8: Technical data of the reference engine [9] [10] ...................................................................... 38 Table 9: Evaluation criteria for launch from creep with simple actuator model.................................. 57 Table 10: Evaluation criteria for launch from creep with complex actuator model ............................. 59 Table 11: Evaluation criteria for launch from brake with simple actuator model................................ 60 Table 12: Evaluation criteria for launch from brake with simple actuator model................................ 62 Table 13: Evaluation criteria for launch on hill with simple actuator model........................................ 64 Table 14: Jerk of upshift from gear 1 to gear 2 for accelerator pedal position 40 %, 70 %, 100 % ...... 68 Table 15: Jerk of upshift from gear 2 to gear 3 for accelerator pedal position 40 %, 70 %, 100 % ...... 71 Table 16: Evaluation criteria for upshift from gear 3 to gear 4 for accelerator pedal position 40 %, 7 0 %, 100 % ................................................................................................................................................ 74 Table 17: Evaluation criteria for power on downshift for gear 4 to gear 3 .......................................... 78 Table 18: Evaluation criteria for power off downshift for gear 4 to gear 3 and gear 2 to gear 1 ........ 81 Table 19: AMESim Signal and Control library ....................................................................................... 90 Table 20: AMESim Mechanical library .................................................................................................. 91 Table 21: AMESim Powertrain library ................................................................................................... 92 Table 22: AMESim Sequential Functional Chart (SFC) library ............................................................... 93


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List of Abbreviations AMESim

LMS Imagine.Lab AMESim

AMT

Automated Manual Transmission

DCT

Dual Clutch Transmission

CVT

Continuous Variable Transmission

App

Application

ICE

Internal combustion engine

OSS

Output shaft sensor

ISS

Input shaft sensor

Acc pedal

Accelerator pedal

CAN

Controlled Area Network

SFC

Sequential Functional Chart

c1, c2

Clutch 1, clutch 2

Tc1, Tc2

Torque of clutch 1, torque of clutch 2

Fc1, Fc2

Actuation force on clutch 1, actuation force on clutch 2

hofer

hofer-pdc GmbH

VW

Volkswagen AG


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Introduction

1

Introduction

1.1 Motivation In the last few years, the number of passenger vehicles with automated transmission is increasing rapidly. At the same time, the design variation of automated transmission is becoming more and more diverse. Besides the well-known automatic transmission with torque converter and planetary gear wheels, other types of transmission such as the continuous variable transmission (CVT), automated manual transmission (AMT) and the dual clutch transmission (DCT) are becoming more popular in the passenger car market. The listed automated transmission types can be further classified according to their realisation concept, such as the wet and the dry variation type of the dual clutch transmission. As a result of the differences in the concept implementation of the listed transmissions, the driving experience also varies according to the different transmission concept. The driving dynamics, as well as the comfort perception are evaluated by the driver in a subjective way. As a way to improve the development of the transmission, objectification of the driver’s subjective perception is the way

forward. As an example, the power interruption period during acceleration with an automated transmission can be used as an evaluation criterion, since the transmission does not allow a power interruption-free shifting. A lot of evaluation criteria such as the one mentioned above are already put into used in the early stages of simulation-based evaluation. The challenge however lies in the complete evaluation using the objective criteria defined, without relying on the subjective perception of the driver. It is also important to make sure that the parameters as well as the control strategies used in the simulation can be implemented realistically on the real transmission.


1

Introduction

1.2 Scope of Work The aim of this master thesis is to evaluate the subjective perception of comfort in different driving situations by means of simulation. These driving situations are focused on vehicles with automated transmission. To start off, a variety of driving situations was listed and defined. In order to evaluate the listed situations, suitable evaluation criteria were determined to produce the required objective results, i.e. representing and explaining the evaluated driving situations in an accurate matter, in other words, to objectify the subjective perception of comfort during driving. To help the author of the thesis to understand more about the driving situations, several test drives were performed. Subsequently, suitable driving situations were chosen considering the time constraint of the master thesis and the difficulty to realise such driving situations on a simulation program. The chosen driving situations were simulated using the simulation program LMS Imagine.Lab AMESim, or simply AMESim. Depending on the necessity, other programs such as Diadem were used to assist the simulation and evaluation process. The simulation model parameters were calibrated to the reference transmission. The simulation model in the early stage was relatively simple and uncomplicated. Depending on the results of the early stage simulation, improvements were made where deemed necessary by increasing the complexity of the model. Using the measurement methods determined in the first part of the task, objective results were obtained and evaluated. By comparing the simulation results achieved from the hofer benchmark, results from the simulation would then be compared with the measurement data from hofer, further improvements of the simulation were made where deemed necessary. Additionally, in order to facilitate the evaluation process of the obtained results, several programming scripts were written. In the end, conclusions are made according to the comparisons and evaluation of the results.


2

Simulation Software Used

2


To carry out the tasks of the thesis smoothly, it is required that the student were able to use the software products such as AMESim and DIAdem. Since both software programmes were not familiar to the student, an “introduction time” to learn and familiarise with the software was needed. The

software products used to carry out this Master Thesis are described in the following sub chapter.

2.1 LMS Imagine.Lab AMESim LMS Imagine.Lab AMESim or simply AMESim is an element or component based simulation software for the modelling and simulation of one-dimensional systems developed and distributed by LMS International. The software package offers a 1D simulation suite to model and analyse the hydraulic, pneumatic, electrical and mechanical behaviour of the 1D system. In its usage AMESim is similar to Simulink. For modelling of the system, AMESim is equipped with approximately 30 libraries. Due to the partnership of hofer-pdc GmbH with LMS, the complete library package is provided. The important libraries for this thesis are controls, mechanical, pseudo-mechanical, hydraulics, electrical, thermodynamics and powertrain. More about the elements used for the simulation in this thesis can be read under appendix.

Figure 1: LMS Imagine.Lab AMESim

The figure 1 shows a standard interface of AMESim. The modelling and simulation of a system is done in four steps: sketch, submodel, parameter and run. These four steps are represented or highlighted by the 4 panels on the left side of the screen. The four steps are:


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

Sketch mode: Components are selected from the library and are linked together to form a system. Drag-and-drop functionality simplifies and accelerates modelling processes.



Submodel mode: Physical submodel associated to each component is chosen.



Parameter mode: The parameters for each submodel of the system are set and compiled.



Run mode: The simulation is run. The run mode also includes the pre-processing mode. The needed curves which show the behaviour of the system (e.g. displacement vs. time) can be viewed after the simulation ended.

AMESim

also

provides

an

“App

Designer” .

The

App

Designer

is

a

pre

and

post-

processing IDE (Integrated Development Environment) that can be used to create user interfaces (which are also called “App”) for the user’s specific needs, use and reuse them within AMESim. The

App Designer uses the already known QT-Platform with several modifications by AMESim to accommodate its users. Normal users can use the available widgets to assist their work. Advanced python users can additionally design their own widgets by writing their own python scripts (e.g. to enable them to automate the obtaining and evaluation of the simulation results).

Figure 2: App Designer in AMESim

The simulations in this thesis are done using this software program. The App Designer is used automate the repeating steps taken to o btain the simulation results.


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2.2 DIAdem DIAdem is a technical software for managing, analysing, and reporting technical data developed by National Instruments. It is used to analyse data sets that are obtained from test drive equipment, provided by National Instrument. With this software it is also possible to use mathematical functions on a data set or a curve such as the average, integration and differentiation function and in the end graphically present it in a report. This software program is used to read and edit the results obtained from the test drives.


5

Theoretical Foundations

3


3.1 Automotive Transmission and Powertrain A transmission plays a very important role in passenger and commercial vehicles. The main task of a transmission is to convert the traction supplied from the power source, mainly the internal combustion engine, to satisfy the requirements of the driving situations performed by the driver. A transmission consists of sets of gears to provide different gear ratios for the mentioned different driving situations. As an example, during start up or launch of a vehicle from stationary, the driver might want to accelerate from stationary to the desired speed. Hence, a combination of gears which provide a high gear ratio is needed to convert the supplied torque from the engine and accelerate the vehicle. In addition, a transmission also plays an important role in respect to fuel consumption, reliability and safety. The following figure depicts a simple gear set in a schematic view. Gearwheel 1 is connected to shaft 1 and gearwheel 2 to shaft 2 respectively. The letters n stands for rotation per minute, T stands for the torque and z stands for the gear teeth number.

Shaft 1

z1

n1, T1 n2, T2

z2

Shaft 2 Figure 3: Gear ratio

The gear ratio of a gear set can be calculated as follows:

      

(3.1)

As mentioned, an automotive transmission consists of several gear sets as depicted above to provide suitable gear ratios respective to driving situations and fuel consumption. In general, an automotive transmission may be in the form of manual transmission, automatic transmission or automated manual transmission.


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3.1.1

Dual Clutch Transmission

To understand this thesis, it is important to get to know the type of transmission used in the test drive and for the simulation. The evaluation of the comfort criteria would be done using a dual clutch transmission (see chapter 4.4). Dual clutch transmissions (DCT) are categorised as automatic transmission with various gear ratios due to their similarities with respect to control and functionality [1]. A DCT combines the characteristics of a manual transmission, such as a high level of efficiency, a broad range of gear ratios and sportiness, with the ease of handling and shifting without power interruption from an automatic transmission. A DCT generally consists of two sub-gearboxes, each connected to the engine through its own cl utch. One sub-gearbox contains the odd gears (1, 3, 5…) while the other contains the even gears (2, 4, 6…). The following figure shows a schematic design of a DCT.

Figure 4: Schematic of a dual clutch transmission [1]

With the help of the figure above, a basic gear shifting process can be explained as follows. While accelerating in the first gear, the idle second gear is preselected and engaged. Since clutch 2 is not engaged during the idle gear synchronisation process, there is no interruption to the torque supplied by the engine. The driver does not notice the synchronisation process. When the speed for the upshift from first gear to second gear is reached, clutch 1 disengages at the same time when the clutch 2 engages. This phase is known as the cross-fading phase. This enables a power-interruptionfree gear shifting. Once the shifting process ends, the next gear, the third gear, can be preselected, while the first gear is disengaged and the same steps is repeated for upshift. This principle is basically the same for upshift and downshift. The dual clutch built in a DCT can be further divided into two variant types, namely the wet dual clutch and the dry dual clutch transmission.


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Dry Dual Clutches Dry dual clutches are normally used in small vehicle with low engine torque not more than 250 Nm. A clutch of a dry dual clutch transmission usually consists of a single friction plate and the torque is transmitted via pressure plate and friction plate of the clutch, like a normal clutch of a manual transmission. However, there are some design differences compared to the clutch of a manual transmission, such as the dry clutch is normally designed to be in open position (disengaged) when no force is applied to the clutch. It is designed that way to fulfil the safety requirement, which requires the clutch to open automatically when the clutch actuation system fails. Another difference is, because of the high actuation force of dual clutches, direct linkage and bearing support on the crankshaft is not feasible due to the high load. So, the clutch needed to be supported at one of the two shafts of the transmission. There are further two known variants of support design on the shaft. The position of the support bearing is preferred to be on the hollow shaft. What differs here is the position of the damper to eliminate or reduce unwanted oscillation between the engine and the clutches. In variant 1, the torque damper is mounted on the crankshaft, and the crankshaft is linked to the clutch via a drive gear. This drive gear is preloaded in circumference direction and can also compensate axial tolerance between the engine and the transmission shafts. In variant 2, torque dampers are integrated to each friction plate of each clutch. The crankshaft is connected to the transmission shafts via a flywheel with a cardanic function. The “cardan joint” is made of elastic elements which can compensate radial

and axial tolerance between the engine and the transmission shafts [2].

Figure 5: Dry dual clutch with external torsion damper (left) and friction disk integrated damper ( right) [2]


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Figure 6: Electromechanical actuator of a dry dual clutch [4]

The figure above presents the electromechanical actuation concept of a dry clutch. The dry clutch is actuated by an E-Motor by means of engagement lever. When the E-Motor is supplied with power, its shaft-rotation would be converted to translational displacement by the ball screw. The roller on the ball screw here acts as a variable pivot. The apply spring supplies the preload on one lever end. By varying the position of the pivot, different actuation can be achieved on the other end of the lever.

Wet Dual Clutches Wet dual clutches are used in vehicles with high engine output, typically 250 Nm and above. The typical design of wet dual clutches consists of multiple friction disks, to accommodate the high input load. They are mounted directly on the t ransmission shafts or in an external clutch carrier connected directly to the transmission. Most of the wet dual clutches currently in the market are actuated by hydraulic means [2]. There are two typical wet dual clutch designs known implemented by automakers, which are 

concentric design (also called radial arrangement)



parallel design (also called axial arrangement)

Figure 7: Concentric (left) and parallel design (right) of a multi disk wet dual clutch [1]


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Concentric clutches are advantageous for short installation spaces. In a vehicle which uses the concentric design, the outer clutch is preferred as master clutch due to its higher thermal capacity, suitable for low gears which need to transfer high input torque. The cooling oil flows first through the outer clutch to the inner clutch. In the contrary, parallel design are applied in transmission which has limited space in the radial direction. The parallel design enables the first gear to be controlled by either the outer or the inner clutches. Another advantage of such arrangement is that the cooling oil can be supplied separately to each clutch. The following figure presents a dual clutches in concentric design. From the figure, it can be identified that the torque damper are arranged in the dry space between the engine and the dual clutch. Another alternative to this design is to integrate the torque damper to the dual clutch plates in the wet chamber, similar to the dry dual clutch design. To actuate the clutches electrohydraulically, an external hydraulic pack is necessary. The hydraulic pack consists of a hydraulic pump, which pumps the cooling oil and the oil to actuate the clutches, and a valve block for controlling. The actuation oil from the pump would flow through the rotary oil passages to the pressure chambers. Parallel to the pressure chambers are compensation chambers, which are needed to compensate the influence of centrifugal oil pressure that builds up from the rotation.

Figure 8: Wet dual clutch [2]

Synchronizer and Gear Actuation In simpler words, synchronisation of a gear in a vehicle with dual clutch transmission can be defined as firstly, friction coupling with non-planar friction plane, that follows with form locking of an idle gear to a sub-gearbox shaft, to transfer power from the input shaft via the now engaged idle gear and sub-gearbox shaft, to the output shaft. Depending on the application in vehicles (passenger


10


vehicle and commercial vehicle), a synchronizer unit may differ in terms of the number of non -planar friction plane (also known as cone) involved during synchronisation process. In practice there can be up to 3 non planar friction planes in a synchroniser unit (single cone, double cone or triple cone). The number of cones is a multiplication factor for the synchronisation capability of a synchronizer [5]. The figure below depicts a single cone synchronizer.

Figure 9: Single cone synchroniser unit [5]

The same or similar synchroniser unit can also be found on each sub-gearbox shaft in a dual clutch transmission. However, in a dual clutch transmission, the far left and far right idle gears should be st

rd

nd

th

consecutive either odd or even number gears (e.g. 1 gear and 3 gear or 2 gear and 4 gear) so that gear pre selection during upshift and downshift can be achieved. During gear change, the gearshift sleeve would be shifted to the desired shift position. The gearshift sleeve is connected to a gear shifter, which can be seen in the figure 10 below.

Figure 10: Gear shifter unit [1]


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Gear shifter can be actuated by means of hydraulics or electric. The gear shifter presented above is a hydraulic operated gear actuator of a dual clutch transmission. As mentioned under the previous chapter (3.1.1 Dry Dual Clutches), Clutches), a hydraulic operated actuator is preferred for wet dual clutch transmission and hydraulically operated DCT-systems. The same applies to eletromechanically operated DCT-systems. One distinguished feature of a hydraulic gear shifter is the locking element, which is needed to supply the locking force to the synchronisation force from the hydraulic piston.

3.1.2

Powertrain

In general, the torque supplied by the engine in a vehicle must pass through several components before the output at the vehicle tyres. The engine torque i s converted through multiplication of each gear ration from these components. The whole combination of the components is called powertrain. A powertrain mainly consist of 4 sections, which is the engine, the coupling element, the transmission and the final drive, as depicted below.

Figure 11: Powertrain structure of a commercial vehicle [1]

The total ratio i A is the multiplication product of the ratio of each the coupling element, the selector gearbox and the final drive.

      

(3.2)

It is important to understand how the powertrain works, since the output torque at the tyres are influenced by the components in each section, as can be seen above.


12


3.2 Subjective Evaluation of Driving Situation The evaluation of “comfort” by a driver, which consists of noise, visible and sensible oscillation, can be developed through his or her subjective perception. Since every driver’s perception can differ from one another, it is important to evaluate how “comfortable” the driving experien ce through a

group of trained evaluators and a group of customers and taking the average marks from each group. The following table presents one of the ways to summarise the evaluation given by the evaluators that is normally used by the major automakers. The scale used is 1 to 10, with 10 representing the best mark and 1 the worst.

Marks

Flaw detection

Evaluation

10

Not detectable by trained evaluators

Excellent

9

Detectable by trained evaluators

Very good

8

Detectable only by critical customers

Good

7

Detectable by all customers

Satisfying

6

Sensed by some customers as disturbing

Acceptable

5

Sensed by all customers as disturbing

Not acceptable

4

Sensed by all customers as faulty

Faulty

3

Complained and claimed by customers

Fail

2

Only partly functioning

Bad

1

Not functioning

Very bad

Marketable

Not marketable

Table 1: Subjective evaluation widely used by Automakers [1]

The subjective evaluation data are already available as reference for this thesis. However, to increase the understanding of how the subjective evaluation is carried out, a simple subjective evaluation was done by the student as an example. The test was carried out using a VW Passat 2.0 which is also equipped with a dual clutch transmission. The evaluation table used by hofer can be found under appendix.


13


3.3 Objectification of Comfort Criteria To evaluate how ‘comfortable’ a person driving is actually a very difficult task because there are no objective guidelines to it. The feeling of ‘comfortable’ is very subjective depending on different

person. Therefore, this chapter would elaborate on the objectification of subjective feelings that would be used to evaluate the driving situation chosen.

3.3.1

Driving Capability vs. Driving Tasks

Before proceeding with the objectification of the subjective criteria, it is important to determine the target type of drivers. The type of driver is typically differentiated in two categories, the sporty driver and the driver that prefers comfortable driving. One way to differentiate between the two types of drivers is by using the following ‘driving capability vs. driving tasks’ curve.

Figure 12: Driving capability vs. driving task

A driver’s capability is determined by: 

competence: Driving licence, extra training, experience



psychological factors: feelings (under stress, anxiety)



substance: under alcohol or drug influence

and many more. Whereby driving tasks are determined by the following factors: 

increase with increasing driving resistance (air resistance, slope, rolling resistance, acceleration)



secondary factors: pedestrian, road regulations and many more [6] more [6]..


14


From the curve, we can clearly conclude that a comfortable driver is a driver that feels that they have the vehicle under control by having more capabilities than tasks, while a sporty driver is a driver that likes to prefer using their own capabilities to take on the driving tasks. Our target driver is the comfortable driver. This means that the drivers in this category would like to have, to some extent, driving assistance to have a smooth and unburdening driving. An example of such driver is a driver that prefers automatic transmission with smooth gear shifts without vibrations.

3.3.2

Driving Situations and the Respective Comfort Evaluation Criteria

Driving situations that affect the comfort of driving were determined, before an appropriate method of evaluating can be assigned. The following driving situations were identified as having most effect on the driving comfort and needed to be evaluated. 

Launch



Creep



Gear upshift and downshift



Hill hold



Judder



Change of mind

To assist the simulation process regarding the driving situations, it also makes sense to identify the participating sub-systems as well as the measurement instruments (sensors and actuators) used for each driving simulations. This information is to be summarised in a table and suitable driving situations can be chosen based on the information of each driving situations.

Launch In a non-technical term, launch is understood as start-up or moving off of a vehicle from stationary condition to the desired speed. In this thesis, launch is further divided into two sub-definition, that is launch and launch hesitation. The definitions used in launch are as following: 

Launch hesitation, tLH: Period between accelerator pedal actuation and maximum vehicle acceleration



Launch, tL: Period between maximum vehicle acceleration and full clutch engagement



Total time, tT: Sum of launch and launch hesitation

The sub-systems taking part during launch are: 

Accelerator pedal


15




Transmission, which includes first gear and final drive ratios and inertias



Clutch, including clutch actuator and clutch control



Engine, including inertia and engine control strategy



Road profile, for plane and hill launch

Figure 13 can help present the definitions of launch and launch hesitations. The upmost curve shows the engine speed in red and shaft 1 speed in green. The middle curve shows the acceleration of the vehicle while the last curve shows the accelerator pedal actuation signal.

Figure 13: Launch, tL and launch hesitation, tLH

To determine the exact time where each section (launch and launch hesitation) starts, a method of quantification is needed. The right parameters must first be determined, and then the respective signals from the respective sensors can be obtained, either from the Transmission Control Unit or through external built sensors. In this case, the accelerator pedal potentiometer can provide the start time of the launch. The time where maximum acceleration is reached, which signals the end of launch and start of launch hesitation, is calculated through the speed signal of the output speed sensor (OSS). The speed of each sub-gearbox is obtained from its own sensors while the input speed sensor (ISS) provides the


16


engine speed. Therefore, the moment of the speed harmonisation of the engine and shaft 1 can be calculated at the time when both speeds reach a common speed with a constant micro slip.

Creep Creep is normally associated with automated transmission. At start-up, when the driver shifts the gear lever into drive mode (D) but without following up by actuating the accelerator pedal or the brake pedal, the vehicle would accelerate on its own until it reaches a certain ‘creeping speed’ an d

moves forward constantly with this speed. This phenomenon is called creep. Creep in a vehicle with dual clutch transmission is usually achieved by actuation of the clutch with a certain amount of slip (the clutch is not 100% closed). Creep is simulated together with launch. Hence, the sub-systems taking part are almost the same as during launch, except that an extra creep control strategy which control the clutch slip during creeping is needed. The figure 13 also shows creep of the vehicle. If the driver still has not actuated the accelerator pedal after 1 s, the vehicle would start accelerating until it reaches the creep speed. This can be seen through the shaft 1 speed (green curve) of the upmost curve in the figure. The speed difference between the engine speed and the shaft 1 speed can be seen here, which indicates the micro slip in the clutch.

Gear Upshift and Downshift One of the main reasons of the introduction of dual clutch transmission into the automotive market was to improve the smoothness of gear shifting. The key to determining the evaluation parameter is the change felt by the driver. When a driver is driving at a constant speed, he or she would not feel any significant vibration with his or her body. Only when the driver is accelerating would the driver feel the change with his body. Therefore, peak to peak acceleration, a pp, as an example, can be used as a criterion for objectification of the subjective comfort feeling felt by the driver during gear shifting. Another criterion widely used by the automakers to evaluate shifting smoothness is shock intensity or jerk (J). Shock intensity or jerk is defined as rate of change of longitudinal acceleration.

       

(3.3) 3

It is widely accepted by automakers that jerk value of ± 5 m/s as comfortable to drivers.


17


Figure 14: Jerk during gear upshift 1 to 2

The figure above shows the acceleration (top) and jerk (bottom) during upshift from gear 1 to gear 2. The acceleration which falls almost instantaneously during the cross-fading from around 6.3 s to 3

6.62 s (labelled A) causes average jerk (in green) of approximately -6 m/s . As the engine speed is reduced to match the shaft 2 speed, the vehicle is moving with a constant acceleration, hence the constant jerk, labelled with B. As the engine speed reaches the speed of the shaft 2, the matching up of the two speeds causes a slight increase in acceleration hence an average jerk around 5 to 10 m/s

3

(labelled with C). Besides the named criteria above, it is also plausible to take into account the torque phase time, speed phase time and the total shifting time [3]. These parameters can be defined as follows.


18




Torque phase time, t TP: The period between start of torque reduction of the off going clutch and until the oncoming clutch fully engaged (also known as clutch cross-fading). It can also be approximately measured from the acceleration curve, from the fall of acceleration until the minimum acceleration (area labelled as A in figure 14)



Speed phase time, t SP: The period of the deceleration of the engine speed to oncoming shaft speed. It can approximately be measured starting from the minimum acceleration until the acceleration rise again (area labelled as B in figure 14)



Shifting time, tS: Total shifting time which is the sum of torque phase time and speed phase time.

The two phases of gear upshift (torque phase and speed phase) are explained in detail in chapter 4.6.2. The following figure presents the evaluation parameters mentioned above, which are torque phase time, speed phase time and shift time.

Figure 15: Upshift from gear 1 to 2

For power on downshift, the order of the t TP and tSP is reversed. The reason is further discussed in chapter 4.6.3. The following figure depicts the evaluation parameters for power on downshift of gear 4 to 3.


19


Figure 16: Power on downshift from gear 4 to 3

The sub-systems taking part during upshift and downshift are: 

Accelerator pedal



Transmission, which includes first gear and final drive ratios and inertias



Transmission control unit, which is responsible for transmission control strategies



Clutch, including clutch actuator and clutch control



ICE , including inertia and engine control strategy



ICE control unit, which is responsible for ICE control strategies



Road profile

Engine speed, shaft 1 and shaft 2 speeds are provided by their own sensors respectively. The vehicle speed is obtained from the OSS and then differentiated by means of evaluation software programs such as DIAdem or even MS Excel to get the evaluation parameters acceleration and jerk respectively. From the acceleration curve, the torque phase time and speed phase time can be calculated.


20


Hill Assist Control Hill assistant is a mechanism that prevents the vehicle from rolling backwards down a hill when the brake pedal is released by the driver. The aim o f the mechanism is to increase driving comfort during hill launch. The implementation of hill hold mechanism for a vehicle with a dry dual clutch transmission is particularly complicated, since the dry clutch would be closed when the brake pedal is released on hill. If the hill is too steep or the time taken to actuate the accelerator pedal is too long, the clutch would get ‘hot’ and subsequently lose its friction coefficient. To avoid any def ect on

the clutch, most of the hill hold strategy used by the automakers is to disengage the clutch and let the vehicle roll backwards. It is plausible to measure the hill hold time, t H, of a vehicle on different angle of slope. Hill hold time can be defined as period between releasing the brake pedal until the vehicle starts rolling backwards. Generally it is preferable to drivers that a vehicle has a long hill hold time, so that a driver could switch from brake to accelerator pedal without rushing, hence avoiding mistakes such as rollback or engine stalling. An additional element which is important for hill hold is the slope sensor. One of the requirements of the activation of the hill start is that the vehicle needs to be on a slope. However, if the slope angle is more than the critical angle, the hill hold mechanism would not be activated at all to avoid ‘hot clutch’. The moment when the vehicle starts to roll backwards can be determined from the

speed signal provided by the OSS.

Judder Judder can be defined as vertical oscillation of a vehicle. Judder usually happens during vehicle motion at low speed and low accelerator pedal actuation level, which is normally lower than 30%. Vehicle controls nowadays are optimised to avoid judder at every speed; hence it is very difficult to get a vehicle to judder intentionally. Evaluation parameters that can be used to measure judder are the peak to peak values of the vehicle speed oscillation and its frequency. Low amplitude judder can however be easily detected at sub-gearbox-shafts mainly during engagement of the clutch due to the unoptimised controller settings.


21


Figure 17: Judder during clutch harmonisation

Figure 17 shows a low amplitude judder during clutch harmonisation at launch. The shaft oscillation does not cause the vehicle to judder, as can be seen from the speed curve of the vehicle most probably due to the damping in the powertrain.

Change of Mind Change of mind can be divided further into two sub categories, which is tip in and let off. Tip in can be defined as quick sudden pressing of the accelerator pedal during deceleration. This can occur in situation such as when a driver, who is on a branch road, is decelerating to find a gap in between vehicles on the main road, and as soon as a gap was found, the driver would press the accelerator pedal quickly to drive his or her vehicle into the gap. The tip in time, t TI, which is the time between the actuation of the accelerator pedal and the moment when the vehicle starts accelerating, can be used as a criterion to evaluate tip in. In contrary to tip in, let off is defined as the sudden releasing of accelerator pedal when accelerating (pressing of accelerator pedal). This can occur when a driver suddenly sees an obstruction in front of him that needs to be avoided, and quickly releasing the accelerator pedal to actuate the brake pedal. At this moment the vehicle control unit should be able to detect the driver’s request and react as

fast as it can to reduce the vehicle speed. The let off time, t LO, is the time period between the release of the accelerator pedal and the moment when the vehicle starts decelerating.


22


The following figure depicts the change of mind driving situation, namely sudden the let off of the accelerator pedal during driving. As can be seen from the figure, the vehicle only starts to slow down after 2 s letting off of the accelerator pedal.

Figure 18: Change of mind (let off)

3.3.3

Summary of the Driving Situations

The driving situations explained in the previous sub-chapters are summarised in the table below. Driving situation

Launch & launch hesitation

Description

Measurement instruments

Measurement method & quantification

Root causes / corresponding subsystems

Test data

Launch hesitation: defined as the period between acc pedal actuation and the moment when the vehicle reaches peak acceleration

-Acc pedal potentiometer -OSS -ISS -Sensors on sub-gearboxes

Potentiometer (sensor) at acc pedal provides time when acc pedal is p ressed, Output speed sensor provides time of acceleration begin.

Acc pedal -Transmission - Clutch - Clutch Actuation (electromot. or hyd.) -Road profile -Engine - Inertia - Delays from ECU (incl. discreteness) - ICE control strategy (ICE speed by control of torque)

Yes

Launch: period from the moment of peak acceleration until clutch is fully closed Total time: Launch hesitation + launch

'Intersection' of engine speed and shaft speed to detect closed clutch. Sensor at sub-gearbox shaft provides shaft speed, input speed sensor (ISS) provides engine speed. Signals from sensors acquired from TCU/ECU through CAN

test variation: -on plane & on slope -from creep & from brake


23


Driving situation

Creep

Description

Measurement instruments

Measurement method & quantification

Root causes / corresponding subsystems

Test data

Creep: Vehicle moves forward when driver shifts to 1st gear, without pressing acc pedal Variation: -zero pedal, from brake release -on fixed grade

-OSS -ISS

Creep speed and acceleration provided by the OSS

-Engine -Acc pedal -Transmission - Clutch controlled in Slip mode? - Control towards target speed?

Yes

Upshift from gear 1 to gear 6 / downshift from 6 to 1 under normal condition.

-Acc pedal potentiometer -OSS -ISS -Sensors on sub-gearboxes

Evaluation of shifting through 'jerk' and velocity curve of the car

-Acc pedal -Transmission - Clutch - Clutch control (slip) -Road profile (flat road) -Engine - Inertia - ICE sontrol strategy (ICE speed by control of torque)

Yes

Hill hold: Vehicle does not slip down when driver switch from brake pedal to acc pedal on hill

-Slope sensor -Acc pedal sensor -OSS -ISS -Sensors on sub-gearboxes -Brake signal

Slope sensor,OSS provides time when the vehicle begins to slip. Signal from acc pedal must be zero!

-Transmission - Thermal model for clutch -Road profile -(brake/brake assistance)

Yes

Judder: Vibration during idle and low vehicle velocity

-Acc pedal sensor -OSS -ISS -Sensors on sub-gearboxes (-Slope sensor)

Evaluation of shifting through 'jerk' and velocity curve of the car

- Transmission - Clutch with "capability for judder"

Yes

Tip in: Quick sudden pressing of the acc pedal during deceleration. Instead of shifting down, stay at the same gear and anticipate next move of the driver.

-Accelerator pedal potentiometer -OSS -ISS -Sensors on sub-gearboxes

Period between tipping the accelerator pedal until the car starts accelerating is calculated. Measurement same as 'launch hesitation'.

-Acc pedal -Transmission - Shift Strategy -(Engine)

Partially available

Gearshift (up-/ downshift)

Hill hold & hill assist control

Measure maximum holding time at certain slope Variation of strategy: -closing of clutch - brake assistance

Judder

Change of mind -tip in -let off

Let off: Qquick sudden releasing of the acc pedal during acceleration. Instead of shifting up, stay at the same gear and anticipate next move of the driver.

Indentify brake assistance strategy through brake signal, clutch closing through shaft speed sensor and ISS

Period between let off of the accelerator pedal until the car starts decelerating is calculated. Measurement same as 'launch hesitation'.

Table 2: Summary of the driving situations

After careful consideration regarding the time frame of the thesis and the workload, the driving situations launch, creep, upshift and downshift were chosen to be simulated.


24

Simulation Model Setup and Parameterisation

4


This chapter deals with the modelling of the simulation sub-systems by using the chosen reference vehicle. Besides the model setup, variable parameters used and model complexity variations are also presented in the following chapters. This chapter would be the prerequisite to understanding chapter 5, which presents the simulation results.

4.1 Overview of the Complexity of the Simulation Model The following table depicts the overview for the simulation done for this thesis. There are a total of 5 “submodels” that needed to be modelled, which are the engine, clutch actuator, car, gears and

synchronizers, and the control system. Launch

Upshift

from Brake

From Creep

Launch on slope

1-4

Basic

x

x

x

Extended

x

x

x

x

x

x

x

x

Complex

x

x

x

Gears & Synchronizer

Basic

x

x

Control

Direct

x

x

Flow diagram

x

x

Engine

Downshift Power off

Power on

(gear 2-1)

(gear 4-3)

x

x

x

x

x

x

x

x

x

x

x

x

x

x

x

x

x

Complex Clutch-Actuator

Basic Extended Complex

Car

x (gear 1-2)

Basic

Table 3: Overview of the simulation model complexity


25


The variable parameters and the evaluation criteria for each driving situations are listed in table 4. Launch

Variable parameters

Evaluation criteria

from Brake

From Creep

x

x

Acc pedal position (1,0 - 0,3) Slope

Upshift Launch on slope

1-4

x

x

Downshift Power off

Power on

(gear 2-1)

(gear 4-3) x

x

Friction coefficient

x

x

x

Launch

x

x

x

Launch hesitation

x

x

x

Cross-fading time

x

x

x

Total shift time

x

x

x

Jerk

x

x

x

Max. creep speed

x

x

x

Max. slope

x

x

x

Table 4: Overview of the variable parameters and evaluation criteria for the simulation

4.2 Simulation Components in AMESim Simulation components in AMESim are called submodel and usually consist of one or more ports. It is important to understand that the number of ports does not necessarily represent the number of input or output. A port can consist of multiple numbers of input or output and at the same time the combination of both input and output.

Figure 19: Rotary load in AMESim

Figure 19 shows as an example a rotary load with two shafts and friction in AMESim. The two shafts represent the ports of the submodel. As can be seen, the both shafts co nsist of inputs and outputs at the same time. The left shaft has an input torque and provides an output angular speed, whereby the right shaft has an input torque, which can occur because of reaction force from other connected component, and an output angular acceleration.


26


4.3 Reference Car The car that is used as a reference for the simulation is one of a b-segment class car. Some of the examples in this class include VW Polo, Ford Fiesta and Peugeot 206. Further information about the engine, transmission and other submodels are presented together with their simulation models in the following sub-chapters. The general information of the car is summarised in the following table. The information is provided by hofer internal. Curb weight Maximum speed Acceleration 1-100 km/h Drag coefficient cw Front cross-section area AF Fuel type Fuel consumption urban/outside urban/combined CO2 Emission Tyres Power density

1110 kg 190 km/h 10.0 s 0.3 2 2.2 m Gasoline 7.8/4.5/5.8 (litre/100 km) 133 g/km 195/50 R 15 H 0.09 kW/kg

Table 5: Technical data of the reference b-segment car

The following figure presents the vehicle model in Amesim. The vehicle submodel is a 2D submodel with 3 degrees of freedoms due to its longitudinal, vertical and pitch translation. Basic geometrical parameters of the vehicle such as its mass, centre of gravity position, pitch inertia, wheelbase and track dimensions, cross-section area and drag coefficient were set in this submodel. The inputs into the vehicle part are the headwind speed, longitudinal forces partly due to the road profile and drive torque from the powertrain via the tyres and shock absorbers. The vehicle submodel provides the output of distance, speed and acceleration of the vehicle in all three degrees of freedoms. Spring and damping coefficient of the car suspensions were set in the respective model, while parameters such as tyre types and tyre dynamic rolling radius were set in the wheel and tyre submodel. The wheel and tyre submodel has the outputs of drive torque on the road and longitudinal and vertical forces on the vehicle carbody. Its inputs are torque from the powertrain, input torque resulting from the friction with the road and the brake signal (can be set as brake torque, brake force or in %). The road profile also enables the user to set the inclination and the condition of the road, such as dry or slippery road.


27


Figure 20: AMESim model of the b-segment car

The inertia of the tyres and wheels as well as the car suspension stiffness and damping were set as following. 2



Wheel and tyres:

J = 1 kgm ;

rdyn = 0.270 m



Car suspension:

c = 20000 N/m;

d = 2000 Ns/m


28


4.4 Reference Transmission: Getrag Powershift 6DCT250

Figure 21: Getrag Powershift 6DCT250 Transmission [7]

The reference transmission used to investigate the comfort criteria of a DCT is the Getrag Powershift 6DCT250 Transmission. It is developed together by Getrag and Ford, and is mainly targeted for the Band C-segment vehicles. The main target of its development is to achieve better fuel consumption than any other automatic transmission. It is claimed that this transmission shows 10-20% improvement of fuel consumption compared to the state of art planetary automatic transmission with torque converter. Besides better fuel consumption, by using electromechanical concept for the clutches and gear actuations, a reduction in CO 2 emission is achieved. The Getrag 6DCT250 is a three shaft design transmission developed for B and C-segment vehicles of Ford and Renault vehicles with front-transverse engine position. Since the transmission uses the dry dual clutch system, the torque capacity of the transmission is limited to maximum 250 Nm. No additional cooling system for the dry dual clutches is required. The important information of this transmission is summarised in the table below. Weight Length Clutches Clutch actuation Clutch torque capacity Gears Drive mode Oil Volume

75kg (without EM) 350-380 mm Dry single plate dual clutches Electromechanical 250 Nm 6 forward gears, 1 reverse gear Automatic, manual (sequential) 1.7-1.9 litre

Table 6: Technical data of Getrag Powershift 6DCT250 transmission [7]

4.4.1

Dry Dual Clutches

The dual clutches in the Getrag transmission were designed according to variant 2 of dry dual clutch (see chapter 3.1.1 Dry Dual Clutches), which uses friction plate integrated with torque damper for each clutch. This design is favourable for applications with low engine excitation such as for petrol


29


application. The clutches are axially supported by a support bearing on the hollow shaft and radially supported on the crankshaft. Friction disks Cover Pressure plate 2 Pressure plate 1 Flywheel Flexplate Support bearing

Torque dampers

Figure 22: Cross-section view of Getrag 6DCT250 dry dual clutches [7]

Figure 23 describes the dual clutch model in AMESim. The clutch is modelled as two rotating bodies with a common rotation axis. It uses the coulomb friction model which is represented as follows. The input of the clutch is set to normal force, F normal .

    

(4.1)

The coulomb friction model is extended with the tanh function that helps eliminate the difﬁculty in determining the friction force at zero sliding speed both at start up and at direction change. This model is more numerically stable than the coulomb-viscous friction model [12]. The friction force developed at the contact can be described as following.

   [  ]

(4.2)

Whereby Vrel is the relative speed of the two rotating bodies and dV is the rotating speed threshold. Fdyn is the coulomb friction force and can be calculated from the input normal force F normal and coefficient of friction µdyn [13]. An inertia-element is connected to each clutch, and this inertia represents the reduced inertia of the clutch and the involving gears on each subgearbox-shaft. Appropriate viscous friction value can also be set if needed. The parameters of the inertia are set as below.


30


2



Reduced inertia, J = (0.005 … 0.01) kgm



Viscous friction coefficient, d = (0.0 … 0.00 1) Nm/(rev/min)

Figure 23: Dry dual clutch model in AMESim

4.4.2

Electromechanical Actuator of the Dual Clutches

Figure 24: LuK actuator unit for Ford 6DCT250 dual clutch transmission [8]

Figure 24 shows the arrangement of the clutch actuator motors on the dual clutch unit. The actuator motor, also known as electronically commutated motor (or simply EC-Motor), has a range of power from 110-170 Watt. However, the limit of continuous loading is approximately 20 Watt of electrical Ahmad Hakim Mohd Sorihan

31


energy input to avoid thermal overload of the EC-Motor. The dual clutch actuator, together with the dual clutch unit, is developed by LuK and uses a simple lever actuator concept. The mechanism of the lever actuator can be explained using the following figure of a simple lever. A preloaded spring will provide the spring force F spring on one end of the lever. By varying the position of the pivot by means of the EC-Motor via a ball screw (see also chapter 3.1.1 Dry Dual Clutches), variable clutch actuation force on the other end of the lever can be achieved.

Figure 25: Lever concept (left) and clutch actuator unit (right) [4]

Thus, by referring figure 25 (left) the clutch actuator force, F clutch , can be calculated as following.

      

(4.3)

Figure 26: Clutch actuator (for launch)

At the early phase of the simulation, a simple PID controller was used to model the slip controlled clutch actuation (figure 26). The actual slip is calculated from the difference between feedback


32


engine speed and shaft speed. The desired slip input is given. The desired slip is located at the negative port to make sure that the input value into the PID controller is positive to further avoid negative output force from the controller. The limiter also serves the same purpose. A PT1-Filter is connected at the output to smoothen the output clutch force. To simulate gear shifting during driving, the clutch actuator was enhanced to include torque controller during clutch cross-fading phase. The torque controller also uses a PID-controller with the input of error between actual and desired torque and the output of clutch actuator force.

Figure 27: Clutch actuator for launch and gearshift

A more complex model of the clutch actuator was later built to take the power limit of the actuator motor and the friction force caused by the normal force acting on the lever pivot into account. The actuator is still slip-controlled; hence using the same input as the previous actuator model. A torque limiter was placed for the actuator motor. The upper and lower torque limit, T U/L is calculated as follows, whereby the maximum power of the actuator motor is set as 110W and the actuator shaft rotation speed, nEM, can be read from the rotational speed sensor element connected to i ts shaft.

 |   | 

(4.3)

The friction torque acting on the pivot is a function of the output clutch force is calculated as follows.

         Ahmad Hakim Mohd Sorihan

(4.4)

33


These defined equations were set as a function in the new actuator model. The implementation of a variable PID-controller enables the user to use only one PID-controller for both slip control as well as cross-fading control. Only the input to the controller needed to be switched using the signal switch element (input error from torque during cross-fading, input error from slip during other condition). Endstop elements were added to set the maximum distance of the actuator. The following figure presents the more complex clutch actuator modelled for the simulation.

Figure 28: Complex model of the clutch actuator

4.4.3

Gears and Gear Actuators

The Getrag 6DCT250 transmission is a 3 shaft design type transmission, which enables a compact design for small and medium size vehicles. It has 6 forward gears and 1 reverse gear. The low gears st

nd

(1 and 2 gears), which requires high torque capacity transfer, are synchronised with double cone synchronisers whereas the rest of the gears are synchronised with single cone synchronisers. The intermediate gear required to change the direction of the vehicle in reverse gear is integrated nd

together with the idle 2 gear, thus saving space of an extra shaft for the reverse gear. st

The input shaft 1, which is actuated by clutch 1, is responsible for the actuation of the odd gears (1 , rd

th

nd

th

th

3 and 5 gear) while the input shaft 2 for the even gears (2 , 4 and 6 gear). The idle gears of the st

nd

th

th

rd

th

1 , 2 , 5 and 6 gear are located on the output shaft 1 while the idle gears of the 3 , 4 and reverse gear are on the output shaft 2. Both output shafts are connected to the differential. Thus, while driving in a certain gear, the power flows from the ICE through the clutch and input shaft, depending either on odd or even gear, and then through the respective output shaft to the


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differential and the tyres. An exception for the reverse gear, the power flows firstly through the nd

intermediate gear which is integrated to the idle 2 gear on output shaft 1, to the idle reverse gear on output shaft 2 and lastly to the differential and tyres. Thus, a change of direction can be achieved. Figure 28 presents the layout of the Getrag 6DCT250 transmission.

Figure 29: Getrag Powershift 6DCT250 transmission layout [7]

The following table summarises the location of each idle gear, the gear ratios for each gear and the total ratio of each gear after multiplication with the final drive ratio.

Input Input shaft 1 (Clutch 1) Input shaft 2 (Clutch 2)

Gears

Output

Gear ratio

Final drive ratio

Total ratio

1 3 5 2 4 6 R

Output shaft 1 Output shaft 2 Output shaft 1 Output shaft 1 Output shaft 2 Output shaft 1 Output shaft 2

3.92 1.44 0.87 2.43 1.02 0.70 3.51

3.89 4.35 3.89 3.89 4.35 3.89 4.35

15.2488 6.264 3.3843 9.4527 4.437 2.723 15.2685

Table 7: Gear ratio of Getrag Powershift 6DCT250 Transmission

The transmission gear sets were modelled using the 3 ports gear submodel and the 4 ports idle gear submodel. The idle gear submodel must be used with together with the half synchroniser submodel for it to fully function. By using the gear submodels provided by AMESim, the user can set the geometry of the gears, as an example the working transverse pressure angle, αtw and helix angle, β. However, the parameters used for the simulation in this thesis were only the working radius and the constant gear efficiency. As mentioned before, the inertia for the participating gear is reduced to a single inertia for each shaft (see chapter 4.4.1). The differential is not modelled for the simulation Ahmad Hakim Mohd Sorihan

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since the driving situations simulated were assumed to have been done on a straight road; hence no influence from the differential. The differential ratio is already included in the final ratio gear. The powertrain model is built as the following in AMESim.

Figure 30: Transmission model in AMESim

The Getrag 6DCT250 transmission uses an electromechanical gear actuator, which comprises two actuator motors, one for each output shaft. A shift drum, which is designed with groove around it, is linked to the actuator motor via two intermediate gears. Each shift drum with groove is responsible for two shift forks on its respective output shaft. During actuation, because of the groove design, the rotating shift drum would slide the shift fork axially and the shift fork, which is attached to the synchronizer, will engage the desired idle gear.

Actuator motors

Shift drums

Shift forks

Intermediate gears

Figure 31: Gear actuator of Getrag Powershift 6DCT250 Transmission [7]


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Due to the pre-select gear mechanism of the dual clutch transmission, gear synchronisation does not have a big influence on the comfort during gear shifting. So, the gear pre-engage and disengaging mechanism was simplified with a logic function of binary signal (0/1) as an input for the AMESim synchroniser model. Attention should be paid to the gear pre-select logic, so that there is no double overlapping of pre-select, to avoid unnecessary energy loss due to the extra inertia. The following figure shows the gear shift diagram for upshift and downshift of the transmission from gear 1 to 4 and vice versa. The input parameters are accelerator pedal position and the actual vehicle speed. The gear shift diagram is designed so that gear upshift occurs at low rpm for low accelerator pedal actuation for a fuel efficient drive. At high accelerator pedal position, which signals a need for high load, the gear upshift occurs at high rpm. The normal lines represent upshift curves while the dotted lines represent the downshift curves.

Figure 32: Gearshift diagram for gear 1 to gear 4


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4.5 Reference Engine: 1.6 Ti-VCT

Figure 33: 1.6 Ti-VCT engine [11]

The ICE used as a reference is the 1.6 TI-VCT Engine. It is a small inline four cylinder engine with 1.6 litre volume. The “Ti -VCT” stands for Twin independent Variable Camshaft Timing . The current

model is from the year 2010 and complies with the Euro 5 emission level. The important information of the engine is summarised in the table below. Design type Cylinder capacity Injection system

Inline four cylinder engine 3 1596 cm Multi point fuel injection with fuel pump and piezo injectors 88 kW @ 6000 rpm 152 Nm @ 4050 rpm 11.0:1 Euro 5

Maximum power Maximum Torque Compression Emission level

Table 8: Technical data of the reference engine [9] [10]

With the information of maximum power and torque provided, a curve of torque vs. rpm at maximum throttle/accelerator pedal position was generated. Automakers normally do not disclose the characteristic curve of their engine torque. Thus, further curves at different accelerator pedal position were estimated using the maximum value curve as reference. These curves altogether represent the engine operating map, which is needed for the modelling of the ICE in the simulation model. The following figure presents the torque vs. rpm curves at different accelerator pedal position for the 1.6 Ti-VCT engine. One non negligible part of the engine operating map is the negative torque at 0% accelerator pedal actuation, which is caused by the drag torque from the engine and rotating components of the powertrain.


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Figure 34: Engine torque characteristic curve

The following figure presents the vehicle engine modelled in AMESim. The engine is modelled as the characteristic curve of figure 34 with the accelerator pedal position as the input, the engine speed and as the feedback input and the engine torque as output. The engine is equipped with a PIController as the engine limit speed controller. The controller has the engine speed input and equipped with a limiter at the input. The lower limit is set at idle speed (800 rev/min) while the upper limit at 6500 rev/min. If the engine speed is outside the allowed range, the error builds up at the summation point and hence the PI-controller is activated. The controller output will act on the accelerator pedal signal (e.g. throttle signal) by increasing the input if the engine speed is below the idle speed or by decreasing the input if the engine speed exceeds the maximum limit speed.

Figure 35: Engine model in AMESim


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4.6 Control System There are two possible methods of modelling the control systems in by AMESim. The first one is modelling with the provided signal blocks. This type of control system is advantageous for a simple and small control loop, since it provides a clear overview of the control loop. Simulation of a driving situation normally requires more than one stage transition. To realise these transitions, the signal switch element needed to be used. The signal switch element however does not allow a rapid change (e.g. step type of signal) due to its “discontinuity handling”. Thus, to implement a control system of a driving situation consisting of a number of stages, control loops with a lot of signal switch elements are not preferred.

Figure 36: “Direct” control system

However if one has more than one control loop with a lot of input and output parameters, control system modelling with the “Sequential Function Chart” (SFC) is preferred. The SFC is a graphical programming language based on GRAFCET. The SFC enables the user to model control system that is split into stages (similar to the flow chart method) and is used with a BUS line. The control model consists of 3 submodel types, the “SFC step and action” , the “SFC transition and condition” and the “control supercomponent”. The “SFC step and action” submodel controls the activation of the


40


supercomponents. One “SFC step and action” submodel represents one stage in a f low chart, but not limited to the number of control supercomponent assigned. When a certain predetermined condition is fulfilled, a transition to the next stage occurs. This condition can be defined in the “SFC transition and condition”. Using the “ jump branch” in combination with the “SFC transition and condition”, a decision junction such as one in a flow chart can be modelled. Control system modelling with the SFC has a big advantage in terms of parameter settings, since the parameters of the submodels from the signal and control library can be re-defined in the control submodels of each stage. Other advantages include a more orderly modelling and can avoid discontinuities that occur from transition modelling with the signal switch.

Figure 37: Control sys tem with AMESim Sequential Function Chart

Figure 36 and 37 show the complete AMESim model for the driving situation launch. Initially the driving situation was modelled “directly”, as presented in figure 36. As a preparation for model

expansion to accommodate other driving situations, modelling wi th the “sequential function chart” was chosen instead of the “direct” control system. The engine speed controller in figure 36, which cannot be seen in figure 37, is actually built in the control supercomponent of stage 2 and is only activated when stage 2 is activated. The submodels which build the clutch actuator are grouped as one supercomponent.


41


4.6.1

Launch / Moving Off

The clutch is predominantly responsible for the vehicle launch in a vehicle with automated transmission. Hence, a clutch control with the right launch strategy needed to be modelled for the vehicle model in AMESim. Before proceeding to the control stages of launch, it is important to know that the simulation model for launch has an extra PID controller to control the engine speed during engagement of the clutch. The engine speed controller maintains a constant engine shaft speed as it “waits” for the clutch to fully engage.

Figure 38: Engine speed controller

The desired engine speed is a function of the accelerator pedal position. The increase in desired engine speed is proportional to the increase of accelerator pedal position. The desired engine speed curve is displayed below.

Figure 39: Desired engine speed curve


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The following figure describes the control stages that take place during vehicle launch. This includes the variation of launch from creep and launch from brake. Initial stage Vehicle stationary Engine speed is on idle speed  

Transition condition: Accelerator pedal not pressed and 1 s delay -If accelerator pedal pressed, bypass stage 1 (launch from brake) Stage 1 First gear selected Clutch engaged with slip to bring vehicle to creep Stage bypassed if accelerator pedal engaged  



Transition condition: Accelerator pedal pressed Stage 2 

Full clutch engagement begins



Engine speed controller limits the engine speed

Transition condition: Target engine speed reached Equality of engine and shaft speed

Stage 3 

Clutch fully engaged

Figure 40: Control stages during launch in flowchart view

The following figure describes the stages of launch from creep observed from the engine and shaft speed.


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Figure 41: Launch from creep

As presented in figure 41, the simulation begins with the engine speed at constant idle speed of 800 rev/min and the shaft 1 speed of 0 rev/min. At 1 s, clutch torque starts to increase indicating the beginning of stage 1. The vehicle accelerates until it reaches the creep speed and continues moving with constant creep speed. At 4 s, the driver actuates the accelerator pedal and stage 2 begins. Clutch torque starts to increase. The engine speed also increases until around 2000 rev/min and stays constant due to the engine speed controller. The engine speed controller is set with a relatively high value for the proportional and integral gain to make sure that the clutch actuation has only little influence on the engine shaft speed and keep the engine speed constant at this stage. The desired value of the target engine speed is set according to the accelerator pedal position, i.e. the higher the accelerator pedal position, the higher the desired value of the target engine speed (see also Figure 39). At 5.2 s, speed equality of engine and shaft 1 is reached and the clutch is fully engaged.


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Figure 42: Launch from brake

During launch from brake, the accelerator pedal is actuated before creep starts, causing the vehicle to move off before it enters the creep phase. From the flowchart (see Figure 40), the control skips stage 1 and goes directly to stage 2. This is also displayed in figure 42, when the accelerator pedal is actuated at 0.9s, and a transition from the initial stage to stage 2 occurs. The rest of the stages are similar to launch from creep.

4.6.2

Upshift

The clutch-to-clutch shifting process, which normally occurs in a DCT, mainly consists of two parts, the torque phase and the speed phase. During the torque phase of upshift, torque is transferred from the off going clutch to the oncoming clutch. This process is also called clutch cross-fading. At the speed phase, the engine speed is synchronised to the shaft speed of the oncoming clutch [3].


45


The upshift (here gear 1 to gear 2) is also modelled with the SFC in AMESim and consists of 5 stages. Initial stage 

Vehicle accelerating constantly in gear 1 due to constant accelerator pedal position

Transition condition: Upshift curve crossed -Upshift curve defined in gearshift diagram (see Figure 32) Stage 1 

Reduction of clutch 1 torque



Control of clutch 2 torque increase to maintain constant micro slip between engine speed and shaft 1 speed



Also known as torque phase /cross-fading phase

Transition condition: Clutch 1 does not transfer torque anymore

Stage 2 

Reduction of engine torque to decelerate engine speed to shaft 2 speed



Complete disengagement of clutch 1



Also known as speed phase

Transition condition: Equality of engine speed and shaft 2 speed Stage 3 

Engine torque brought back to normal



Clutch 2 fully engaged

Transition condition: Time delay of 0.05 s to 0.1 s

Stage 4 

Disengagement of gear 1



Preselect gear 3

Figure 43: Control stages during upshift from gear 1 to g ear 2 in flowchart view


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During stage3, the reduction of the engine torque (also known as torque down) is controlled by a PID-controller. The desired engine torque is a function of accelerator pedal position. The input of the engine torque controller is the error between desired engine torque and actual engine torque and the output is the accelerator pedal position.

Figure 44: Engine torque controller

The following figure describes the stages of upshift from gear 1 to gear 2 based on the engine speed and shaft speed curves as well as the torque curves.

Figure 45: Upshift from gear 1 to gear 2


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At the initial stage, the vehicle is accelerating constantly due to the constant accelerator pedal position. Shaft 2 is also rotating due to the pre-engaged gear 2. Stage 1 starts at 2.75 s, as the torque of the off going clutch (clutch 1) starts reducing its torque capacity, due to the upshift curve of gear 1 to gear 2 crossed (see Figure 32). To maintain a micro slip between the engine and shaft 1, the oncoming clutch (clutch 2) engagement begins. At 3.05 s, clutch 1 is fully disengaged, signalling the transition to stage 2. The engine torque is reduced to decelerate the engine speed to the speed of shaft 2. Stage 3 begins as soon as speed equality of the engine and shaft 2 is reached at 3.48 s. The engine speed is brought back to normal and clutch 2 is fully engaged. At 3.54 s (stage 4), gear 1 is disengaged and gear 3 is pre-engaged. The disengagement and pre-engagement of gears causes a little disruption to the torque progression, as can be seen at gear 3.54 s.

4.6.3

Downshift

Downshift of a gear during driving is further differentiated to two types, precisely the power on downshift and the power on upshift. Similar to the upshift process of a DCT, downshift also consists of the known two parts, the torque phase and the speed phase. However, the order of the two phases is reversed for power on downshift but is the same for power off downshift. The control strategy of a downshift process is also completely different compared to upshift [3]. The following figure can be used to differentiate between power on downshift and power up downshift.

Figure 46: Gearshift diagram areas


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Power on downshift occurs when more torque is needed compared to the maximum torque offered at the actual gear. In this case, the driver may have further pressed the accelerator pedal in a short period, signifying the need of more torque for acceleration. Power on downshift can only occur at the positive torque region. In the case of this thesis’s simulation model, the positive torque is above

the 15 % of accelerator pedal actuation, as highlighted in figure 46. Using figure 46 as a reference, the driver presses the accelerator pedal from 50 % further to 100 % at vehicle speed of 55 km/h. The downshift-line is crossed at around 60 % of accelerator pedal actuation and hence a downshift from gear 4 to gear 3 occurs. In special cases a crossing of more than one downshift-line can occur, causing a multiple gear downshift. Power off downshift can only take place during deceleration of a vehicle. A vehicle deceleration can occur because of braking or simply due to a very little or no actuation of actuator pedal, which causes a negative torque on the clutch. Referring to figure 4 6, at 0 % accelerator pedal actuation, the vehicle is decelerating. As the vehicle decelerates, it crosses the downshift-line at around 10 km/h and a power off downshift from gear 2 to gear 1 occurs. As the engine shaft reaches the idle speed, the clutch is disengaged to avoid the engine from stalling.


49


Power On Downshift The power on downshift (here gear 4 to gear 3) was modelled in AMESim and consists of 4 stages. Power on downshift differs from upshift in terms of the order of torque and speed phase, which is reversed. A control strategy during speed phase of increasing the engine torque to accelerate the engine speed to the speed of oncoming shaft is implemented by automakers. However it works only if there is available reserve torque (accelerator pedal is not fully pressed by the driver during acceleration demand). The engine torque increase strategy is not implemented in this simulation. The following flowchart presents the 4 stages during power on downshift. Initial stage 

Vehicle acceleration increase due to accelerator pedal actuation from 50 % to 100 %



Next low gear (gear 3 is preselected)

Transition condition: Downshift curve crossed due to increase in accelerator pedal position Stage 1 

Clutch 2 torque reduction to half of its original value to accelerate engine speed to oncoming shaft speed (helped by engine torque increase due to demand from driver)




Transition condition: Speed equality of engine and shaft 1 Stage 2  

Clutch 2 torque further reduced to 0. Clutch 1 torque increased through control for micro slip between engine and oncoming shaft

Transition condition: Clutch 2 does not transfer torque anymore

Stage 3 

Clutch 1 fully engaged

Figure 47: Control stages during power on downshift for gear 4 to gear 3 in flowchart view


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Figure 48: power on downshift for gear 4 to gear 3

Figure 48 can assist the description of power on downshift stages of gear 4 to gear 3. In the initial stage, the driver is actuating the accelerator pedal at a constant position. At this stage, it is still possible to accelerate to a certain degree, as can be seen from the engine torque curve. However, at around 2.68 s, the desired torque exceeds the torque capacity supplied by the current gear, thus a power on downshift to the next low gear begins. At stage 1, which is the speed phase, the torque capacity of clutch 2 is reduced to half of its initial amount to assist the acceleration of engine speed to shaft 1 speed. As soon as speed equality of the engine and shaft 1 is reached, transition to stage 2, also called the torque phase, begins. The torque capacity of clutch 2 is further reduced to zero and at the same time, clutch 1 is engaged to bring micro slip between the engine and shaft 1. Finally at stage 3, clutch 1 is fully engaged as soon as clutch 2 disengages.


51


Power Off Downshift The order of the torque phase and speed phase during power off downshift is the same as that of upshift. For the simulation of this driving situation in AMESim, clutch actuation force is controlled during clutch cross-fading instead of the of clutch torque, since the negative torque on the clutch due to the drag torque is difficult to control. Negative value of desired slip is used for slip control due to the transmission shafts rotating “faster” than the engine shaft. The clutch control during power

off downshift consists of 4 stages. Initial stage 

Vehicle decelerating in gear 2 at 0 % accelerator pedal actuation



Gear 1 preselected

Transition condition: Downshift curve crossed due to vehicle deceleration Stage 1 

Clutch 2 actuation force reduced to 0



Clutch 1 slip controlled to maintain the current condition (moving with clutch 2’s speed)



Also known as torque phase

Transition condition: Clutch 2 disengages, does not transfer any more torque Stage 2 

Clutch 1 force further increased as clutch 2 completely disengages




Transition condition: Speed equality of engine and shaft 1 Stage 3 

Clutch 1 fully engages and slip controlled

Figure 49: Control stages during power on downshift for gear 2 to gear 1 in flowchart view


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Figure 50 presents the power off downshift from gear 2 to gear 1 simulated in AMESim.

Figure 50: Power off downshift for gear 2 to gear 1

In the initial stage, the vehicle is decelerating due to the low accelerator pedal actuation and at 1.1 s a complete release of the accelerator pedal. At around 1.58 s, stage 1 begins. The actuation force of clutch 2 is reduced constantly while clutch 1 is actuated through control to make sure that there the engine shaft is still rotating with the same speed as the shaft 2 speed. As clutch 2 completely disengages, which signals the start of stage 2, clutch 1 actuation force is further increased to accelerate the engine shaft to the speed of shaft 1. As the engine speed reaches approximately the same value as the shaft 1 speed, the clutch 1 is slip-controlled again, and clutch 1 adapted its actuation force to keep a constant slip between engine shaft and shaft 1.


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Evaluation of Results

5


A variety of results were obtained from the simulation due to the variable parameters and simulation model variation. These results are the pre-defined evaluation criteria as defined under chapter 3.3.2 (see also Table 4).

5.1 Script/Apps for Evaluation of Results Python scripts were written for each simulation model to design “ Apps” for each model (see also

chapter 2.1). With one click of the App, the needed simulation results can be extracted. In addition to the simulation results, the respective areas where the comfort criteria are evaluated can also be displayed.

Figure 51: Form window for base App (right) and scripted App for evaluation (left)

Figure 51 displays the App designed for the evaluation. On the right is the Form window for the base App. The form window must contain a “Qwidget”, which prepares an area for the plot. A script was written to display the curves used for evaluation in the Qwidget. From the curves, the results of the evaluation criteria were calculated and displayed in the ready-made LCD widgets. The calculations however were also done in the script written for the App. The script is written in the programming language Python and its extensions (scipy, numpy). The python scripts for evaluation of each driving situation can be found in th e appendix.


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5.2 Evaluation of Simulation Results To investigate the effect of variable value of a parameter on the result, a simulation must be run for each value of the parameter. AMESim however have a function called “batch run”, which enables the user to run simulations with variable parameter values sequentially.

The variable “batch

parameters” are set in parameter mode, and in simulation mode the user can select “batch run”.

This enables the presentation of different curves of a simulation due to the variable parameter values, and is really useful for comparison of results for variable values of a parameter. This function is used to vary the accelerator pedal between 30 % to 100 % actuation, the road slope value and the friction coefficient value of the clutch plates.

5.2.1

Launch/Moving Off

In this thesis, the situation of launch is further divided to three variations, launch from brake, launch from creep and launch on hill. Launch from brake and launch from creep were done on plane. For launch on hill, it was simulated from creep. This way, the maximum creep speed on a slope as well as launch hesitation and launch time can be measured.

Launch from Creep During launch from creep, the accelerator pedal is actuated only after a constant creep speed is reached (see also 4.6.1). To evaluate launch, the evaluation parameters launch hesitation, t LH and launch, tL and total time t T are used (see chapter 3.3.2 Launch and 3.3.3). Figure 52 displays launch from creep for 4 cases due to the different accelerator pedal position. Some of the description of the figures in the following chapters might use the terms such as “stage” and “transition condition” which are already defined in chapter 4.6 and its subchapters. The 2

acceleration fall to 0 m/s at the end for each curve is due to the speed limiter at maximum speed. Gear upshift to the next gear (here gear 2) is not simulated in the simulation model to evaluate launch / moving off.


55


Figure 52: Launch from creep with simple clutch actuator model (see chapter 4.4.2)

The accelerator pedal is actuated within a period of 0.2 s in each case. Only the end actuation position is different, which are 100 %, 70 %, 50 % and 30 % respectively. Referring to the engine and shaft 1 speed, there is a difference between the curve gradient of each case after the clutch is fully engaged. This is due to the different loads from the engine as a respond to the accelerator pedal position. The maximum acceleration of each case is also different due to the same reason. At 4 s, stage 2 begins due to the accelerator pedal pressed. The engine speed should have increased instantaneously to the desired engine speed as soon as the stage begins. However, the clutch actuation which also begins quite early “drags” the engine shaft and reduces the gradient.


56


Launch from creep with simple actuator model Accelerator pedal position [%]

Launch hesitation tLH [s]

Launch tL [s]

Total time tT [s]

100

0.30

0.52

0.82

70

0.32

0.61

0.93

50

0.31

0.83

1.14

30

0.15

1.02

1.17

Table 9: Evaluation criteria for launch from creep with simple actuator model

The clutch full engagement time is longer the lower the accelerator position is. For accelerator pedal position higher than 50 %, the launch hesitation time of around 0.3 s is achieved. The launch time increases with decreasing accelerator pedal position. Even though a low desired engine speed is set for low accelerator pedal position, the low load causes a late full clutch engagement. Initial acceleration until 30 % accelerator pedal actuation in each case is the same. This is confirmed from the acceleration curve of figure 52 and the launch hesitation at 30 % pedal actuation is half of the rest. The same evaluation was done using the complex clutch actuation model and the following figure 53 was obtained.


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Figure 53: Launch from creep with the complex clutch actuator model (see chapter 4.4.2)

The engine shaft accelerates quite instantaneously to the desired speed as soon as stage 2 begins because of the late begin of clutch actuation. From the shaft 1 speed, it can be deciphered that the maximum creep speed is also reached a little later compared to that of the simple clutch actuator model. The clutch full engagement time is in total longer for every case. The controller parameter for pedal position of 30 % at full clutch engagement is not fully optimised, as displayed by the high acceleration overshoot at around 7 s.


58


Launch from creep with complex actuator model Accelerator pedal position [%]


Launch tL [s]

Total time tT [s]

100

0.75

0.42

1.12

70

0.78

0.49

1.27

50

0.81

0.63

1.44

30

0.74

2.01

2.75

Table 10: Evaluation criteria for launch from creep with complex actuator model

The total time confirms the generally late full clutch engagement from the application of the complex clutch actuator. The launch hesitation is more than twice longer compared to that of the simple clutch actuator due to the limited power and hence the torque limit of the actuator motor. The total time increases around 0.3 s if compared with the total time using the simple clutch actuator model. The application of the complex clutch actuator for the simulation of launch is thus more realistic.


59


Launch from Brake For launch from brake, the accelerator pedal is actuated before vehicle creep begins. One preassumption that can be made is that the full clutch engagement period is longer than that of the launch from creep because of the bigger speed difference between the engine shaft and shaft 1 in the beginning. Launch from brake was also done using both clutch actuator model. The results would then be compared to launch from creep according to their respective actuator model (launch from creep with simple actuator model with launch from brake with simple actuator model, and the same for complex actuator model).

Figure 54: Launch from brake with simple clutch actuator model

Table 11 presents the evaluation criteria for launch from brake with the simple actuator model.

Launch from brake with simple actuator model Accelerator pedal position [%]


Launch tL [s]

Total time tT [s]

100

0.3

0.91

1.21

70

0.33

1.05

1.38

50

0.32

1.43

1.75

30

0.13

2.50

2.63

Table 11: Evaluation criteria for launch from brake with simple actuator model


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The launch hesitation is approximately the same compared to launch from creep, meaning the vehicle reaches the maximum acceleration at approximately the same time. Due to the big speed difference of the engine shaft and shaft 1 in the beginning, the launch period is longer and hence causes the total time taken until full clutch engagement longer too. Launch from brake was also simulated using the complex actuator model. The following curves (figure 55) are obtained as a result of the simulation. The same trend of relation between accelerator pedal position and the full clutch engagement time is detected here. Comparing the following figure to figure 53 (launch from creep with complex clutch actuator), one can see that the controller is optimised for accelerator pedal position 30 %. However, for accelerator pedal position 2

2

of 100 %, an instantaneous sudden deceleration from 5.6 m/s to 4 m/s is detected as soon as the clutch is fully engaged at 2.25 s. This can cause an uncomfortable feeling for the driver due to the high negative rate of change of acceleration. The controller parameter must still be optimised for a more comfortable driving.

Figure 55: Launch from brake with complex clutch actuator model


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For launch from brake with the complex actuator model, launch hesitation and launch is longer compared to launch from creep of the same clutch actuator model. However one anomaly can be detected for the accelerator pedal position 30 %, whereby the values are lower than the values from launch from creep. This is because the controller parameters were not optimised for the situation of launch from creep.

Launch from brake with complex actuator model Accelerator pedal position [%]


Launch tL [s]

Total time tT [s]

100

0.81

0.46

1.27

70

0.87

0.76

1.53

50

0.79

1.03

1.82

30

0.66

1.72

2.38

Table 12: Evaluation criteria for launch from brake with simple actuator model


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Launch on Hill As stated before, launch on hill is simulated from creep. The evaluation criteria for this driving situation are maximum creep speed, time to maximum creep, launch hesitation, t LH and launch, tL. Launch on hill is simulated by varying the gradient (6 %, 10 % and 15 %) at the same accelerator pedal position (here 100 % pedal position). It is simulated using the simple clutch actuator model.

Figure 56: Launch on hill with simple actuator model

The negative acceleration at 1 s is due to the let off of the brake pedal (not shown in the figure). The vehicle starts to creep at 1 s and the accelerator pedal is pressed from 0 % to 100 % at 5 s (also not shown in figure). The speed gradient and the maximum acceleration decreases with increasing hill gradient. As hill gradient increases, the grade resistance on the vehicle increases, resulting in extra drive torque used to overcome the resistance. Hence the rest of the drive torque available for acceleration is reduced. Grade resistance is defined as follows.

    Ahmad Hakim Mohd Sorihan

(5.1)

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where mcar is the mass of the vehicle, g is the standard gravity and α the gradient angle. The vehicle speed has a proportional relation with shaft 1 speed. It is expected that the launch, t L and total time t T is longer due to the low speed gradient at high hill gradient. The following table presents the evaluation criteria for launch on hill.

Launch on hill with simple actuator model Gradient [%]

Time to max. creep [s]

Max. creep speed [km/h]


Launch tL [s]

Total time tT [s]

15

3.02

4.33

0.33

1.02

1.35

10

2.06

5.10

0.31

0.79

1.10

6

1.60

5.12

0.34

0.63

0.97

0 (on plane)

1.16

5.26

0.30

0.52

0.82

Table 13: Evaluation criteria for launch on hill with simple actuator model

The maximum creep speed decreases and the time taken to reach maximum creep increases with an increase in hill gradient. While the launch hesitation are approximately the same for all cases, the launch and hence the total time increases due to the mentioned increase of gradient resistance. The following conclusions were made from the simulation of all variations of launch: 

Launch hesitation is determined by the accelerator pedal position, the control settings of the clutch actuator and the power limit of the actuator itself. Power limit of an actuator motor limits the reaction time of the actuator and hence the time to reach maximum acceleration



By setting a smooth clutch full engagement as a pre-condition, the criterion launch is determined by the accelerator pedal position. The higher the accelerator pedal position, the higher the engine load and subsequently the shorter the launch period.



The criterion launch values for launch from brake are longer than launch from creep due to the big speed difference between the engine shaft and shaft 1 in the beginning.



A smooth full clutch engagement can be detected from a smooth vehicle acceleration profi le during the engagement.



For launch on hill, gradient resistance increases the resistance on the vehicle causing longer period of launch hesitation and launch. Maximum creep speed decreases and time to maximum creep are longer for increasing gradient.


64


5.2.2

Upshift

Comparison of Gear Upshift between the Simple and the Complex Actuator Model To explain the figures and curves under this chapter, terms such as torque phase and speed phase are used. These terms are already explained under chapter 3.3.2 Gear Upshift and Downshift and chapter 4.6.2. Before proceeding with the simulation of upshift, the clutch actuator model must first be chosen between the simple and the complex model. The simple clutch actuator has the advantage of easier parameter setting and short simulation and calculation time. The complex clutch actuator, even though more realistic, consumes a lot of effort for parameter fine tuning and simulation time as well as RAM memory. Simulation for upshift for gear 1 to gear 2 was done using both clutch actuator model. Both results were compared.

Figure 57: Comparison of upshift of gear 1 to gear 2 between the simple and complex actuator model


65


Figure 57 presents a comparison of upshift from gear 1 to gear 2 of the simple and complex clutch actuator. At first glance, the engine and the shaft speeds from simulation are very similar. However, by observing the torque curves closely, one can see that torque phase of the simple actuator is shorter than that of the complex actuator. Clutch 2 torque of the complex actuator are also difficult to control, as can be seen from the overshoot and undershoot around 6.65 s. There is a different control strategy implemented in the torque phase for both clutch actuator model. For the simple actuator model, the torque capacity of clutch 2 is controlled to achieve a constant torque in the speed phase. Clutch 2 is only allowed to increase its torque capacity at the end of its speed stage. For the complex actuator model, clutch 2 torque capacity is allowed to increase to help the deceleration of the engine shaft to the shaft 2 speed. At the end of its speed phase, clutch 2 torque capacity decreases slowly to its normal value. Based on the similar torque and speed profiles of both clutch actuators, it is decided that the simple clutch actuator would be used. A comparison of upshift of gear 1 to gear 2 is already sufficient since the upshift involves a change of high gear ratio between gear 1 and gear 2. The next gear upshifts involve smaller ratios between the gears change and does not have a very significant influence on the torque and speed curves.

Sequential Upshift from Gear 1 to Gear 4 The driving situation upshift was simulated from gear 1 to gear 4 simultaneously. The evaluation of upshift was done and displayed separately, meaning from gear 1 to gear 2, from gear 2 to gear 3 and from gear 3 to gear 4, for different accelerator pedal positions. Figure 58 presents the speed and torque curves of upshift from gear 1 to gear 2 for the accelerator pedal position of 100 %, 70 % and 40 %. The average jerk during the torque and speed phase was calculated using the evaluation App (see chapter 5.1) and the average value was “drawn” as a straight green line to display it alongside the jerk curve. Figure 59 displays the jerk and average jerk curves taken from the evaluation App.


66


Figure 58: Upshift from gear 1 to gear 2 for accelerator pedal position position 40 %, 70 %, 100 %


67


Figure 59: Jerk of upshift from gear 1 to gear 2 for accelerator pedal position 40 %, 70 %, 100 %

Table 14 presents the evaluation criteria for upshift gear 1 to gear 2.

Upshift gear 1 to gear 2 Accelerator pedal [%]

Torque phase time tTP [s]

Speed phase time tSP [s]

Total time tT [s]

Average Jerk [m/s3]

40

0.296

0.39

0.685

-1.15

70

0.3

0.336

0.636

-1.18

100

0.302

0.454

0.756

-1.64

Table 14: Jerk of upshift from gear 1 to gear 2 for accelerator pedal position position 40 %, 70 %, 100 %


68


Through a careful observation of the speed curves, one can notice that the torque phase begins at different engine and shaft 1 speed. This is due to the gearshift diagram (see Figure 32). 32). The reason of the implementation of the gearshift diagram is also explained under the chapter 4.4.3. chapter 4.4.3. During the torque phase, clutch 1 is controlled to reduce its torque capacity constantly until zero in the duration of 0.3 s. A value around 0.3 s is achieved for torque phase time of all three accelerator pedal position, signifying a good controller parameter setting. The speed phase time meanwhile is influenced by the value of engine torque down. The lower the engine torque is brought down during the speed phase, the higher the deceleration of the engine speed to the shaft 2 speed, thus the shorter the speed phase is. This is also seen from the slight difference of the engine speed gradient. The different values of speed phase time for different accelerator pedal positions reflects the different desired value of engine torque down used in each case. Thus, the shifting time is also different for each case. The average jerk is calculated for the duration of the shifting time, which is around 0.7 s. This method is deemed sufficient because the period of torque and speed phase altogether (around 0.7 s here) is quite short. A more accurate improvement to the calculation of average jerk can be done by calculating the average jerk during torque phase and speed phase separately. This is because high amplitude jerk is detected mainly during the torque phase. Jerk during the speed phase has a value nearly zero except for 3 to 4 high amplitude peaks. An early observation from the jerk values can be made, whereby the higher the accelerator pedal position is, the bigger the value of jerk. For upshift from gear 2 to gear 3, clutch 1 is controlled to increase its torque capacity from zero to the initial level of clutch 2 torque in 0.3 s, while clutch 2 is controlled to maintain a constant slip between the engine shaft and shaft 2. The torque phase is set to end when clutch 2 torque is equal and smaller than zero. Thus, the torque phase time here is determined by how accurate the controller parameter setting of clutch 2. The duration of speed phase is determined by the engine torque down level as before. Figure 60 presents the speed and torque curves during the upshift while figure 61 presents the respective jerk curves.


69


Figure 60: Upshift from gear 2 to gear 3 for accelerator pedal position 40 %, 70 %, 100 %


70



Table 15 presents the evaluation criteria for upshift gear 2 to gear 3.




Shifting time tS [s]

Average Jerk [m/s3]

40

0.349

0.385

0.734

-0.71

70

0.324

0.308

0.632

-1.1

100

0.313

0.421

0.734

-1.2

Table 15: Jerk of upshift from gear 2 to gear 3 for accelerator pedal position 40 %, 70 %, 100 %

The evaluation parameters for all cases do not show a definite trend in relation with the accelerator pedal position, besides the parameter jerk. The torque phase time is considerably longer compared to the value during upshift from gear 1 to gear 2. This shows a slightly unoptimised clutch 2


71


controller parameters during upshift from gear 2 to gear 3 and consequently causes the shifting time longer. The values of average jerk during this upshift are lower than during upshift of gear 1 to gear 2. The controlled reduction of clutch 2 torque is not as uniform as expected in the beginning and at the end of the torque phase. This is the main cause of the high amplitude jerk during the torque phase. Improvement can be made through better PID-controller settings. An easier improvement is by altering the desired torque increase of clutch 1, so that the increase in the beginning is at a lower gradient and then followed by a steep increase to the desired torque. Finally, the upshift for gear 3 to gear 4 was simulated. Clutch 1 is controlled to reduce its torque capacity to zero while clutch 2 is controlled to maintain a constant micro slip between the engine shaft and shaft 1, as during the upshift from gear 1 to gear 2. Figure 62 presents the speed and torque curves during the upshift while figure 63 presents the respective jerk curves.


72


Figure 62: Upshift from gear 3 to gear 4 for accelerator pedal position 40 %, 70 %, 100 %


73



Table 16 presents the evaluation criteria for upshift of gear 3 to gear 4.





Average Jerk 3 [m/s ]

40

0.296

0.336

0.632

-0.59

70

0.302

0.249

0.551

-0.81

100

0.305

0.373

0.678

-0.95

Table 16: Evaluation criteria for upshift from gear 3 to gear 4 for accelerator pedal position 40 %, 70 %, 100 %

The torque phase time achieved for upshift from gear 3 to gear 4 is similar to upshift from gear 1 to gear 2 as expected. The speed phase time for accelerator pedal position 70 % is clearly lower


74


compare to other accelerator pedal position, causing also consequently a short shifting time. This is due to the high torque down value during its speed phase which is also displayed in figure 62. The amplitude of jerk during this gear upshift is smaller compared to the other lower gear upshifts. 3

As displayed in figure 63, the maximum jerk amplitude is around ± 5 m/s . As before, the high amplitude jerk mainly takes place during the torque phase. From the simulation of the sequential upshift from gear 1 to gear 4, several conclusions can be made: 

The torque phase time is influenced by the controller parameter settings. An optimised controller settings can bring the torque phase time nearer to the desired value.



The speed phase time is determined by the value of engine torque down. The higher the engine torque down, the shorter the speed phase time is.



High amplitude jerk takes place mainly during the torque phase. An improvement of average jerk calculation can be done whereby the average jerk is calculated separately for torque and speed phase.



Taking only an upshift into account (e.g. upshift gear 1 to 2), the average jerk increases the higher the accelerator pedal position is, due to a high value of torque transferred from the off going to the oncoming clutch during the upshift.



Taking all upshifts into account at the same accelerator pedal position, upshift at high gears displays lower average jerk compared to upshift at lower gears. This is due to the low gear ratio jump high gear upshift.


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5.2.3

Downshift

Two types of downshift are simulated in this thesis, namely the power on downshift and power off downshift. Information about both downshift controls and descriptions are available under chapter 4.6.3.

Power On Downshift Power on downshift is simulated from gear 4 to gear 3 due to extra torque demand from the driver. The torque demand is recognised from the abrupt accelerator pedal position change that crosses the downshift curve in the gear shifting diagram (see chapter 4.6.3 and Figure 46). The speed phase takes place first, and then the torque phase during power on downshift. Two variants of power on downshift were simulated. Both variants begin at 50 % accelerator pedal position at the same time but one ends at 70 % and the other at 100 % accelerator pedal position. The period of the increase in accelerator pedal position is 0.3 s. The differences in torque phase time, speed phase time and average jerk values were o bserved. Figure 62 presents the speed and torque curves during the power on downshift while figure 63 presents the respective jerk curves.


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Figure 64: Power on downshift from gear 4 to gear 3


77


Figure 65: Jerk of power on downshift from gear 3 to g ear 4 for different accelerator pedal position change

Table 17 presents the evaluation criteria for the power on downshift from gear 4 to gear 3 .

Power on downshift from gear 4 to gear 3 Accelerator pedal [%]

Speed phase time tTP [s]

Torque phase time tSP [s]


Average Jerk [m/s3]

50-100

0.159

0.247

0.406

3.399

50-70

0.171

0.227

0.398

1.888

Table 17: Evaluation criteria for power on downshift for gear 4 to gear 3

The torque curves with the accelerator pedal position change from 50 % to 100 % displays a steeper increase of engine torque due to its higher end position compared to the change from 50 % to 70 %. Due to the same reason too it enters the speed phase and subsequently the torque phase first. Clutch 2 torque capacity reduction to approximately half of its initial value helps to increase the engine speed to shaft 1 speed without the help from engine torque (engine torque up). The speed phase time and torque phase time are approximately the same in both cases, differing no more than 0.02 s. Thus, the shifting times for both of them are also approximately the same. The Ahmad Hakim Mohd Sorihan

78


relation between jerk and torque is also seen here, whereby the average jerk at higher engine torque is higher compared to at lower engine torque. The average jerk values recorded are also bigger compared to those recorded during upshift. The jerk curves show the same trend as in other gear shifting situation: high amplitude jerk mainly takes place during the torque phase, which starts around 2.84 s for the accelerator pedal position change from 50 % to 100 % and 2.98 s in the other case. Several conclusions can be made from the simulated power on downshift: 

Speed phase time is determined by the initial reduction level of the off going clutch capacity and the engine torque increase due to the acceleration demand from the driver



Torque phase time is determined by the controller parameter settings. The faster the oncoming clutch controller reacts, the shorter the torque phase time is.



Higher average jerk occurs at higher accelerator pedal position change.

Power Off Downshift During power off downshift at 0 % accelerator pedal position, the clutch experiences a negative torque due to the drag torque of the ICE and the rotating components of the powertrain. The simulation of power off downshift was done for the gear 4 to gear 3 and the gear 2 to gear 1. Contrary to power on downshift, the torque phase takes place first followed by the speed phase during power off downshift. The speed and torque curves for the power off downshift are as displayed in Figure 50 under chapter 4.6.3 Power Off Downshift. Figure 66 displays the torque and speed curves for the power off downshift of gear 4 to gear 3 while figure 67 displays the jerk curves for both cases simulated. The simulations were done at 0 % accelerator pedal position.


79


Figure 66: Power off downshift for gear 4 to gear 3

Figure 67: Jerk during power off downshift from gear 4 to gear 3 and gear 2 to gear 1

Table 18 presents the evaluation criteria for the power on downshift from gear 4 to gear 3 .


80


Power off downshift at 0 % accelerator pedal position Gear shift [-]

Torque phase time tSP [s]

Speed phase time tTP [s]


Average Jerk 3 [m/s ]

4-3

0.34

0.75

1.09

-0.17

2-1

0.33

0.55

0.88

-0.27

Table 18: Evaluation criteria for power off downshift for gear 4 to gear 3 and gear 2 to gear 1

The torque phase time in both cases are approximately the same. However, the speed phase time for power off downshift gear 2 to gear 1 is shorter compared to the other. As one can see, the clutch actuation forces in both Figure 66 and Figure 50 are at the same level. The power off downshift occurs at a slightly higher engine speed, causing a slightly higher negative torque at the beginning of the speed phase. Thus, a slightly longer time is needed to synchronise the engine shaft to the shaft 1 speed if the same actuation force is applied. The high jump in ratio at low gear power on downshift causes the higher average jerk values compared to the high gear power off downshift. High amplitude jerks of power on downshift gear 4 3

to gear 3 are in the range of ± 5 m/s while the power off downshift of gear 2 to 1 exhibits jerks 3

bigger than ± 10 m/s . As before, high amplitude jerks occur mainly during the torque phase. The following points are concluded from the simulated power off downshift: 

Torque phase time is determined by the controller parameter settings. The faster the oncoming clutch controller reacts, the shorter the torque phase time is.



For the same level of clutch actuation force, speed phase time is determined by the engine speed dependant torque.



The trend of high average jerk at low gear shifting also applies here.


81


5.3 Comparison of Simulated Driving Situations with the Real Measurement Data The simulation results were briefly compared to the measured data from test drives provided by hofer. In fact, comparisons were done continually as a mean to improve the parameter settings of the simulation model. The comparisons of launch and upshift from gear 1 to gear 2 are displayed in this chapter.

Figure 68: Launch comparison between measured data and simulation

Figure 68 presents speed and acceleration curves during launch. The speed curve gradient and acceleration differences between the measured and the simulation results indicate a slightly different torque capacity of the clutch. This can be due to the difference in the engine torque map or


82


the difference in inertia of the rotating components and the whole weight of the vehicle. The oscillation of the shaft 1 speed of the measured results indicates a considerably low stiffness at the transmission shaft side. The measured acceleration curve displays a sink in acceleration during full clutch engagement. The acceleration fall however does not occur at a steep gradient and hence will not be felt as “uncomfortable” for the driver.

The evaluation criteria defined for the simulation can also be implemented for measured data results. Even though the measured acceleration of the vehicle seems to increase gradually until the full clutch engagement, constant acceleration can still be seen around 5.1 s. Thus, launch hesitation time can be calculated. Launch time can be calculated normally.

Figure 69: Upshift gear 1 to gear 2 comparison between measured data and simulation


83


Figure 69 presents speed and engine torque curves for the upshift of from gear 1 to gear 2. The measured engine torque displays a slightly lower engine torque value compared to the simulated engine torque. The reason is the same as the reason for the acceleration difference of the figure 68. Unfortunately, the torque on each sub gearbox shaft is not available. However, the speed phase can be approximately identified from the torque down duration.

Figure 70: Comparison of acceleration between measured and simulation

Another way to approximately identify the torque phase is to observe the acceleration during gear upshift. The torque phase starts as the acceleration begins to reduce and ends as the acceleration reaches the minimum value. Likewise the speed phase begins at the minimum acceleration and ends before the acceleration falls again as the engine and the oncoming shaft reaches speed equality and the oncoming clutch fully engage (see also chapter 3.3.2 Gear Upshift and Downshift).


84

Conclusions and Future Improvements

6


This chapter discusses the contribution and the tasks that this thesis cover on the simulation based comfort evaluation of an automated transmission. The simulation was done using a type of an automated transmission, namely the dry dual clutch transmission (DDCT). At the end of this chapter, future improvements and possible research activities are examined. This thesis mainly consists of 3 main parts. In the first part of the thesis, several driving situations where the “driving comfort” is mainly affected were listed. Several test drives were performed for

the author to experience the listed driving situations and hence increase the understanding of evaluation of comfort through subjective perception. Then, relevant evaluation criteria were defined for the driving situations. Subsequently, suitable driving situations were chosen considering the time constraint of the thesis and the difficulty to realise such driving situations in a simulation program. In the second part of this thesis, the simulation process was firstly planned. The evaluation criteria, variable parameters and simulation model complexity were listed in a table for each chosen driving situations to give an overview of the simulation tasks. Next, simulation models were built for each driving situation and continually improved in terms of model component parameters and control parameters. The improvement of the simulation model was not done intensively to not stray away from the main objective of this thesis. Scripts were written to automate the obtaining of evaluation parameters from the simulation. In total there were 3 main driving situations simulated with further variations of them. They are launch, upshift and downshift. Launch is further divided into launch from creep, launch from brake and launch on hill. Downshift is divided to power on downshift and power off downshift. With the application of the app, the evaluation criteria were successfully read and extracted from the simulation. The extracted evaluation criteria of each driving situation return plausible values that can be related to the simulation parameters. Conclusions were made at the end of each simulation to determine the relationship of the evaluation criteria to the simulation parameters. An improvement can be made to one evaluation criterion during gear shift, namely the jerk (or shock intensity). In the simulation, total average jerk of both torque and speed phases were calculated. F rom the simulation results, it was recognised that jerk levels during torque phase are significantly higher in amplitude compared to during the speed phase. So, average jerk can be calculated separately for each phase. To conclude the evaluation chapter, a short comparison was done between the simulation results and the test drive results from hofer.


85


A lot of improvements in terms of parameter settings can be made for the simulation model to achieve more realistic results. In addition to the parameter settings, the simulation model can be used as a basis for simulation of further driving situations listed under the chapter 3.3.2. Alternative control strategy can also be implemented and the results can be compared to the present control strategy. One example is the torque down control during gear upshift. Instead of using torque down, the oncoming clutch torque capacity is increased to accelerate the engine shaft to the oncoming shaft speed. Such measures can be done by using the current simulation model as basis. The long term aim of this thesis is to provide a knowledge of simulation based comfort evaluation. This thesis does provide the foundation knowledge of the listed simulated driving situations and also the simulation model basis of the powertrain of a vehicle.


86

Reference Index

7 [1]

Reference Index Naunheimer, H.; Bertsche, B.; Ryborz, J.; Novak, W.: Automotive Transmissions Fundemental, Selection, Design and Application. Springer Heidelberg Dordrecht London New York, 2011

[2]

Kimmig, K.L.; Agner, I: Double clutch – Wet or dry, that is the question, LuK SYMPOSIUM, 2006

[3]

Dong, P.; Tenberge, P.; Qu, W.; Dai, Z.: Optimized shift-control in automatic transmissions with respect to efficiency, shift loads and comfort. VDI-Berichte 2158, 2012

[4]

Wagner, U.;Berger, R.; Ehrlich, M.; Homm, M.: Electromotoric actuators for double clutch transmissions - Best efficiency by itself. LuK SYMPOSIUM, 2006

[5]

Kirchner, E.; Leistungsübertragung in Fahrzeuggetrieben – Grundlagen der Auslegung, Entwicklung und Validierung von Fahrzeuggetrieben und deren Komponenten. Springer Berlin Heidelberg New York, 2007

[6]

Werneke, J.; Kassner, A.; Vollrath, M.: An Analysis of the Requirements of Driver Assistance Systems – When and why does the driver like to have assistance and how can this assistance be designed?, DLR Journal 2007

[7]

Aversa, P.; DeVincent, E.: Evolution and outlook - Powershift DCT250. Getrag Presentation 2010

[8]

Kimmig, K.L.; Bührle, P; Henneberger, P.; Ehrlich, M.; Rathke, G.; Martin, J.: Dry Double Clutch – Success with efficiency and comfort. LuK SYMPOSIUM 2010

[9]

Wikipedia article: Ford Sigma engine – Duratec Ti-VCT. Link: http://en.wikipedia.org/wiki/Ford_Sigma_engine#Duratec_Ti-VCT (22.12.2012)

[10]

PDF from Internet: Der neue Ford Focus – Auto Online Magazin. Link: http://www.auto-online-magazin.de/pdf/focus_technik.pdf (22.12.2012)

[11]

Internet article: The New Ford Duratec 1.6L Ti-VCT Engine – ATZ Online Link: http://www.atzonline.com (22.12.2012)

[12]

Andersson, S.; Söderberg, A.; Björklund, S.: Friction Models for Sliding Dry, Boundary and Mixed Lubricated Contacts. Journal Science Direct 2006


87

Reference Index

[13]

LMS.Imgine.Lab AMESim Help Files (Internal Software) Link: http://nupet.daelt.ct.utfpr.edu.br/_ontomos/paginas/AMESim4.2.0/doc/ (19.02.12) LMS.Imgine.Lab AMESim Website: http://www.lmsintl.com/LMS-Imagine-Lab-AMESim (19.02.12)

[14*]

Hagerodt,A: Automatisierte Optimierung des Schaltkomforts von Automatikgetrieben. Shaker Verlag 2003

[15*]

Ni, C.; Lu, T.; Zhang, J.: Gearshift Control for Dual Clutch Transmissions. WSEAS 2009

[16*]

Wang, Y: Optimal Gear Shifting Strategy for a Seven-speed Automatic Transmission Used on a Hydraulic Hybrid Vehicle. Master Thesis University of Toledo 2012

[17*]

Kulkarni, M.; Shim, T.; Zhang, Y.: Shift Dynamics and Control of Dual Clutch Transmissions. Science Direct 2006

[18*]

Goetz, M.; Levesley, M. C.; Crolla, D. A.: Dynamics and Control of Gearshifts on Twin-Clutch Transmissions. Dissertation University of Leeds 2005

[19*]

Dylla, S.: Entwicklung einer Methode zur Objektivierung der subjektiv erlebten Schaltbetätigungsqualität von Fahrzeugen mit manuellem Schaltgetriebe. Dissertation Institut für Produktentwicklung KIT 2009

[20*]

Schäfer, J.: Wirkungsgraduntersuchung verschiedener Steuerungs- und Aufbaukonzepte eines automatisch geschalteten KFZ-Triebstrangs mit Hilve von Simulationmodellen. Diplomarbeit hofer/Universität Stuttgart 2008


88

Appendix

8 Appendix 8.1 AMESim Submodels Used in Simulation Figure

Name

Description

Constant signal source

Used to generate constant value

Piecewise linear signal source

Consist of 8 stages and used to generate piecewise linear signals like ramp, steps, squares etc.

Gain

The output signal is formed by multiplying the input signal at input by a user specified gain

Variable gain

Variable gain is activated by binary 1

Differentiator

Produces a signal output which is an approximation of the derivative of the input by using a time constant

First order lag

First order lag with gain and positive time constant setting

PID controller Variable PID controller

PID controller with options of P,PI, PD, and PID modes & anti windup methods. Variable PID enables user to change parameter values during simulation

Function of input block

Output as a function of input. Function can be defined by user. Also support logic function.

Function of input defined by ASCII file

Output as interpolation of input. Interpolation function provided by user‘s ASCII file

Continuous delay

Continuous delay with user specified delay time.

Saturation Variable saturation

Dead band Inverted dead band

Two way ideal switch signal


Upper and lower limit specified by the user. Variable saturation enables variable limit during simulation.

User defined maximum and minimum dead zone value. Useful to avoid zero division.

Contains 2 input s, 1 output and 1 input treshold. Output controlled by input treshold

89

Appendix

Figure

Name

Description

Bus source

Ideal bus source channel. Creates a BUS line.

Add signal to BUS

Adds an item or another BUS line into a BUS.

Get signal from BUS

Get value from BUS

Transmitter Receiver

Transfer and receive variables without any visible connection

Table 19: AMESim Signal and Control library

Figure

Name

Description

Zero force source

Supplies zero/no force.

Zero linear speed source

Zero velocity source. It also provides constant values of displacement and acceleration.

Force transducer

Force transducer equipped with offset and gain function. Supplies force value in signal.

Linear displacement transducer

Linear displacement transducer with offset and gain function. Supplies displ. value in signal.

Power, energy and activity sensor

To calculate energy/power dissipation of a part, connect it before and after the part and substract values with each other.

Linear mechanical node

Enables two linear shafts to be connected to another

Linear elastic end stop

Elastic contact between two bodies capable of linear motion. Gap between bodies defined by user

Torque converter

Converts signal input to torque

Torque transducer

Torque transducer equipped with offset and gain function. Supplies torque value in signal.

Rotary speed transducer

Rotary speed transducer equipped with offset and gain function. Supplies speed value in signal.

Zero torque source

Supplies zero/no torque.

Zero angular speed source

Zero angular speed source

Rotary load with friction


Rotary load model with external torques applied to ist two ports. There is provision for viscous friction, Coulomb friction and stiction.

90

Appendix

Figure

Name

Description

Rotary spring and damper

Ideal spring damper system for a rotary shaft. Spring stiffness and damper rating set by user.

Variable friction between rotary parts Variable friction between a fixed part and a rotary part

Rotary friction torque generator modelled by Coulomb friction with option of tanh, dahl or Lugre model. Signal form: 0 to 1, force or friction torque

Rotary mechanical node

Enables two rotary shaft to be connected to one another

Rack and pinion

converts a linear displacement x1 into an angle theta2 , a linear velocity v1 into an angular velocity w2 and a torque t2 into a force f1 with a transformation ratio set by the user

Table 20: AMESim Mechanical library

Figure

Name

Description

Half synchronizer

Ideal synchronizer for coupling with idle gears. 2 synchronizer model available: Friction model & stiction model

Rotary link

Rotary link between gears. Normally used for linkage in planetary gear system.

Tyre and wheel

Models and generates the contact force at the tyre/road interface and rolling resistance for driving situations

Road profile

Presents the option of modelling with pl ane or slope road. Slope road is also provided with radius of cyllindric section of the slope

Vehicle suspension

models suspension damping and stiffness as well as spindle mass in 3 planar degrees of fr eedom: longitudinal translation, vertical translation and self-rotation


91

Appendix

Figure

Name

Description

Gear with 2 rotary ports and 1 gear port

Gear with 2 rotary ports and 2 gear port

Simple gear model with radius setting. Used to construct simple gear systems in a transmission

Idle gear with 3 rotary ports and 1 gear port

Idle gear model with radius sett ing. Used to construct simple gear systems in a transmission. Used together with half synchronizer and other gear submodels

2D vehichle with 2 axles

2d Submodel of carbody with 2 axles &3 DOF: - Pitch rotation - Longitudinal translation - Vertical translation Use with suspension & tyre and wheel submodel

Table 21: AMESim Powertrain library

Figure

Name

Description

SFC enable/disable action supercomponent

Logical condition to enable/disable a supercomponent. Connected to SFC control interface

SFC action end

Used to model an end to a step

SFC control interface

Used to enable/disable supercomponent interface. Must be embedded in the supercomponent.

SFC step

Used to model a step. In usage connected to SFC action

SFC initialization and system connexion

Connects the SFC to the BUS and define action states value of the SFC system

SFC alternative branch

Provides an alternative branch of different transition condition

SFC transition

Used to model a transition. Connected to SFC logic for transition condition definition

SFC logic condition expression

Used to model logical transition condition set by user

SFC action

Used to model action executions when a step is activated


92

Appendix

Figure

Name

Description

SFC parallel branch

Used to split process to two parallel branches

Table 22: AMESim Sequential Functional Chart (SFC) library

8.2 Table for Subjective Evaluation of Driving Situations Vibrations on the gear selection lever in neutral position Rating 1

2

3

4

5

6

7

8

9

10

Idle RPM

x

2000 1/min

x x

4000 1/min

Notes:

Gear engagement and disengagement noises Rating 1

2

3

4

5

6

7

8

9

0 --> 1

x

1 --> 0

x

0 --> R

x

R --> 0

x

10

Notes:

Shunting Rating 1

2

3

4

5

6

7

8

9

10

9

10

x

0 --> 1 --> 0 --> 1

Notes:

Engagement when launching on 2nd gear Rating 1

2

3

4

5

6

7

8

Low gas pedal (F)

x

High gas pedal (F)

x

Low gas pedal (R)

x

High gas pedal (R)

x

Notes:


93

Appendix

Forward launch on grade (uphill) Rating 1

2

3

4

5

6

7

Low gas pedal

x

High gas pedal

x

8

9

10

8

9

10

8

9

10

8

9

10

Notes:

Reverse launch on grade (uphill) Rating 1

2

3

4

5

6

7

Low gas pedal

x

High gas pedal

x

Notes:

Hill hold-launch Rating 1

2

3

4

5

6

Low gas pedal

x

High gas pedal

x

Notes:

7

Launch with hill hold is a little aggressive with clear jerk

Creep Rating 1

2

3

4

5

6

7

uphill (F)

x

uphill (R)

x

Notes:

Creep at 2nd gear with approx same speed. If more torque needed, auto-shift to 1st gear

Shifting (upwards) Rating 1 1 --> 2

2

3

4

5

6

7

8

9

10

x

2 --> 3

x

3 --> 4

x

4 --> 5

x

Notes:


94

Appendix

8.3 Simulation Model Basis

Figure 71: Simulation model basis for launch


95

Appendix

Figure 72: Simulation model basis for upshift


96

Appendix

Figure 73: Simulation model basis for downshift


97

Appendix

8.4 App Interfaces

Figure 74: App interface for launch with s imple actuator model

Figure 75: App interface for launch with complex actuator model


98

Appendix

Figure 76: App interface for upshift (shift time)

Figure 77: App interface for upshift (jerk)


99

Appendix

Figure 78: App interface for power on downshift (shift time)

Figure 79: App interface for power on downshift (jerk)


100

Appendix

Figure 80: App interface for power off downshift (shift time)

Figure 81: App interface for power off downshift (jerk)


101

Appendix

8.5 Python Code Snippets

Figure 82: Code snippet for basic plotting app class


102

6DCT250_INGLES.pdf

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