University of Technology, Sydney

Faculty of Engineering and Information Technology

Research of Two Speed DCT Electric Power-train and

Control System

A thesis submitted for the degree of

Doctor of Philosophy

Bo Zhu

(June 2015)

1

CERTIFICATE OF ORIGINALITY I certify that the work in this thesis has not previously been submitted for a

degree nor has it been submitted as part of the requirements for a degree

except as fully acknowledged within the text.

I also certify that the thesis has been written by me. Any help that I have

received in my research work and the preparation of the thesis itself has

been acknowledged. In addition, I certify that all information sources and

literature used are indicated in the thesis.

Bo Zhu

29 June 2015

2

ACKNOWLEDGEMENTS

My sincere thanks is extended to my supervisor, Professor Nong

Zhang His extraordinary knowledge and guidance has been invaluable. I

am also very grateful for the expert assistance of my co-supervisors Dr

Paul Walker and Dr ZhanWenzhang from the BAIC Group. Their ongoing

assistance and support throughout my candidature has been greatly

appreciated.

The team at NTC Powertrains –Simon Fitzgerald, Joef Lou Villanueva

and Simon Cowling were enormously helpful. Not only did they provide

information that was critical to the success of this work but their ideas

effectively inspired this project. I am also indebted to my UTS colleagues:

JiaGengRuan, Xingxing Zhou, Holger, Christopher, ZhenLuo,

YuWang,Cliff, Guangzhong Xu and many others along the way. Their

advice, humour and knowledge sustained me throughout my studies.

My deep gratitude goes to my parents for their constant love, support

and advice.

Finally, I wish to acknowledge my beautiful and loyal wife, YaWei

Sun, who was constantly there for me over the course of my candidature. I

couldn’t have done this without her.

Financial support for this project was provided jointly by BAIC Motor

3

Electric Vehicle Co.Ltd, the Ministry of Science and Technology, China,

and the University of Technology, Sydney.

4

CONTENTS

CERTIFICATE OF ORIGINALITY ...................................................... 1

ACKNOWLEDGEMENTS ..................................................................... 2

CONTENTS .............................................................................................. 4

LIST OF FIGURES.................................................................................. 8

LIST OF TABLES ................................................................................. 14

GLOSSARY OF TERMS AND NOTATION ....................................... 16

ABSTRACT ............................................................................................ 20

Chapter 1 Introduction ............................................................................. 22

1.1 Background ................................................................................... 22

1.2 Project Statement ......................................................................... 30

1.3 Project Objectives ......................................................................... 31

1.4 Project Scope ................................................................................ 31

1.5 Presentation of This Thesis ........................................................... 33

1.6 Publications .................................................................................. 35

References .......................................................................................... 37

Chapter 2 Two Speed DCT Electric Power-train Structure Analysis and

Manufacture ............................................................................................. 40

2.1 Introduction of Two Speed DCT Structure .................................. 40

5

2.2 Simulation Platform Building ..................................................... 41

2.3 Two Speed DCT Power-train Matching and Analysis .................. 45

2.3.1 Q60EV-DCT Matching and Calculation ................................. 45

2.3.2 C70EV-DCT Matching and Calculation .................................. 50

2.3.3 Matching Conclusion ............................................................ 51

2.4 Two Speed DCT Prototype Manufacture .................................... 51

2.4.1 Two Speed DCT Prototype Manufacture .............................. 51

2.4.2 Spin Testing .......................................................................... 57

2.5 Novel Two Speed Electric Power-train System ........................... 61

2.5.1 Two Motor Two Speed System Matching ............................. 61

2.5.2 Simulation Result and Analysis ............................................. 65

2.6 Conclusions ................................................................................ 67

References .......................................................................................... 68

Chapter 3 Multi-Speed Electric Power-train Shifting Schedule ................. 69

3.1 Introduction .................................................................................. 69

3.2 Dynamic Shift Schedule Development for Multi-Speed Pure Electric

Vehicles .............................................................................................. 70

3.3 Economic Shift Schedule development for Multi-Speed Pure

Electric Vehicles .................................................................................. 73

3.4 Simulations and Analysis ............................................................... 77

3.5 Conclusions ................................................................................... 82

References .......................................................................................... 83

6

Chapter 4 Two Speed DCT Shifting Control Strategy ................................. 85

4.1 DCT Shifting Control Analysis ........................................................ 85

4.1.1 Shifting Process of PEV DCT ................................................. 86

4.1.2 Shifting Quality Criterion ...................................................... 89

4.2 Two Speed DCT Transient Modeling .............................................. 91

4.3 Shifting Control Strategy ............................................................... 94

4.3.1 Power-on up-shift control .................................................... 95

4.3.2 Power-off up-shift control .................................................. 103

4.3.3 Power-On Down-Shift Control ............................................ 107

4.3.4 Power-off Down-shift control ............................................. 110

4.4 Shifting Control Strategy with Motor Braking Torque Control ..... 114

4.5 Conclusions ................................................................................. 116

References ........................................................................................ 117

Chapter 5 Rig Testing .............................................................................. 120

5.1 Testing Rig Design and Analysis ................................................... 120

5.1.1 Introduction of Testing Rig ................................................. 120

5.1.2 Testing Rig Parameter Matching ......................................... 122

5.2 Testing Rig Development ............................................................ 128

5.2.1 Frame development ........................................................... 128

5.2.2 Power supply development ................................................ 131

5.2.3 Installation ......................................................................... 132

5.3 Rig Testing Criterion .................................................................... 133

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5.4 Control System Development Based on Rapid Control Program . 135

5.4.1 Rapid Control Prototyping Technology Introduction .......... 135

5.4.2 Hardware ........................................................................... 137

5.4.3 Software ............................................................................ 140

5.5 Rig Testing Results and Analysis .................................................. 146

5.5.1 Shifting Control Testing ...................................................... 146

5.5.2 Temperature Testing .......................................................... 150

5.5.3 Driving Cycle Testing .......................................................... 152

5.5.4 Efficiency Testing ................................................................ 156

5.6 Conclusions ................................................................................. 158

Chapter 6 Vehicle Integration and Road Testing ..................................... 160

6.1 Vehicle Integration ...................................................................... 160

6.2 On Road Calibration .................................................................... 163

6.3 On Road Testing .......................................................................... 166

6.3.1 Dynamic Performance Testing ............................................ 166

6.3.2 Economic Performance Testing .......................................... 166

6.4 Conclusions ................................................................................. 167

Chapter 7 Thesis Conclusions ................................................................. 168

7.1 Summary of the Thesis ............................................................... 168

7.2 Summary of Findings and Contributions ..................................... 169

7.3 Limitations to Research .............................................................. 172

7.4 Future Research .......................................................................... 173

8

LIST OF FIGURES

Fig.1- 1 Dual Clutch Transmission Structure .................................... 27

Fig.2- 1 Two Speed DCT Electric Power-train.................................... 40

Fig.2- 2 DQ250 ................................................................................. 41

Fig.2- 3 Simulation platform based on Simulink-AMESim ................ 44

Fig.2- 4 Motor Working Points NEDC/UDDS .............................. 47

Fig.2- 5 Constant Working Points Efficiency Analysis ........................ 48

Fig.2- 6 Removed 3nd shaft (5th and 6th gear) and reverse shaft. ... 52

Fig.2- 7 Gear Modification ............................................................... 52

Fig.2- 8 Lubrication System .............................................................. 52

Fig.2- 9 Original Hydraulic System .................................................... 53

Fig.2- 10 Modified Hydraulic System ................................................ 54

Fig.2- 11 Hydraulic Valve Body Modification .................................... 55

Fig.2- 12 Original Oil Pump .............................................................. 55

Fig.2- 13 12 Volt E-motor drives pump ............................................ 55

Fig.2- 14 Original Pressure Sensor ................................................... 56

Fig.2- 15 Speed Sensor..................................................................... 56

Fig.2- 16 Oil temperature sensor ..................................................... 56

Fig.2- 17 E-Motor side (long shaft) and Gearbox side (short shaft) .. 57

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Fig.2- 18 Spin Testing Rig ................................................................. 57

Fig.2- 19 Clutch Pressure Sensor Testing .......................................... 58

Fig.2- 20 Output Speed Sensor Testing ............................................ 58

Fig.2- 21 Hydraulic Testing (with mains powered pump).................. 59

Fig.2- 22 Main Solenoid Testing ....................................................... 60

Fig.2- 23 Clutch Activation Solenoid Testing ..................................... 60

Fig.3- 1 Vehicle acceleration curves for establishing single parameter

shift map.................................................................................... 72

Fig.3- 2 Dynamic upshift and downshift map for PEV ...................... 73

Fig.3- 3 Efficiency MAP of motor in 1st and 2nd gear relative to vehicle

speed ......................................................................................... 75

Fig.3- 4 Economic shifting points for output Torque T0 .................... 76

Fig.3- 5 Economic shifting schedule curve for PEV ........................... 76

Fig.3- 6 Adjusted Economic shifting schedule curves for PEVs ......... 77

Fig.3- 7 Acceleration Performance of PEV for different shift schedules

.................................................................................................. 79

Fig.3- 8 NEDC Cycle .......................................................................... 80

Fig.3- 9 UDDS Cycle .......................................................................... 80

Fig.3- 10 Motor Working Points in NEDC ...................................... 81

Fig.3- 11 Motor Working Points in UDDS ......................................... 81

Fig.4- 1 Dynamic Model of Pure Electric DCT ................................... 92

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Fig.4- 2 Shifting Condition Judgment ............................................... 95

Fig.4- 3 Power-On Up-Shifting Process Analysis ............................... 96

Fig.4- 4 Control Algorithm of Power-on Up-shift .............................. 98

Fig.4- 5 Power-on Up-shift Control .................................................. 99

Fig.4- 6 Power-on Up-shift Simulation Results ............................... 100

Fig.4- 7 Simulation results under different clutch slip rotate speed 101

Fig.4- 8 Simulation results under different motor minimum torque limit

................................................................................................ 101

Fig.4- 9 Simulation results under different motor minimum torque limit

................................................................................................ 102

Fig.4- 10 Power-Off Up-Shift Process Analysis ............................... 103

Fig.4- 11 Control algorithm of Power-Off Up-Shift ......................... 105

Fig.4- 12 Power-off Up-shift Control .............................................. 105

Fig.4-13 Power-off Up-shift Simulation Results ............................. 106

Fig.4- 14 Power-On Down-Shift Control Process Analysis............... 107

Fig.4- 15 Control algorithm of Power-on Down-shift ..................... 108

Fig.4-16 Power-on Downshift Control ........................................... 109

Fig.4- 17 Power-on Downshift Simulation Results .......................... 109

Fig.4- 18 Power-off Down-shift Control Process Analysis ............... 111

Fig.4-19 Control algorithm of Power-off Down-shift .................... 112

Fig.4- 20 Power-off Downshift Control ........................................... 113

Fig.4- 21 Power-off Downshift Simulation Results ......................... 113

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Fig.4-22 Power-On Up-Shift control with Motor Braking Control .. 114

Fig.4-23 Power-on Up-shift Control with Motor Braking Torque ... 115

Fig.4-24 Power-on Up-shift Simulation Results (With Motor Braking

Torque) .................................................................................... 116

Fig.5- 1 Schematic of Two speed DCT Power-train Rig ...................... 121

Fig.5- 2 Horiba-WT190 ...................................................................... 123

Fig.5- 3 Motor and Controller Used on Rig ........................................ 123

Fig.5- 4 Characteristics Matching of Motor and Dynamometer ........ 124

Fig.5- 5 Rig Inertia Flywheel Group ................................................... 126

Fig.5- 6 Cooling Pump (Left) and Characteristics Curve (Right) ......... 128

Fig.5- 7 Power-train Mounting Sub-Assembly 1 ................................ 129

Fig.5- 8 Power-train Mounting Sub-Assembly 2 ................................ 130

Fig.5- 9 Detailed sub assemblies for knuckle and wheel mounting ... 130

Fig.5- 10 Final Power-train and Rotating Inertia Assembly ................ 131

Fig.5- 11 Power Supply layout motor and controller are located after

the DC filter ................................................................................ 132

Fig.5-12 Power supply assembly, (left) Isolator and mains contactor,

(right) Inductor, capacitors and DCS550 4Q drive ........................ 132

Fig.5- 13Power-train Rig at University of Technology, Sydney ........... 133

Fig.5- 14 “V” Development Mode Based on Rapid Control Prototyping

................................................................................................... 137

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Fig.5- 15 DSPACE MicroAutoBox Left /RapidPro Right ..... 138

Fig.5- 16 DCT Control System base on MicroAutoBox ....................... 138

Fig.5- 17 Electrical Schematics of DCT Testing Rig ............................. 139

Fig.5- 18 Signals Definition of Rig Control System ............................. 141

Fig.5- 19 DCT Control Program ......................................................... 143

Fig.5- 20 Vehicle Monitor .................................................................. 144

Fig.5- 21 Motor Monitor ................................................................... 145

Fig.5- 22 DCT Shift Monitor and Calibration ...................................... 145

Fig.5- 23 Motor Control Software ..................................................... 146

Fig.5- 24 Power-On Up-Shifting 1000r/min30Nm ...................... 146

Fig.5- 25 Power-On Up-Shifting 3000r/min25Nm ...................... 147

Fig.5- 26 Power-Off Up-Shifting (500r/min ................................... 147

Fig.5- 27 Power-Off Up-Shifting 4000r/min ................................ 148

Fig.5- 28 Power-On Down-Shifting 500r/min25Nm .................... 149

Fig.5- 29 Power-On Down-Shifting 3000r/min25Nm .................. 149

Fig.5- 30 Power-Off Down-Shifting(500r/min) .................................. 150

Fig.5- 31 Power-Off Down-Shifting(3000r/min) ................................ 150

Fig.5- 32 DCT Temperature Testing Results 1st gear ................... 151

Fig.5- 33 DCT Temperature Testing Results 2nd gear .................. 152

Fig.5- 34 NEDC Driving Cycle ............................................................. 153

Fig.5- 35 UDDS Driving Cycle ............................................................. 154

Fig.5- 36 Motor Working Points NEDC/UDDS ............................. 156

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Fig.5- 37 Efficiency MAP of Two Speed DCT Power-train includes

Motor and Controller 1st gear/2nd gear ............................. 157

Fig.5- 38 Efficiency MAP of Single Reducer Power-train (includes Motor

and Controller) ........................................................................... 158

Fig.6- 1 Q60FB Prototype Car ............................................................ 160

Fig.6- 2 Q60FB Compartment Left and DCT sample Right .... 160

Fig.6- 3 Vehicle Layout Scheme ......................................................... 161

Fig.6- 4 Compartment Layout ........................................................... 162

Fig.6- 5 Layout of Batteries ............................................................... 162

Fig.6- 6 Layout of Charger Port ......................................................... 163

Fig.6- 7 Installation of the Real Car ................................................... 163

Fig.6- 8 On Road Testing and Calibration ........................................... 164

Fig.6- 9 Power-On Up-Shift Results Motor Status 100-Drive 200-

Brake .................................................................................. 165

Fig.6- 10 Power-Off Up-Shift Results .............................................. 165

Fig.7- 1 Two Speed Electric Power-train used Mechanical Pump ... 174

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LIST OF TABLES Table.2- 1 Q60FB-DCT/C70EV-DCT Parameters ................................ 45

Table.2- 2 Q60EV-DCT Performance ................................................. 45

Table.2- 3 Motor Efficiency Analysis ................................................. 48

Table.2- 4 Analysis of Motor Working Points Adjustment ................ 49

Table.2- 5 Q60EV-DCT Improved Dynamic Performances ................ 49

Table.2- 6 Q60EV-DCT Improved Economic Performance ................ 50

Table.2- 7 C70EV-DCT Performance Parameters .............................. 50

Table.2- 8 Two Motor Two Speed Matching Parameters .................. 64

Table.2- 9 Vehicle and Battery Parameters ....................................... 65

Table.2- 10 Performance Comparison .............................................. 65

Table.3- 1 Paramters of C70GB ........................................................ 77

Table.3- 2 Economic Performance.................................................... 81

Table.4- 1 Shifting Classification in different situations .................... 87

Table.4- 2 Shift Process Classification .............................................. 88

Table.5- 1 Vehicle Driving Resistance Analysis................................ 122

Table.5- 2 Vehicle Rotating Inertia ................................................. 125

Table.5- 3 the Existing Rig Inertia ................................................... 125

Table.5- 4 Economic Performance.................................................. 155

Table.5- 5 Compare of Efficiency Area ........................................... 158

Table.6- 1 Dynamic Performance Results (Q60EV-DCT) ................. 166

15

Table.6- 2 Economic Performance Results (Q60EV-DCT) ................ 166

16

GLOSSARY OF TERMS AND NOTATION

ABBREVIATIONS USED IN THIS THESIS

PEV -- Pure Electric Vehicle

HEV -- Hybrid Electric Vehicle

ICE -- Internal Combustion Engine

MT-- Manual Transmission

AT --Automatic Transmission

CVT --Continuously Variable Transmission

DCT-- Dual Clutch Transmission

AMT --Automated Manual Transmission

EVT --Electrically Variable Transmission

PDK -- Porsche Doppelkupplungsgetriebe (English: dual-clutch gearbox)

NEDC -- New European Driving Cycle

UDDS -- Urban Dynamometer Driving Schedule

ECE -- Economic Commission for Europe

EUDC-- Extra Urban Driving Cycle

FTP-72 -- Federal Test Procedure 72

FF -- Front mount Front drive

SOC – State of Charge

VCU --Vehicle Control Unit

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MCU – Motor Control Unit

TCU—Transmission Control Unit

NOTATION

Chapter three

MT -- Drive torque of motor;

gi -- Gear ratio of transmission;

0i -- Gear ratio of final drive;

T -- Efficiency of the whole driveline from the motor to the driven wheel;

r --Radius of the driven wheels;

G --Weight of vehicle;

f --Rolling resistance coefficient;

-- Road angle;

DC -- Aerodynamic drag coefficient;

A --Vehicle front area;

-- Rotational inertia factor;

'V --Speed of downshift point;

V --Speed of up-shift point;

nA -- the offset coefficient,

Chapter Four

-- Vehicle speed;

18

δ -- Vehicle gyrating mass conversion factor;

-- Vehicle mass;

-- Gear ratio;

-- Final ratio;

η -- Transmission efficiency;

-- Wheel rolling radius;

-- Clutch friction torque;

-Sliding friction loss power;

ω – Speed;

m – Motor;

1 –Clutch 1;

2-- Clutch 2;

t1 –Clutch 1 engage or disengage time;

t2 –Clutch 2 engage or disengage time;

θ—Rotational displacement;

I – Inertia element;

C – Damping coefficient;

K –Stiffness coefficient;

T— Torque;

n -- Number of friction plates;

X – piston displacement;

X0 – Minimum displacement required for contact between friction plates;

19

μ — dynamic friction;

μ –Static friction;

–Outside diameters of the clutch plates;

– Inside diameters of the clutch plates;

— Pressure load on the clutch;

— Average torque;

J— Conversion to the moment of inertia at the wheel 2mkg

ω—Flywheel angular velocity rad/s

M—Vehicle Mass kg

v—Vehicle Speed km/h

—Flywheel radius m

SOC0—Initial SOC value;

MAXCAP -- Battery capacity;

outV --Battery output voltage;

V-- Real voltage value;

I-- Real current value;

20

ABSTRACT

The research for this thesis is based on an international cooperation

project with BAIC Motor Electric Vehicle Co.Ltd, UTS and AVL/NTC. It

aims to develop a sample of a two speed DCT used in an electric drive

system.

For the dual clutch’s structural characteristics, one clutch is connected

with one gear, so it is very simple to realize two speed driving. Simulation

models are built in a co-simulation platform using AMESim and Simulink.

Gear ratio selection is processed during the matching of Q60FB and

C70GB vehicles. The ratios selected are 2nd and 3rd gear, and the ratios are

8.45 and 5.36. The prototype is modified from a VW 6spd DCT to operate

at 2 speeds. The work primarily involves modification of the mechanical

part of the gears and shaft, and changing the hydraulic parts.

To optimize vehicle dynamics and economic performance, a shifting

schedule calculation method for PEVs is provided. This uses a graphical

development method and is adapted for the purposes of simulations and

experimental work. As long as gear shifts are initiated according to the

schedule, the EM will be maintained at a higher efficiency operating region.

As a result, the proposed method provides more efficient operations of the

PEV.

Study of the control algorithm, including the vehicle control algorithm

and shift control algorithm, is the core of this thesis. To investigate shift

control and its calibration of a two speed DCT electric drive power-train,

this thesis analyzes the shifting process. The vehicle control

algorithm section follows the judgment of the pure electric multi-

mode algorithm. The shift control section analyses the traditional DCT

control shifting algorithm. In addition, the shifting control algorithm is

21

based on motor active braking control. Detailed shifting control algorithms

are developed which include power-on and power-off

methods. Corresponding simulation analysis has also been carried out.

The rig test uses the UTS power-train test bench for the purposes of

modification. Calibration and testing works are employed for processing

and the test rig mainly calibrates the shift control algorithm,

DCT temperature testing, and NEDC and UDDS drive cycles testing.

Vehicle integration and testing are finished at BJEV. This is based on

the BAIC independent brand car of Q60FB, with two gear DCT prototype

mounting and road test calibration. Finally, the project tests dynamic and

economic performance.

22

Chapter 1 Introduction

1.1 Background

THE number of cars on the road has increased from five million after

the Second World War to nearly one billion today and this is expected to

reach two billion in the next 20 years [1]. In recent years, developing

countries such as China and India, have contributed significantly to the

rapid growth of vehicle production. The biggest problem in relation to the

growing number of vehicles is the pressure placed on limited oil resources

and the corresponding environmental pollution. It is believed that the motor

vehicle sector is currently responsible for nearly 60% of the total world oil

demand and it will be the strongest growing energy demand sector in the

future. Between 2006 and 2030, around three quarters of the projected

increase in oil demand is expected to come from transportation [2][3]. In

the USA, light duty passenger vehicles consumed the highest amount of

energy (57%) in 2007. As a result, it is expected that about 4.1 billion

metric tons of carbon dioxide will be released into the atmosphere from

2007 to 2020. In the transportation sector, carbon dioxide (CO2) represents

the major green house gas emission and it accounts for 95% of the gas

produced [4].

Owing to serious energy shortages and environmental pollution, the

pure electric vehicle PEV and hybrid electric vehicle (HEV) have

been identified as alternatives to conventional passenger vehicles. This is

because of their low energy consumption and the fact that they produce

zero or low emissions on the road. Many countries, and almost every

automobile company, are conducting research and development into

23

electric power train systems. The drive to produce PEV’s is considered to

be a growing trend in the automobile industry.

Electric vehicles first came into existence in the mid-19th century

when electricity was one of the preferred methods for motor vehicle

propulsion, providing a level of comfort and ease of operation that could

not be achieved by the gasoline cars of the time. These vehicles differ from

fossil fuel-powered vehicles in that the electricity they consume can be

generated from a wide range of sources, including fossil fuels, nuclear

power, and renewable sources such as tidal power, solar power, and wind

power or any combination of these. The electricity may be stored on

vehicles using a battery, flywheel, or supercapacitor. Vehicles making use

of engines working on the principle of combustion can usually only derive

their energy from a single source or a few sources, and these are usually

non-renewable fossil fuels. Another advantage of electric or hybrid electric

vehicles is regenerative braking (i.e. their ability to recover energy

normally lost during braking as electricity is restored to the on-board

battery [5]).

Although the PEV appeared very early even before the internal

combustion engine vehicle, it never gained widespread use in the

automotive industry and the power-train structure is not as developed as the

internal combustion engine (ICE) power-train. This is as a consequence of

the limited battery capacity, limiting the running range. It is hard to run a

long distance, so most of the PEVs have been confined to operating as city

shuttles or taxis. The main purpose is to drive to the place of work and back

home, but not out of city. Thus the main driving condition of the electric

power train system considered during design is city driving cycle. Also

dynamic performance is weaker compared to internal combustion engine

vehicles.

The single reducer drive-train is the most popular structure of the PEV

24

as it uses the wide speed range of the motor to realize all drive speeds of

the vehicle. Take the Nissan LEAF, for example, it is coupled with a single

reducer, a high output torque in the low speed region and a high rotating

speed of the motor and this allows it to achieve the grade-ability and

maximum vehicle speed.

In the development of electric vehicles with the use of single reducers

it is hard to satisfy the requirement for the dynamic performance of PEVs,

especially in some luxury vehicles. This is because the motor cannot be too

big. In the limited maximum output torque and power, it is still difficult to

realize the same dynamic performance of the engine given the same power

level. Meanwhile, as there is no transmission to alter the speed ratio; it is

difficult to optimize the working point of the motor to high efficiency areas.

This restricts the performance of the PEVs, especially as the running range

is limited. The economic performance of the vehicle must be considered in

order to make it run longer with restricted energy resources. More and

more, research and application studies are beginning to pay attention to

multi-gear transmission application in PEVs, and this will become the

development trend of the PEV power-train system [6].

As we all know, multispeed transmissions are commonly used in

vehicles where the transmission adapts the output of the internal

combustion engine to the drive wheels. Such engines need to operate at a

relatively high rotational speed which is inappropriate for starting, stopping,

and slower travel. The transmission reduces the higher engine speed to the

slower wheel speed, increasing torque in the process. The original type of

transmission is Manual Transmission (MT), which needs a driver to operate

it. To reduce vehicle fuel consumption and tailpipe emissions, automotive

manufacturers have been developing new technologies for power-train

systems. Emerging technologies such as automatic transmission (AT),

continuously variable transmission (CVT), dual clutch transmission (DCT),

25

automated manual transmission (AMT), and electrically variable

transmission (EVT) have appeared in the market. The basic function of any

type of automotive transmissions is to transfer the engine torque to the

vehicle with the desired ratio smoothly and efficiently. [7][8]

The application of multispeed transmissions for PEVs has the

potential to improve average motor efficiency and enhance range, or even

reduce the required motor size [12]. There is a range of transmissions

available for application to PEVs for multispeed drives. In [13] Rudolph, et

al. suggested that DCTs have higher efficiency than other automatic drives,

making them particularly suitable.

DCT was invented by Frenchman Adolphe Kégresse just

before World War II but he never developed a working model. The first

development of the dual-clutch transmission started in the early part of

1980 under the guidance of Harry Webster at Automotive

Products (AP), Leamington Spa with prototypes built into the Ford

Fiesta Mk1, Ford Ranger & Peugeot 205. Initially, the control systems

were based on purely analogue/discrete digital circuitry with patents filed

in July 1981.[17] All of these early AP Dual Clutch installations featured a

single dry clutch & multi-plate wet clutch. After that, DCT work continued

from Porsche in-house development, for Audi and Porsche racing cars later

in the 1980s[15], when computers which were used to control the

transmission became compact enough: the Porsche

Doppelkupplungsgetriebe (English: dual-clutch gearbox) (PDK) used in

the Porsche 956 and 962 Le Mans race cars from 1983,[15] and the Audi

Sport Quattro S1 rally car.[18][19]

A dual-clutch transmission eliminates the torque converter as used in

conventional epicyclic-geared automatic transmissions.[14] Instead, dual-

clutch transmissions that are currently on the market primarily use two oil-

26

bathed wet multi-plate clutches, similar to the clutches used in most

motorcycles, though dry clutch versions are also available.[20]

The first series production road car to be fitted with a DCT was the

2003 Volkswagen Golf Mk4 R32.[15][21][22]

As of 2009, the largest number of sales of DCTs in Western Europe

were by various marques of the German Volkswagen Group,[23] though

this was anticipated to lessen as other transmission makers and vehicle

manufacturers began to make DCTs available in series production

automobiles.[15][24] In 2010, on BMW Canada's website for the 3 series

coupé, it is described both as a 7-speed double-clutch transmission and as a

7-speed automatic transmission. It is actually a dual-clutch semi-

automatic.[25][26]

In DCTs where the two clutches are arranged concentrically, as Fig.1-

1,the larger outer clutch drives the odd numbered gears, while the smaller

inner clutch drives the even numbered gears.[14][15][21] Shifts can be

accomplished without interrupting torque distribution to the driven road

wheels [14][15][16][18][21] by applying the engine's torque to one clutch

at the same time as it is being disconnected from the other

clutch.[15][18] Since alternate gear ratios can pre-select[14][15][16][18] an

odd gear on one gear shaft while the vehicle is being driven in an even

gear,[16] (and vice versa), DCTs are able to shift more quickly than cars

equipped with single-clutch automated-manual transmissions (AMTs), a.k.a.

single-clutch semi-automatics. Also, with a DCT, shifts can be made more

smoothly than with a single-clutch AMT, making a DCT more suitable for

conventional road cars.[27]

27

Fig.1- 1 Dual Clutch Transmission Structure

DCT has recently become the most popular type of transmission; it

seeks to combine the advantages of the conventional manual shift with the

qualities of a modern automatic transmission by providing different

clutches for odd and even speed selector gears. When changing gears, the

engine torque is transferred from one gear to the other continuously,

thereby providing gentle, smooth gear changes without losing power or

jerking the vehicle. DCT vehicles feature the convenience and comfort of

AT vehicles along with fuel economy. In such a way, they are even better

than MT vehicles. In addition, dual clutch transmission is less costly to

manufacture in comparison to automatic transmission since it shares similar

structures and components with the MT. Due to these advantages, DCT has

attracted extensive development interests in the automotive industry in

recent years.

Ricardo determined that the DCT and AMT are suitable for the

Chinese market for three reasons:

1. They can use the Chinese manual transmission manufacture

foundation;

28

2. They can get better fuel economy performance and shifting comfort

as compared to other types of transmissions;

3. They both have mature technologies and productions.

Fig.1-2 represents the efficiency analysis and a comparison between

different transmissions. Clearly the highest efficiency belongs to DCTs,

especially the dry clutch DCT.

Fig.1- 2 Transmission Efficiency under NEDC Drive Cycle

Meanwhile DCT is the type which can be easily designed with new

energy vehicles, especially the hybrid vehicle. For example, DCT can

satisfy start and stop functions without modification. Further, it is the only

type that can realize the pure electric drive in the P1 hybrid. Indeed, it can

start the engine without the starter. Some types of hybrid are in Fig.1-3.

Transmission Efficiency (%)

Engine Efficiency(%)

Total Efficiency(%

DCT is the best in total Efficiency

DCT can get the same shifting comfort of AT

29

Fig.1- 3 Hybrid Power-train Designs based on DCT

The advantages of the hybrid using DCT are as follows:

a) One of the difficulties of the hybrid power-train industry is its

complexity and the question of how to reduce cost of the drive-

train. Because DCT canceled the torque converter, meanwhile its

compact volume, the weight became lighter and this, in turn made

the assembly requirements of the DCT hybrid vehicle far smaller

than other vehicles;

b) The cancellation of engine idling is one of the major energy-

saving methods of hybrid cars. Owing to the fact that the DCT’s

engine and transmission can be completely separated in parking,

this completely cancels the engine idling which further improves

vehicle fuel economy;

c) By means of DCT's structural features, shifting can be finished

without power interruption and motor efficiency can be improved.

In addition, the drive motor in the vehicle arrangement is more

flexible;

d) When the hybrid vehicle design and parameters match, there is

sometimes a sacrifice in terms of shifting dynamics and this is in

30

exchange for vehicle fuel economy. The DCT notably has the

advantage of power shifting to make up for this defect.

In China some universities and companies are also researching the

power-train with DCT, such as JiLin University which has developed a

pure electric vehicle using dry DCT (Fig.1-4 Left). BYD “Qin” is a hybrid

with a single motor (Fig.1-4 Right). Its hybrid vehicle with dual motors is

also under testing.

Fig.1- 4 Prototype vehicle of Jilin University (Left) and BYD (Right)

1.2 Project Statement

The research of this paper is supported by BAIC Motor Electric

Vehicle Co. Ltd, the Ministry of Science and Technology, China, and the

University of Technology, Sydney.[2011DFB70060] The research project is

an international cooperation with BAIC Motor Electric Vehicle Co. Ltd,

UTS and AVL/NTC.

This project started in January2012 and finished in July 2014.

The author of this thesis is the project manager and core technology

staff member in terms of the development. The author has undertaken

control system development control strategy design and simulation, part

of prototype DCT manufacture testing rig design and building, vehicle

integration and on road testing.

31

1.3 Project Objectives

The aim of this project is the development of two speed DCT which

will be used in pure electric vehicles. In addition, the project is concerned

with rig testing and vehicle testing. The pre period has been processed in

Australia, and this has involved theoretical study, prototype development

and rig testing. A later period involved two prototypes being shipped to

China, while vehicle integration, calibration and on road testing were

processed in BAIC Motor Electric Vehicle Co. Ltd.

1.4 Project Scope

This thesis mainly studies two speed DCT pure electric driving system

structure and control methods. In addition, it covers the rig test,

vehicle installation and calibration.

Two speed DCT design needs gear ratio selection calculation and

analysis. Accordingly, a vehicle simulation model has been established, and

dynamic and economic performances have been simulated. The impact

of different gear ratio selections on vehicle performance has also been

analyzed.

The two speed DCT prototype has been restructured from the mature

DQ250 wet dual clutch transmission. This prototype has been finished

mainly in Australia by NTC Company. Restructuring work has mainly

involved the removal of excess gear group structures, and fixing the

toretain part of the gear necessary structure. The hydraulic system has been

reformed, and the electric hydraulic pump has been used to replace

the engine driven hydraulic pump. This has correspondingly led to the

closure and locking off of the hydraulic module circuit. When the motor

speed was higher than that of the engine, the cooling system was modified

32

in order to ensure an excellent temperature range. This internal forced

bearing cooling canceled the external cooling cycle in the original machine.

Research into the shifting schedule is the foundation of the power-

train dynamic and economic performance. Because the

motor characteristic is different to the conventional

engine, the utility to shift the schedule of the engine is no longer suitable

for the motor. However, there is currently no completely shifting schedule

method which is suitable for the motor. This thesis attempts to use

the rules of the economy shift design method based on motor efficiency.

Further, the dynamic shift design method is analyzed, and the

corresponding simulations are finished.

Study of the control algorithm, including the vehicle control algorithm

and the shift control algorithm, forms the core of this thesis. The vehicle

control algorithm section follows the judgment of the pure electric multi-

mode algorithm. The shift control section analyses the traditional DCT

control shifting algorithm along with the shifting control algorithm which

is based on motor active braking control. A corresponding simulation

analysis has been carried out.

The rig test uses the UTS power-train test bench for modification

purposes and calibration and testing works are employed. The test

rig mainly calibrates the shift control algorithm, DCT temperature testing,

and NEDC and UDDS drive cycles testing.

Vehicle integration and testing is finished at BJEV. This is based on

the BAIC independent brand car of Q60FB, with two gears DCT prototype

mounting and road test calibration. Finally the dynamic and economic

performances are tested.

33

1.5 Presentation of This Thesis

The chapter one is the introduction. This describes the overall trend

of development in terms of new energy vehicles and the pure electric

vehicle developmental direction. The biggest problem of the current

electric vehicle is its limited running range. A multi-speed gear box for use

in the electric drive system is proposed. This will improve the efficiency of

the motor and the dynamic and economic performance. Because DCT has

good shifting comfort and high efficiency, it is widely believed to be the

best choice for new energy vehicle development. Therefore, this thesis is

based on the DCT structure of the two speed pure electric drive system

research.

Chapter two examines the two speed DCT electric power-train

structure and its manufacture. Firstly, to introduce the two speed DCT

structure, the simulation platform is built using AMESim and Simulink.

After matching and analysis under this platform, the 2nd and 3rd gears are

selected as the target gear ratios of the two speed DCT. The sample DCT is

modified from the mature DQ250 wet clutch transmission, and this is

manufactured in the NTC Company. The modifications mainly concern

mechanical and hydraulic parts. Relative spin testing is then processed.

Finally a novel two motor, two speed power-train structure is introduced

which can further serve to improve motor efficiency.

Chapter three studies the multi-speed electric power-train shifting

schedule, and the dynamic and economic shifting schedule.

To optimize vehicle dynamics and economic performance, a shifting

schedule calculation method for PEVs is provided in this chapter which

uses a graphical development method, and this is adapted to be used in

simulations and experimental work. Using the acceleration curve of two

34

gears in the same throttle degree the intersection creates the ideal dynamic

shifting point. This is necessary for the downshifting line to have hysteresis

and thereby avoid shifting hunting. The economic shift schedule is

developed by taking a constant output torque across a number of vehicle

speeds. This determines the efficiency of the electric machine and generates

an efficiency curve. Where the two efficiency curves intersect is the point

of transition from higher efficiency in one gear to higher efficiency in the

other gear. Therefore, this is the optimum shift point to maximize the

operating efficiency of the PEV. As long as gear shifts are initiated

according to this schedule, the EM will be maintained at a higher efficiency

operating region and as a result the proposed method will maintain more

efficient PEV operations.

Chapter four represents research into the two speed DCT shifting

control strategy. The shifting control theory is analyzed. Generally shifting

is divided into a torque phase and an inertial phase. Torque transfer is

finished in the torque phase at the point of disengaging the clutch to

engaging the clutch, and the motor speed synchronization is in the inertial

phase. The three most common used quality criterions: shifting time, jerk

and sliding friction power are provided in this thesis. Two control methods

are provided. The first is the traditional dual clutch transmission shifting

control algorithm, which uses the overlap of torque transfer from one

clutch to another to overcome the torque hole of the shifting. Also, to

evaluate the control parameter’s influence on vehicle performance, a series

of slip values and minimum motor output torques are set in the power-on

up-shift control. After comparison of simulation results, correct control

parameters can be selected in the control system.

Another control method by motor active braking torque in the inertial

phase is researched in the final part of this chapter. Using motor braking

torque can shorten the shifting time significantly, but this can worsen the

35

shifting comfort. The on vehicle testing results are provided in Chapter Six.

Chapter five covers the rig testing. This part of the work is finished in

UTS using the rig of a power-train lab after it has been modified to adapt to

electric driving needs. The control system is developed using the

MicroAutobox of Dspace.

On the rig, shifting controls are calibrated and tested; also the driving

cycles of NEDC and UDDS are simulated on the rig. During the testing, a

good shifting control performance can be achieved, and a shifting control

algorithm can adapt to the driving cycle well. DCT temperature testing is

also processed, the maximum temperature of 1st and 2nd gears are 71.6

and 104.6 . This does not exceed the maximum limit of 140 .

Chapter six investigates the vehicle integration and road testing. The

two speed DCT is mounted on the Q60FB pure electric vehicle, and the

work is finished in BJEV. After installation, the road calibration of the

shifting control is undertaken and the dynamic and economic performance

testing is tested.

Chapter seven puts forward the thesis conclusions.

1.6 Publications

Journal [1] BoZhu,NongZhang,Paul.Walker,WenzhangZhan,XingxingZhou,JiagengRuan; Two

Speed DCT Electric Power-train Shifting Control and Rig Testing;Advances in

Mechanical Engineering;2013.11

[2] Bo.Zhu,N.Zhang,WenZhang.Zhan,P.D.Walker, YueyuanWei,

XingxingZhou,NanjiKe; Gear Shift Schedule Design for Multi-Speed Pure Electric

Vehicles; Journal of Automobile Engineering;2014.7 (SCI)

[3] Bo.Zhu Nong.Zhang Wenzhang.Zhan Yueyue.Wei Nanji.Ke Modelling and

Test Verification of DCT Hydraulic System; Automobile Technology; No.4,2014 (Serial

No.463)

[4] Bo.Zhu, Nong.Zhang, Wenzhang.Zhan, Yueyue.Wei, Nanji.Ke, Jiageng.Ruan; Study

36

of multi-gear pure electric vehicle economic shifting schedule and rig testing; Advanced

Technology of Electrical Engineering and Energy; Dec. 2013, Vlo.32, Suppl.

Conference [5] BoZhu, NongZhang, WenzhangZhan, YueyuanWei, NanjiKe, JiagenRuan; Study of

multi-gear pure electric vehicle economic shifting schedule and rig testing;

The 17th Conference of China EV; July 6-7, 2013, Beijing

[6] BoZhu, NongZhang, Paul.Walker, WenzhangZhan, XingxingZhou, JiagenRuan,

YueyuanWei; Two Speed DCT Electric Power-train System Rig Testing; NEV-DCS

2013,Beijing, Feb 6th 2013

[7] Bo.Zhu, NongZhang, WenzhangZhan, Paul.Walker, YueyuanWei, XingxingZhou,

NanjiKe; Two Motor Two Speed Power-train System Research of Pure Electric Vehicle;

SAE2013 Conference 2013.4

[8] BoZhu, Paul.Walker, NongZhang, WenzhangZhan, JiagenRuan, XingxingZhou,

YueyuanWei Study of Two Motor Multi-speed Electric Power-train System

TMC2013 SuZhou China March 2013 Excellent Paper Of the Conference

[9] BoZhu, WenzhangZhan, NongZhang, YueyuanWei, NanjiKe, XingxingZhou; Study

of Two Speed DCT Electric Power-train Up-shifting Control; Oct.2012; GuiLin,China;

LMS2012 Annual Conference.

Patent (1) A synchronous device; 201220196680.5

(2) A pure electric vehicle brake energy regeneration control system and method based

on DCT 201210188756.4

(3) A dual clutch hybrid system control unit, method and system 201210464909.3

(4) A Triple clutch transmission device and electric vehicle 201210143297.8

(5) A control system of pure electric vehicle brake energy regeneration based on DCT

201220270944.7

(6) A dual motor two speed electric drive system 201220266467.7

(7) Shift method using pure electric vehicle economy shift rule 201310201305.4

(8) Dual motor drive system and driving control method CN201210520830

(9) Multi mode of dual motor drive system and its driving method CN201210520862

(10) A novel hybrid system and its driving method CN201210520864

(11) Multi speed and multi mode hybrid system and its driving method

37

CN201210520863

(12) Multi mode hybrid system and its driving method CN201210520865

References

[1] International Council on Clean Transportation (ICCT). Passenger vehicles; 2010. [Online] Available at: http://www.theicct.org/passenger-vehicles/

[2] U.S. Energy Information Administration. International energy outlook 2010; 2010. [Online] Available at: http://www.eia.doe.gov/oiaf/ieo/pdf/ 0484%282010%29.pdf

[3] An F. Global overview on vehicle fuel economy and emission standards; 2010. [Online] Available at: http://www.un.org/esa/dsd/susdevtopics/sdtpdfs/meetings/egm0809/feng Global%20Overview%20UN%20NYC%20Aug- 09.pdf

[4] Atabani AE, et al. A review on global fuel economy standards, labels and technologies in the transportation sector. Renew Sustain Energy Rev (2011), doi:10.1016/j.rser.2011.07.092

[5] Electric Vehicle, From Wikipedia, http://en.wikipedia.org/wiki/Electric_vehicle

[6] Huang Juhua, Xu Shihua, Xie Shikun, “The Design of Automatic Transmission Control System of Electric Vehicle”, Journal of Jinggangshan University (Natural Science) Vol.32 No.1 Jan.2011

[7] Zongxuan Sun and Kumar Hebbale, Challenges and Opportunities in Automotive Transmission Control, 2005 American Control Conference June 8-10, 2005. Portland, OR, USA

[8] Xingyong Song PHD Thesis, DESIGN, MODELING, AND CONTROL OF AUTOMOTIVE POWER TRANSMISSION SYSTEMS; June, 2011; The University of Minnesota

[9] LiuYonggang, Datong Qin, HongJiang, YiZhang, “A Systematic Model for Dynamics and Control of Dual Clutch Transmissions” Journal of Mechanical Design,June 2009,Vol.131/061012

[10] Matthes,B., 2005, “Dual Clutch Transmissions—Lessons Learned and Future Potential,” Proceedings of the Transmission and Driveling Systems Symposium-4WD/AWD, SAE Paper No. 2005-01-1021.

[11] Wheals, J., Turner, A., Ramasy, K., O’Neil, A., 2007, “Double Clutch Transmission (DCT) Using Multiplexed Linear Actuation Technology and Dry Clutches for High Efficiency and Low Cost,” SAE Paper No. 2007-01-1096.

[12] P. D. Walker, S. Abdul Rahman, N. Zhang, W. Zhan, B. Zhu, and H. Du, MODELLING AND SIMULATION OF A TWO SPEED

38

ELECTRIC VEHICLE, International Conference on Sustainable Automotive Technologies 2012 to be held in Melbourne 21-23 March 2012.

[13] Rudolph F. Schafer M. Damm A. Metzner F T. and Steinberg I. (2007) The Innovative Seven Speed Dual Clutch Gearbox for Volkswagen’s Compact Cars 28th Internationales Wiener Motorensymposium

[14] "Powertrain — transmissions: Shift in power to the gearbox" (PDF). AMS (UnofficialBMW.com). September–October 2003. Retrieved 31 October 2009.

[15] "Automatic-shifting dual-clutch transmissions are poised to grab share from traditional transmissions thanks to their combination of efficiency and convenience" (PDF). AEI-online.org (DCTfacts.com). June 2009. Retrieved 31 October 2009.

[16] "Porsche Doppelkupplung (PDK)". Porsche.com. Retrieved 31 October 2009.

[17] "Patent GB2101243 - Control system for a vehicle automatic gearbox".Espacenet.com. Retrieved 2012-06-14.

[18] "Dual clutches take the lead". EurekaMagazine.co.uk. 13 March 2009. Retrieved 31 October 2009.

[19] "The Porsche Transmission". Lüfteknic.com. Retrieved 28 October 2009.

[20] "Wet Clutch or Dry Clutch?". DCTfacts.com. The Lubrizol Corporation. Retrieved 30 October 2009.

[21] "Volkswagen DSG — World's first dual-clutch gearbox in a production car". Volkswagen-Media-Services.com (Press release).Volkswagen AG. 22 November 2002. Retrieved 30 October 2009.

[22] "The 7-speed DSG — the intelligent automatic gearbox from Volkswagen". VolkswagenAG.com. Volkswagen AG. 21 January 2008. Retrieved 30 October 2009.

[23] "Volkswagen Group extends reach of dual clutch transmissions". DCTfacts.com. The Lubrizol Corporation. 8 May 2009. Retrieved 31 October 2009. "Some 311,000 light vehicles were produced in Western Europe with dual-clutch transmissions in 2008, according to data from JD Power; of these, the overwhelming majority were Volkswagen Groupmodels."[dead link]

[24] "Dual-Clutch technology voted Transmission of The Future at CTI Symposium". TheAutoChannel.com. Gordon Communications. 11 March 2009. Retrieved 9 November 2009.

[25] Interone Worldwide GmbH (2008-06-17). "BMW 3 Series Coupé — Highlights | BMW Canada". Bmw.ca. Retrieved 2010-12-14.

39

[26] Interone Worldwide GmbH (2008-06-17). "BMW 3 Series Coupé : 7-Speed Double Clutch Transmission". Bmw.ca. Retrieved 2010-12-14.

[27] "How dual-clutch transmissions work". HowStuffWorks.com. Retrieved 28 October 2009.

40

Chapter 2 Two Speed DCT Electric Power-train

Structure Analysis and Manufacture

2.1 Introduction of Two Speed DCT Structure

Fig.2- 1 Two Speed DCT Electric Power-train

Fig.2-1 presents the structure of a front wheel drive two speed dual

clutch transmission electric power-train. It is comprised of the motor,

coupled clutches, transmission gear train, differential and output to wheels.

The unique aspects of the DCT power-train are the application of clutches

and the arrangement of the gear train. The two clutches have a common

drum attached to the input shaft from the motor, and the friction plates are

independently connected to the first and second gears, respectively. With

only two gear pairs and a final drive gear in the two speed gearbox it is a

comparatively simple transmission, without the requirement to engage

alternate gears using synchronizer mechanisms [1,2]. For just two gear

ratios, gear shifting is realized through the dual clutch control alone.

41

Additionally, as the motor has the capability to reverse rotation, it can

reverse the vehicle, thus the reverse shaft is also eliminated. As a result,

the two speed dual clutch transmission equipped EV power-train is

relatively simple.

This project modified a VW 6spd DCT to operate in 2 speeds. The

type is DQ250 which is manufactured by Borgwarner, as show in Fig. 2-2.

Fig.2- 2 DQ250

The specifications of DQ250 are

Gear Six drive with one reverse

Clutch Multi disc wet dual clutch

Gear Ratio 14.26/8.45/5.36/3.71/2.86/2.39

Max Torque 350Nm

Weight 94kg

Oil 7.2L G052 182

2.2 Simulation Platform Building

Currently, electric vehicle simulation primarily adopts

42

MATLAB/Simulink modeling methods. It is modeled based on

mathematical equations, and provides high precision and control system

design capability. However, for the complex nonlinear system, there is a

conflict between model simplicity and precision. Also the modeling process

of battery and motor etc. creates a multi-subject system with high

complexity.

AMESim, the software by LMS Company is an alternative solution

for multi-subject field modeling and simulation. It provides a systems

engineering design platform. Users can build complex multi subject system

models in the platform and move to simulation rapidly. Meanwhile deep

analysis can be performed based on simulation results. It is suitable for

engineers to use. Its graphic modeling method can realize multi subject

combined platform building. And model extension and change can be

achieved through the graphical user interface which simplifies the

complicated modeling process. Engineers can be liberated from math

modeling and they can pay more attention to the physical system design.

More and more applications of AMESim have been found in the

automotive industry in recent years [3][4][5].

Although AMESim does not have great control system design ability,

it provides a communication interface with Simulink. A combined

simulation platform can be built in AMESim and Simulink to realize

efficient co-simulation. The co-simulation platform is proving to be one of

the most efficient developmental methods in the multi subject simulation

area.

During this project, we found that the commercial software is a quick

method to realize modeling and simulation in the initial period of matching

and analysis. It is not just quick, but it is also flexible in modeling different

structures and key parameters selection. Through this method, we can see

that companies are mainly using commercial software in matching, but

43

Universities are mainly using Matlab/Simulink to model, especially in

complex transient modeling and in structure dynamic analysis.

The model platform built in this chapter is in Fig.2-3. All matching

and simulation are finished in this platform. It includes the Vehicle model,

Drive model, Driving cycle, Battery model, Accessory model, Dual clutch

models, Reduce gears, and Vehicle Control Unit (VCU). The VCU model is

interfaced with Simulink, and detailed control algorithms are developed in

Simulink.

44

Fig.2- 3 Simulation platform based on Simulink-AMESim

45

2.3 Two Speed DCT Power-train Matching and Analysis

The matching is based on Q60EV which is the BAIC EV independent

brand developed from saab9-3. Meanwhile, to consider the power-train

using flexibility in different platforms of the C70EV, the matching of

C70EV is processed. Detailed vehicle parameters are in Table.2-1.

Table.2- 1 Q60FB-DCT/C70EV-DCT Parameters

Items Parameters

Vehicle Parameters

Vehicle Type Q60EV-DCT C70EV-DCT Drive Type FF FF

Length (mm) 4647 4860 Width (mm) 1762 1820 Height (mm) 1450 1461

Wheelbase (mm) 2675 2755 Front Track (mm) 1524 1522 Rear Track (mm) 1506 1528

Weight kg 1780 1925 Tire 225/45R17 235/45R17

Power-train Parameters

Peak Power kW 80 80 Rated Power kW 40 40 Peak Torque Nm 255 255 Rated Torque Nm 127 127 Rated Speed rpm 3000 3000 Max Speed rpm 9000 9000

Battery Parameters

Type Lithium ion Lithium ion Voltage V 372 372

Capacity Ah 60 60 Energy kWh 22.32 22.32

2.3.1 Q60EV-DCT Matching and Calculation

Preliminarily gear selections are 2nd and 3rd gear, ratios are 8.45 and

5.36, and performance simulations are as follows:

Table.2- 2 Q60EV-DCT Performance

Items Q60EV-Original Q60EV-DCT Improvement

Dynai Max Speed

km/h Long-term 140 146 4.28% Short Time 140 187 33.57%

46

0 100km/h Acceleration Time (s)

15 14 6.67%

0 50km/h Acceleration Time (s) 5.5 4.9 10.91% 50~80km/h Acceleration Time (s) 4.8 4.6 4.17% Grade Ability

(%) Long-term 12 14 16.67% Short Time 25 31 24%

4% Hill Speed(km/h) 140 147 5% 12%Hill Speed (km/h) 85 86 1.17%

Economic

Running Rangekm

NEDC 89.3 95.4 6.83% 60km/h 137.3 146 6.34%

Power Consumption per 100kmkWh/100km 60km/h

14.6 13.7 6.16%

As in Table.2-2, the two speed DCT can improve dynamic and

economic performance significantly as compared to the single reducer

power-train with the same motor power. The economic performance can

also increase to 6%.

Because the matching above is based on DQ250 and the ratio selection

are two closed gears, nothing optimization has been considered in the

process. From Fig.2-4 we can see that the motor working points are not

distributed in the high efficiency areas. The system efficiency still has

space to improve.

47

Fig.2- 4 Motor Working Points NEDC/UDDS

To improve the power-train efficiency, an analysis of the constant

speed points is done, as follows:

48

Table.2- 3 Motor Efficiency Analysis

Speed km/h Running Range

km

Motor Working PointsTorque/Speed

Motor Efficiency(%)

20 159.3 8.8Nm/1476rpm 71.3% 30 178.4 14.4Nm/1400rpm 80.5% 40 169.5 15.9Nm/1875rpm 79.8% 50 158.4 17.7Nm/2347rpm 81% 60 146.1 20Nm/2820rpm 83% 70 133.7 22.5Nm/3290rpm 85.2% 80 120 25.5Nm/3760rpm 85.9% 90 109.3 28.7Nm/4232rpm 88.2% 100 99.2 32.3Nm/4702rpm 90.2%

Motor working points under constant speed are as follows:

Fig.2- 5 Constant Working Points Efficiency Analysis

In Fig.2-5 we can see that from 20km/h to 100km/h the motor working

points in constant speed are all in the low efficiency area ( points in

figure). If the working points move to the near high efficiency area (▲

points in figure), the efficiency can be improved and the running range can

be extended. The analysis of working points adjustment is in Table.2-4, and

the 2nd gear ratio is mainly optimized for it influences the performance

49

significantly. From Fig.2-5, the ideal speed range can be estimated, as show

in Table.2-4. To achieve the ideal speed range, the gear ratio should be

adjusted, the adjusted value can be calculated, for example at the speed

30km/h, adjusted value is 1400/(500-1000)=(2.8-1.4). From all the speed

range, we can choice the adjusted value 2.5-1.69 to satisfy all speed

requests. So the ideal gear ratio can be 5.36/(2.5-1.69)=(2.144-3.17).

Table.2- 4 Analysis of Motor Working Points Adjustment

Speedkm/h

Original Working Points

Ideal Speed Range

Ratio Ratio Adjusted

Value

Ideal Ratio

20 8.8Nm/1476rpm 500-750 8.45 30 14.4Nm/1400rpm 500-1000

5.36

2.8-1.4 2.144- 3.17 40 15.9Nm/1875rpm 750-1500 2.5-1.25

50 17.7Nm/2347rpm 850-1500 2.76-1.56 60 20Nm/2820rpm 850-1800 3.31-1.57 70 22.5Nm/3290rpm 900-2000 3.66-1.65 80 25.5Nm/3760rpm 1300-2300 2.89-1.63 90 28.7Nm/4232rpm 1500-2500 2.82-1.69 100 32.3Nm/4702rpm 1800-2800 2.61-1.68 Based on analysis of the table above, 8.45/2.86 or 14.26/3.71(or 2.39)

can be selected as the optimized gear ratios. However, we should consider

that the ratio difference between the two gears can’t be too big as it will

cause shifting difficulty. Accordingly, 8.45/2.86 is the first selection, and

the dynamic and economic performances under optimized gear ratios are as

follows in Table.2-5 and Table.2-6:

Table.2- 5 Q60EV-DCT Improved Dynamic Performances

Items Q60EV-Original Q60EV-DCT Optimized gear selection

Max Speedkm/h

Long-term 140 146 149 Short Time 140 187 194

0 100km/h Acceleration Time (s) 15 14 19 0 50km/h Acceleration Time (s) 5.5 4.9 5 50~80km/h Acceleration Time (s) 4.8 4.6 8.3 Grade Ability (%) Long-term 12 14 14

Short Time 25 31 31 4% Hill Speed(km/h) 140 147 149 12%Hill Speed (km/h) 85 86 50

50

Table.2- 6 Q60EV-DCT Improved Economic Performance

Original Running Range km

Original Motor

Efficiency

Improved Running Range

km

Efficiency Improvement

Improvement Rate

NEDC 95.4 98.5 3.25% UDDS 93.2 96.0 3.00% 20km/h 159.3 71.3% 159.4 71.3% 0% 30km/h 178.4 80.5% 186.6 82.7% 2.2% 40km/h 169.5 79.8% 184 85.6% 5.8% 50km/h 158.4 81% 173.5 86.7% 5.7% 60km/h 146.1 83% 159.3 87.3% 4.3% 70km/h 133.7 85.2% 143.3 86.8% 1.6% 80km/h 120 85.9% 129.4 87.8% 1.9% 90km/h 109.3 88.2% 116.4 88.6% 0.4% 100km/h 99.2 90.2% 104.6 89.4% -0.8%

After optimization of the gear ratios, the grade-ability is not changed

when the max speed is improved but the acceleration time is increased. The

0-100km/h acceleration time is 19s, and 50-80km/h acceleration is 8.3s.

This is greater than the original single reducer power-train so it is

unacceptable.

The economic performance can be improved by 6% in constant speed

points from 40km/h to 60km/h. Meanwhile 3% can be improved in the

NEDC drive cycle. But in the low speed and high speed area, the economic

efficiency is not improved.

2.3.2 C70EV-DCT Matching and Calculation

C70EV is 150kg in weight greater than Q60EV when the same power-

train is selected, so the 2nd and 3rd gears are matched. The simulation results

are in Table.2-7.

Table.2- 7 C70EV-DCT Performance Parameters

Items Q60EV-Original

Q60EV-DCT C70EV-DCT

Dyna

mic

Max Speedkm/h

Long-term 140 146 146 Short Time 140 187 191

0 100km/h Acceleration Time (s) 15 14 15.2

51

0 50km/h Acceleration Time (s) 5.5 4.9 5.3 50~80km/h Acceleration Time (s) 4.8 4.6 5.1 Grade Ability

(%) Long-term 12 14 13 Short Time 25 31 28

4% Hill Speed(km/h) 140 147 144 12%Hill Speed (km/h) 85 86 82

Economic

Running Range

km

NEDC 89.3 95.4 88.2 60km/h 137.3 146 141.5

Power Consumption per 100kmkWh/100km 60km/h

14.6 13.7

From Table.2-7, the max speed of Q60EV-DCT and C70EV-DCT are

almost the same, but the acceleration performance and economic

performance of C70EV-DCT are worse than the Q60EV-DCT.

2.3.3 Matching Conclusion

The gear selections are 2nd and 3rd gears (8.45/5.36). Dynamic

and economic performance are improved compared to the single

reducer structure, and economic performance is improved by 6%;

If selections are 2nd and 5th gear (8.45/2.86), economic

performance can be improved by 3% to 6%, but acceleration time

increases. As a result, it is not an ideal selection;

This selection still has room to improve. Indeed, if we can totally re-

match the motor and gearbox, the vehicle performance can be considerably

improved.

2.4 Two Speed DCT Prototype Manufacture

2.4.1 Two Speed DCT Prototype Manufacture

This project modified a VW 6spd DCT to operate in 2 speeds. The

ratios selected are 2nd and 3rd gear, and the ratios are 8.45 and 5.36. The

works are mainly modifications of the mechanical parts of the gears and

52

shaft. Meanwhile the hydraulic parts are changed.

Fig.2- 6 Removed 3nd shaft (5th and 6th gear) and reverse shaft.

Fig.2- 7 Gear Modification

Remove the 1st and 4th gear with their synchronizers on the second

shaft. Mechanically lock in 2nd and 3rd gear with spacers. The hydraulic

gear change pistons are removed.

Fig.2- 8 Lubrication System

Forced lubrication to bearing on 2nd shaft. Some unused injectors are

closed off on the lubrication system to concentrate/improve lubrication on

the 2nd and 3rd gear.

53

Fig.2- 9 Original Hydraulic System

54

Fig.2- 10 Modified Hydraulic System

55

Fig.2- 11 Hydraulic Valve Body Modification

Valve Body modifications, and blocked positions in the plate are

employed to improve oil flow where needed. An orifice is placed to reduce

flow to the oil cooler.

Fig.2- 12 Original Oil Pump

Fig.2- 13 12 Volt E-motor drives pump

56

The original oil pump cannot be used because the electric motor will

run in the reverse direction. An electric driven oil pump is used. The

original pump is replaced by a block to bring the suction and pressure

connections outside the gearbox. The strategy is to use one pump for

lubrication, clutch cooling and clutch activation. Another 12 Volt E-motor

drives the pump for clutch pressure and lubrication (Fig. 2-13).

Fig.2- 14 Original Pressure Sensor

Fig.2- 15 Speed Sensor

Fig.2- 16 Oil temperature sensor

The existing pressure sensors in the transmission valve body have

been rewired and these are used to provide clutch pressure to the controller

(Fig. 2-14). Output speed and oil temperature sensors have been added to

provide the respective signals to the controller (Fig. 2-15 and Fig. 2-16).

57

Fig.2- 17 E-Motor side (long shaft) and Gearbox side (short shaft)

Modified driveshaft flanges (short and long) are employed as a

reduced drive joint diameter is required for clearance to the E-Motor.

2.4.2 Spin Testing

To test the modified DCT prototype, spin testing is finished in the

NTC’s spin rig, see Fig. 2-18. The aim of spin testing is to test that the

prototype can work and spin well in the unloaded condition. The testing

results are shown in Fig. 2-19 to Fig.2-23.

Fig.2- 18 Spin Testing Rig

58

Fig.2- 19 Clutch Pressure Sensor Testing

Fig.2- 20 Output Speed Sensor Testing

From Fig.2-19 and Fig.2-20, the pressure sensor and output speed

sensors can work well. Also the calibration data will be used to control the

program.

59

2 4 6 8 10 12 14Pressure(Bar)

Fig.2- 21 Hydraulic Testing (with mains powered pump)

60

Fig.2- 22 Main Solenoid Testing

From Fig.2-21 and Fig.2-22, the hydraulic system and solenoid valve

work well.

Fig.2- 23 Clutch Activation Solenoid Testing

61

Fig.2-23 is the clutch solenoid control testing. In the figures we can

see that the solenoid real current can follow the ideal current well,

meanwhile the pressure reacts very quickly to follow the requirements in

the ramp and pulse current control signals respectively. All this data will be

incorporated in the control program as the hydraulic calibration parameters.

2.5 Novel Two Speed Electric Power-train System

2.5.1 Two Motor Two Speed System Matching

How to improve drive-line efficiency? The conventional internal-

combustion engine (ICE) power-train uses the solution of multi-gear

transmissions. These transmissions can improve engine efficiency through

gear shifting, which has the effect of shifting the operating speed of the

engine to an efficient region, and it achieves an optimized driving

characteristic field to match the vehicle requested characteristic field [6].

As an alternative to using the gearbox to change engine speed and torque,

another efficient way is the application of multiple power sources, such as

the hybrid electric systems. The hybrid power-train introduces a motor into

engine drive system; it can split the conventional single driving power into

two power sources. The motor can assist driving or regenerate power, to

adjust engine working points to optimized areas and realize efficient engine

driving power. The full hybrid system can realize gasoline savings of

30~50% [7] [8].

Through the concept of split power hybrid vehicles, we can try to split

our PEV’s power source from one drive motor into two motors, with the

associated torque control of the two motors resulting in better overall

vehicle performance.

The structure of the two motor two speed system is shown in Fig.2-26.

62

Motor 1 and motor 2 output torque is shown in parallel and the output shaft

is like the dual clutch’s output shaft. The outer hollow shaft connects to the

motor 2 output shaft, and the inner solid shaft connects to motor 1 output

shaft. For performance optimization, two motors connect with different

gear ratios. Motor 1connects to the 1st gear and motor 2 connects to the 2nd

gear. There is no shifting clutch in the system; the gear shift is achieved

through motor control. So the structure is simple. There is a one-way clutch

between motor 1 and motor 2 to prevent motor 1 rotating when motor 2 is

the only one working. Also, for the different ratios of the two gears, the

one-way clutch can protect motor 1 to run over speed when motor 2 is in

high speed rotation. It can therefore protect the system and prolong its

useful life.

• Figure 2-26. Two motor two speed power-train structure

As in Fig.2-26, motor 2 is the more frequently used driving motor. The

2nd gear ratio is less than the 1st gear. It is configured to provide efficient

driving in common drive cycles. Motor 1 assists the driving motor, and it is

63

mainly focused on high torque output during high dynamic performance

requests. Thus, it can realize two speed driving and improve vehicle

dynamic performance. Through the output torque shifting between two

motors, motors can work in more efficient areas. As a consequence, system

efficiency can be improved and this extends the running range.

The conventional one motor single reducer pure electric power-train

system needs large backup power to satisfy dynamic performance. This is

because there is no multi-gear transmission available to optimize the motor

working point, and the average operating efficiency of the motor is low.

Fig.2-27 is the motor working point of one motor single reducer in the

NEDC driving cycle. In the figure, the blue line is the max motor output

torque, the red line represents efficiency contour lines and the green points

are motor working points. Obviously the motor’s working points cannot

spread in the high efficient area under these driving conditions.

• Fig. 2-27. Motor working points of one motor single reducer

64

• Fig.2-28. Motor working points of two motor two speed

Fig.2-28 displays the motor working points of the optimized two

motor two speed system. The red solid line is the motor 2 max torque line

and the purple dashed line is the motor 1 max torque line. Detailed power-

train parameters of two motor two speed system are shown in Table.2-8. All

the motors we have selected in the simulation are permanent magnet

motors because they have high efficiency over a wide operating range.

Motor 2 is the frequently used motor, the max power of which is

40kW. The 2nd gear ratio of 7 is selected, which includes the final gear ratio.

The Motor 2 parameter matching is according to the motor working point

under the common driving cycle. Motor 1 is dynamic assistant motor; the

max power is also 40kW. The 1st gear ratio of 10.5 is chosen. The Motor 1

parameter matching is achieved from the torque request in acceleration and

the climbing condition, thus the two motor designs are able to run at higher

efficiencies under normal conditions, and they can meet higher torque

demands as requested.

Table.2- 8 Two Motor Two Speed Matching Parameters

65

Items Parameters Motor 1 Motor 2

Peak Power(kW)kW 40 40 Rate Power(kW)kW 20 20

Peak Torque(Nm) Nm 191 100 Rate Torque(Nm) Nm 95.5 50 Basic Speed(rpm) rpm 2000 3750

Max Speed(rpm) 9000 9000 Gear Ratio(Include final gear) 10.5 7

2.5.2 Simulation Result and Analysis

To validate the two motor two speed system, simulations are

conducted in the above platform. Vehicle parameters are shown in Table.2-

9.We chose the vehicle platform of Q60FB from the BAIC Motor Electric

Vehicle Company, which is the front motor front drive (FF) pure electric

vehicle with 26.78kWh battery power.

Simulations include dynamic testing and economic testing, whereby

dynamic testing includes max speed, 0-100km/h acceleration time, and max

grade-ability, and economic testing is running in a range under the NEDC

driving cycle. Simulation results are summarized in Table.2-10.

Table.2- 9 Vehicle and Battery Parameters

Items Parameters Vehicle Type Q60FB Drive Type FF Length*width*height(mm) 4647*1762*1450 Wheelbase (mm) 2675 Track (mm) 1524/1506 Weight kg 1780 Tire 225/45R17

Battery Voltage V 372 Capacity Ah 72 Energy kWh 26.78

Table.2- 10 Performance Comparison

Items A B C

Dynamic Max Speed km/h 152 142 198 0-100km/h Acceleration Time (s) 12.2 13.3 13.1 Grade Ability (%) 42 28 33

Economic Running Range km 127.34 118.18 124.56

66

(NEDC) Range improve rate 7.75% 0% 5.4% Motor Average efficiency 85.29% 79.83% 82.73%

(*A- Two motor two speed; B- One motor single reducer; C-One motor two speed;) To compare performance with one motor single reducer and one motor

two speed structure, another two structures are matched and simulated as B

and C in Table.2-10. For the one motor two speed structure, the max power

of the motor is 80kW; gear ratios are 8.45 and 5.36. For the one motor

single reducer, max motor power is 80kW and the single gear ratio is 7.5.

From Table.2-10 we can see that the two motor two speed system can

improve dynamic and economic performance as compared to the one

motor single reducer system. Further, the max speed, acceleration time and

grade ability are all improved significantly. The running range in NEDC

cycle is improved, from 118km to 127km, this equates to an increase of

7.75%. The average efficiency statistic also shows a 6% improvement of

motor efficiency under these driving conditions. Also it is clear that while

the one motor two speed system can obviously improve the dynamic

performance, the two motor two speed system can achieve a better

economic performance than the one motor two speed system under the

same total motor power. So in relation to the running range and economic

behavior, the two motor two speed system can be considered to have more

competitive abilities.

If we consider the cost of the system, the single reducer is

significantly lower in terms of cost. The cost of the two motor two speed

system will increase primarily as a result of the use of two motors and their

controllers. However, if the mass production of such a platform is

considered, the cost will dramatically decrease and it may even be lower

than the one motor two speed system because other ancillary systems such

as clutches are not necessary.

67

2.6 Conclusions

Single reducer transmissions are the most popular form of PEVs. They

are is simple and low in terms of cost, but they do not always satisfy the

dynamic and economic performance requirements of such PEVs, especially

in regard to some luxury vehicles. Meanwhile, as there is no transmission

to alter the speed ratio it is difficult to optimize the working point of the

motor to high efficiency areas. This restricts the performance of the PEVs,

especially as the running range is limited.

The two speed DCT electric power-train system is introduced in this

chapter because of the dual clutch’s structural characteristics, i.e. one clutch

connects with one gear, so it is very simple to realize two speed driving.

Simulation models are built in the co-simulation platform using AMESim

and Simulink. Gear ratio selection is processed during matching of Q60FB

and C70GB vehicle. The ratios selected are 2nd and 3rd gear, the ratios are

8.45 and 5.36. The prototype is modified from VW 6spd DCT to operate in

2 speeds. Mainly the works are modifications of the mechanical parts of the

gears and shaft, meanwhile the hydraulic part changes.

Finally a new pure electric structure of the two motor two speed

power-train is provided in this chapter. Two motors are used instead of the

former single motor design. One of the two motors is the primary drive

motor which is matched accordingly to optimize motor working points.

Another motor is the dynamic assist motor, which is matched accordingly

to provide the large torque requirements. Thus the two motors have two

different gear ratios to achieve these different needs. Simulation results

show the system can not only improve dynamic performance, but also the

economic performance and prolonged running range. The structure is

simple compared to conventional multi-gear transmission and it provides a

similar degree of design flexibility.

68

References

[1] Walker, P.D., Zhang, N., Zhan, W.Z., Zhu, B. “modeling and simulation of gear synchronization and shifting in dual clutch transmission equipped powertrains”, Proceedings of the Institution of Mechanical Engineers, Part C: Journal of Mechanical Engineering Science, Feb; vol. 227, 2013. [2] Galvagno E, Velardocchia M and Vigliani A. Dynamic and kinematic model of a dual clutch transmission. Mechanism and Machine Theory vol. 46 no. 6, 2011.[11] M. Goetz, M. Levesley, D. Crolla, Dynamics and control of gearshifts on twin-clutch transmission, Proceedings of the institution of mechanical engineers, Part D: Journal of Automobile Engineering 219 (2005) 951–963. [3]XieFei, SongChuanxue, LiuMingshu, ZhangYoukun,LuYanhui; Research on Co-simulation Platform Based on AMESim and Simulink for Dual State CVT; Automotive Technology; Vol.8 2008 [4]Wang Pengyu,Wang Qingnian, Hu Anping,YuYuanbin;Analysis of regenerative brake system of hybrid bus based on Simulink-AMESim co-simulation; Journal of Jilin University(Engineering and Technology Edition) Vol.38 Sup. Feb.2008 [5]Liao Linqing,ZHANG Dongfang, Qu Xiang, Ke Jingjing; Simulation on the Start-up and Shift Process of Dual Clutch Transmission Vehicle Based on AMESim; Journal of Chongqing University of Technology (Natural Science), Vol.25 No.1,Jan.2011 [6]AnLin Ge, “Vehicle Auto Transmission Theory and Design”, China Machine Press, 1991.10 [7]Yimin Gao, Ali Emadi, Mehrdad Ehsani, “Modern Electric Hybrid Electric and Fuel Cell Vehicles Fundamentals Theory and Design”, CRC Press, 2010 [8]QingquanChen, FengchunSun Modern Electric Vehicle TechnologyBeijing Institute of Technology Press; 1 Jan 2002

69

Chapter 3 Multi-Speed Electric Power-train

Shifting Schedule

3.1 Introduction

Gearbox shift schedule design is a rule based strategy to determine the

appropriate driving gear according to driver input and vehicle driving

conditions. It influences the economic and dynamic performance of the

vehicle dramatically and is one of the key technologies of gearbox control

systems. Generally, there are three types of shift schedules: single

parameter, dual parameter and three parameter schedules. The single

parameter shift schedule is controlled by vehicle speed. It is simple but

does not take into account both the economic and dynamic performance

requirements of modern vehicles and so is rarely used nowadays. Dual

parameter shift schedules are based on speed and throttle input from the

driver. These are currently widely used and performance requirements are

generally met under most driving conditions. Lastly, three parameter

shifting is controlled according to speed, acceleration and throttle demand

[1]. This reacts to vehicle dynamics and can produce the best vehicle

performance. The downside is a complicated program requirement and

implementation in-online vehicle control [2], and it is consequently

excluded from this research.

Gear shifting theory for ICE vehicles is well developed. Ge [1]

describes the calculation method and evaluation indexes of gear shifting in

detail. Huang [3] provides a graphical method for the dynamic and

economic shift schedule design for ICE vehicles. However, there is little

70

research regarding shifting schedules for PEVs. Yang [4], through the

establishment of numerical models of the battery and motor systems for

electric vehicles, analyzed the optimal-power shift schedule. Owing to the

fact that the maximum discharge power of the battery decreases with the

decrease of SOC, three -parameter and four -parameter dynamic shift

schedules were developed. These schedules operate when the maximum

discharge power of battery is greater than or less than the maximum input

power of the motor.

It is obvious that we can use some of the basic methods and

calculation algorithms of the internal combustion engine shifting theory to

establish PEV shift schedules, but as the driving power source has changed,

the characteristics and efficiency of the traction motor will produce

significantly different results as compared to ICE powered vehicles. It is

therefore necessary to develop alternative shift schedules for PEVs which

provide the theory to support PEV power-train system matching and

optimization.

In this chapter, the gear shift schedule is developed for a power-train

platform equipped with a two speed DCT, however this method can be

extended to all kinds of multi-geared PEV power-train systems. The

shifting schedule includes dynamic and economic schedules. Through

simulations results it can be seen that the economic method, in particular,

can improve vehicle performance significantly.

3.2 Dynamic Shift Schedule Development for Multi-Speed

Pure Electric Vehicles

There are two general methods for the development of dynamic

shifting schedules for internal combustion engines: graphical or analytical.

The graphical method for determining the best dynamic shifting point

71

chooses the intersection points of the adjacent gear acceleration curves

under the same throttle conditions. It should be mentioned that the drive

torque curve is not used because the nonlinear running cycle causes

variation in acceleration. This makes it difficult to achieve the best

dynamic performance for this shift schedule with just the driving torque

curve. Although there are differences between the torque characteristic

curves of motors and engines, it is possible to apply the same method to

determine the dynamic shift schedule for a PEV. The second option is the

analytical method. Through application of computer programs the point of

the same acceleration value is calculated in two different gears, and shifting

points are determined [1].

The output torque of an electric motor is very regular. At low speed,

voltage supply to the motor increases proportionally to speed through the

electronic converter while flux is kept constant at the point where the base

speed voltage of the motor reaches the source voltage. Beyond the base

speed, the voltage of the motor is kept constant and the flux is weakened,

dropping hyperbolically with increasing speed. Consequently the peak

motor torque also drops hyperbolically with increasing speed [5]. As a

result of this regularity the graphical method is simpler than the analytical

method and it can obtain reliable results with relative ease.

Hereinafter is the detailed application of the graphical method for

single parameter schedules. According to vehicle driving equation, we can

calculate the vehicle acceleration as follows [6]:

)sin15.21

cos(1 20 GuACGfriiT

mdtdu

aDTgM (3-1)

Where, MT is drive torque of motor, gi is gear ratio of transmission, 0i is

gear ratio of final drive, T is the efficiency of the whole driveline from the

motor to the driven wheel, r is the radius of the driven wheels. G is the

weight of vehicle, f is the rolling resistance coefficient, is the road angle,

72

here we set it to zero. DC is the aerodynamic drag coefficient, A is the

vehicle front area, is the rotational inertia factor.

From the above equation (3-1), acceleration curves calculated at

different throttle percentages are developed. These are shown in Figure 3-1.

Fig.3- 1 Vehicle acceleration curves for establishing single parameter shift map

In Fig.3-1, there are six acceleration lines, i.e. three lines for 1st gear,

with different throttles of 100%, 70% and 30%, while the other three lines

are for 2nd gear at the same throttle input values. The intersection point for

the two gear ratio acceleration curves at the same throttle input are

displayed for the 100% throttle input at A, and 70% and 30% are B and C,

respectively. Linking these points, the dynamic up-shift line is obtained.

This resulting line for the motor is a straight line and this can be considered

an up-shift line while for an engine it is a curve. This is as a result of the

maximum torque curve for a motor being more regular than an engine.

Any shift schedule must avoid excessive repetition of gearshifts,

known as shift hunting. To avoid this repetition the down-shift line must be

sufficiently offset from the up-shift line. Therefore, a buffer region between

the up-shift and down-shift is used to avoid shift hunting. The zone

between the two shift lines creates shifting delay, and helps to avoid

73

circular shifting hunting phenomenon in situations such as hill climbing.

Reference [7] gives the down-shift schedule calculation equation as follows:

VAV n )1(' (3-2)

Where 'V is the speed of the downshift point while V is the speed of

the up-shift point in the same throttle percentage input, nA is the offset

coefficient, it can be generally set to 0.4~0.45, we make it nA = 0.40 in this

paper, and the down-shift line for the shift schedule is as follows in Fig.3-2,

where the solid line is the up-shift line and the dash-dotted line is the

down-shift line.

Fig.3- 2 Dynamic upshift and downshift map for PEV

3.3 Economic Shift Schedule development for Multi-Speed

Pure Electric Vehicles

As the running range is limited for PEVs, economical performance is

probably the most important factor considered in PEV power-train system

design and control. Much like an ICE driven vehicle, multiple gears are

used to improve the operating region of the motor. The shift schedule

significantly influences this operating region and therefore the driving

efficiency. As a consequence, the development of an appropriate shift

schedule that considers vehicle economy is very important to multi-speed

EV development as it directly influences the vehicle running distance.

The economic shift schedule design for engine driven vehicles is

74

derived mainly from the fuel consumption rate MAP [1]. Through

calculation of the fuel consumption curve in different throttle degrees at the

same drag torque, the intersection point of two curves in the neighboring

gear ratio is the shifting point for a given vehicle speed and input throttle.

The power use of an electric motor is different from that of an engine,

and the drive characteristics also differ significantly. The economic

performance of the motor mainly depends on the efficiency of its operating

region. If motor operation can be maintained in the high efficiency regions,

the economic performance of the system can be improved. The method to

determine economical shift schedule for a motor is as follows:

1) Plot motor characteristic curves and efficiency MAP of the

consecutive gear ratios in the same figure, as shown in Figure 3-3.

The x-axis is vehicle speed, and y-axis is output torque at the

transmission output shaft;

2) Draw a constant traction torque line of T0 in the overlapping region

of Fig.3-3, as in the overlap area where there are two driving gear

ratios available;

3) Calculate the motor efficiency along output torque line T0 for

different vehicle speeds in both gear ratios, and plot the two lines in

the same figure, see Fig.3- 4. The solid line is the efficiency line for

1st gear, and the dash-dotted line is the efficiency line for 2nd gear.

The intersection point of these lines is denoted as A. It should be

understood that before point A the efficiency of the 1st gear is higher

than the 2nd gear, and after this point the efficiency of the 2nd gear

is higher than the 1st gear. Obviously, before point A the gearbox

should be operating in 1st gear while after that it should be in 2nd

gear. So the point A must be the shifting point at the given operating

output torque and vehicle speed;

4) However, the point A for 1st and 2nd gear at the same vehicle speed

75

will have two different throttle values, depending on the gear

selected. These are shown in Fig.3-5 as point a and a’. These can be

considered the up-shifting point and down-shifting point for a given

speed;

5) By repeating the same procedure above for different output torque

values, all the up-shifting and down-shifting points at these torques

and resulting speeds are then linked to produce the up-shifting and

down-shifting lines, as in Fig.3-5. The solid line is the up-shift line

from 1st gear to 2nd gear, and the dash line is the down-shift line

from 2nd gear back to 1st gear.

constant traction torque line T0

efficiency MAP of 1st gear

efficiency MAP of 2nd gear

Fig.3- 3 Efficiency MAP of motor in 1st and 2nd gear relative to vehicle speed

76

Fig.3- 4 Economic shifting points for output Torque T0

Fig.3- 5 Economic shifting schedule curve for PEV

In Fig.3-5 it is shown that the up-shift and down-shift lines are too

close and this may cause frequent shift operations under certain driving

conditions. This problem is eliminated by adjusting the down-shift

schedule [3] using equation (3-2) and setting An = 0.4. After modification,

the shift schedules are as shown in Fig.3-6, where the solid line is the up-

shift line and the dash line is the down-shift line.

77

Fig.3- 6 Adjusted Economic shifting schedule curves for PEVs

3.4 Simulations and Analysis

For the simulations presented in this chapter, the vehicle platform is

based on a mid-class saloon car C70GB, which is an independent brand of

the Beijing Automobile Group China. The details of the parameters used in

this simulation are set out in Table.3-1. The simulation was conducted on

the co-simulation platform built in Chapter two, in which the vehicle and

power-train were built in AMESim, and the control models were in

Simulink.

Table.3- 1 Paramters of C70GB

Items Parameters Vehicle Type C70GB

Basic Parameters Length (mm) 4860 Width (mm) 1820 Height(mm) 1461

Wheelbase(mm) 2755 Front Track(mm) 1522 Rear Track (mm) 1528

Body Mass kg 1780 Passenger Number 5

Tire Type 235/45R17 Power-train System Max Power kW 80

Rate Power kW 40 Max Torque Nm 255 Rate Torque Nm 127 Base Speed rpm 3000 Max Speed rpm 9000

Gear Ratio(Include final drive) 8.45 / 5.36

78

Voltage V 372 Capacity Ah 66 Energy kWh 24.55

Dynamic performance is generally evaluated by acceleration time,

grade-ability and maximum speed. As the grade-ability and maximum

speed have been confirmed in power-train matching [8], the only

performance which relates to shifting is acceleration time. In this chapter to

validate the dynamic performance, the acceleration test from 0 to 100km/h

is selected.

For comparison of the dynamic performance in different shift

schedules, three acceleration tests were performed using the different

shifting processes. These are the dynamic shifting and economic shifting

schedule developed above, and another general shifting schedule which

was used in simulation before the shift schedule was optimized. The

simulation results are displayed in Fig.3-7. The results presented in Fig.3-7,

show the solid line as the acceleration time in the dynamic shifting

schedule, and the dashed line and dash-dotted line are the results with the

economic shifting and general shifting schedule, respectively. Results

demonstrate the dynamic shifting schedule achieves better acceleration

performance in the 0-100km/h acceleration of 13.3s. While the economic

shifting and general shifting are 13.45s and 13.78s respectively.

79

General Shifting

Fig.3- 7 Acceleration Performance of PEV for different shift schedules

To validate the economic performance, driving cycles are used to

judge the energy consumption and running distance for a given range of the

battery SOC. In this chapter selected cycles are NEDC and UDDS. The

NEDC (New European Driving Cycle) is the regulated European cycle for

defining the specific fuel consumption and emissions of passenger cars.

The entire cycle includes four ECE segments, followed by one EUDC

segment, shown in Fig.3-8. Its average speed is 33.6 km/h, the maximum

speed is 120 km/h, and the total distance is 11 km. UDDS (Urban

Dynamometer Driving Schedule) is also known as the U.S. FTP-72

(Federal Test Procedure) or the LA-4 cycle. It is the simulation of an urban

driving route approximately 12.1 km (7.4 miles) long and takes 1,369

seconds (approximately 23 minutes) to complete, shown in Fig.3-9. In

Fig.3-8 and 3-9 there are two lines, the solid line is the target speed of the

running cycle, while the dashed line is the actual running speed in the

simulation, and we can see that they are essentially identical, indicating the

powertrain model is capable of meeting the required driving patterns.

80

Fig.3- 8 NEDC Cycle

Fig.3- 9 UDDS Cycle

The economic performance is judged here in a running distance of the

same given battery depth of discharge, which is from 95% to 10%. The

simulation results are presented in Table.3-2. Fig.3-10 and Fig.3-11 present

the operating points of the electric motor using the economic shift schedule

for both NEDC and UDDS cycles, respectively. From Table.3-2, in the

economic shifting schedule, a longer running range can be achieved in both

the NEDC and UDDS cycles, which are 118.68km and 112.97km,

respectively. This demonstrates that the economic shift schedule can

optimize the working point of the motor and improve system efficiency.

This is also demonstrated in Fig.3-10 and Fig.3-11, the working points with

the economic shift schedule can be optimized to comparatively high

efficiency areas which improves the average power-train operating

efficiency. It should be noted that the simulations herein are conducted

81

using a constant temperature motor model. In a conventional motor, the

operating efficiency is greatly influenced by the motor operating

temperature. The impact and control of the motor temperature will be

considered in further studies.

Table.3- 2 Economic Performance

Economic Performance Dynamic Shifting Schedule

Economic Shifting Schedule

General Shifting Schedule

Range with NEDC cycle(km)

111.73 118.68 114.09

Range with UDDS cycle(km)

107.97 112.97 110.36

Fig.3- 10 Motor Working Points in NEDC

Fig.3- 11 Motor Working Points in UDDS

82

3.5 Conclusions

To optimize vehicle dynamic and economic performance, a shifting

schedule calculation method for PEVs was provided in this chapter using a

graphical development method, and this is adapted to be used in

simulations and experimental work. Using the acceleration curve of two

gears in the same throttle degree the intersection creates the ideal dynamic

shifting point and this is necessary for the downshifting line to have

hysteresis to avoid shifting hunting. The economic shift schedule is

developed by taking a constant output torque and across a number of

vehicle speeds which determine the efficiency of the electric machine and

generate an efficiency curve. Where the two efficiency curves intersect is

the point of transition from higher efficiency in one gear to higher

efficiency in the other gear. This is therefore the optimum shift point to

maximize the operating efficiency of the PEV. As long as gear shifts are

initiated according to this schedule the EM will be maintained at the higher

efficiency operating region and as a result the proposed method will

maintain more efficient operations of the PEV.

The shifting schedule which is used on vehicle should consider both

the dynamic and economic performance. Actually in the low throttle value,

the driver expects economic performance. But in the big throttle value, we

can see a short acceleration time is expected by driver, so more dynamic

performance should be considered. Actually from the Fig.3-6, in the area of

throttle greater than 40%, the shifting schedule line is increasing very quick,

it can satisfy general dynamic acceleration demand.

To demonstrate the effectiveness of the economical shifting schedule,

a PEV model was built in the AMESim environment, which includes the

battery, motor, transmission, vehicle, and driver models, and control

models, all of which were built in Matlab/Simulink. Through acceleration

83

and running cycle test simulations the results show that the shifting

schedule developed in this chapter can improve the vehicle dynamic and

economic performance significantly.

This chapter is based on a two speed DCT development project;

consequently just 2 gear ratios are used in the simulation. It provides

theoretical support to PEV power-train system matching and optimization,

and can be extended to transmissions with more than two gears to evaluate

the extent of the application of multi-geared transmissions to PEV power-

trains. Also in this chapter, only the motor’s driving efficiency has been

considered. This does not include the generating efficiency but the brake

regeneration is a unique characteristic of motor driven vehicles and the

shifting schedule in the braking condition is also a key point for improving

system efficiency. Further work will be done on these key points.

References

[1] AnLin Ge. Vehicle Auto Transmission Theory and Design. China Machine Press 1991.

[2] ZhongHua Lu. Research on the Control Technology of Double Clutch Transmission with Two Gears Based on Pure Electric Car. PHD paper, Jilin University, China, 2010.

[3] Huang Ying, Shi Xianlei, XuShili, et al. Design of Gear Shifting Rules on the Basis of Power Performance and Fuel Economy and Experimental Study on Them. Automobile Technology 2004; 11.

[4] YangYi, Jiang Qinghua, ZhouBing, et al. A Study on the Optimal-Power Shift Schedule for Electric Vehicle. Automobile Technology 2004; 3.

[5] Yimin Gao, Ali Emadi, Mehrdad Ehsani. Modern Electric Hybrid Electric and Fuel Cell Vehicles Fundamentals Theory and Design. USA: CRC Press, 2010.

[6] YuZhiSheng. Automobile Theory. Beijing: China Machine Press, 2000.

[7] Huang Juhua, Xu Shihua and Xie Shikun. The Design of Automatic Transmission Control System of Electric Vehicle. Journal of Jinggangshan University (Natural Science) 2011; 32.

84

[8] P. D. Walker, S. Abdul Rahman, N. Zhang, et al. MODELLING AND SIMULATION OF A TWO SPEED ELECTRIC VEHICLE. International Conference on Sustainable Automotive Technologies 2012, Melbourne, Australia, 21-23 March 2012.

85

Chapter 4 Two Speed DCT Shifting Control

Strategy

4.1 DCT Shifting Control Analysis

The main consideration of the DCT shifting will be the clutch-to-

clutch shift control. In the process of shifting, one clutch is disengaged and

another engaged. For the strong nonlinearity of the clutches during the

process of shifting, and the torque coupling in the torque transfer, it is

almost impossible to calculate the transfer torque. Furthermore with the

hysteresis and nonlinear nature of the hydraulic system, implementing

clutch-to-clutch shifting is exceedingly complex. Appropriate methods to

control the shifting process have thereby become crucial to technological

development. Indeed, it significantly affects the shifting time and comfort

levels.

There are several control methods that have been applied to dual

clutch transmissions, ranging from basic open loop methods through to

fuzzy control techniques. A basic method applied by Goetz [24] and also by

Zhang, et al, [25] and Kulkarni, Shim & Zhang [26] is to use a controlled

signal to perform the shift. Goetz [24] demonstrates both speed and torque

control techniques, while Zhang [25] and Kulkarni [26] adopt pressure

profiles to perform the shift. In this way it is somewhat similar to speed

based control techniques for automatic transmissions.

Open loop control methods have be adopted by Song [27] for heavy

vehicle applications. There is demonstrated level of success for the

capability to perform shifts, and the results are similar to those of a single

86

clutch transmission. Simulations demonstrate that there is reduced speed

loss during the shift from the DCT. However, there are still significant

improvements that can be made to the shift responses to improve the

engagement, noting particularly the interruption of the vehicle speed during

shifts. Fuzzy control techniques have been applied by Xuexun [16] for

shift control of a DCT. While it has been proved that such a method

achieves a reasonable shift in the DCT, Wu, et al, [30] makes use of

feedback linearization to reduce power-train complexity and applied PID

control in the resulting model with some success. Analysis of the clutch

force by Liu, et al, [28] has enabled an increase in shift quality, while [31]

uses the applied force as a measure of shift quality, combining the clutch

and engine as control variables.

As electric vehicles do not require the idle status capability which is

necessary feature of internal combustion engines, the process of the launch

is simplified. First gear is engaged and the drive motor directly launches

the vehicle. The launch control becomes easier and will not influence the

ride comfort. Thus, in this chapter it is only necessary to study the up-shift

and down-shift control algorithms and simulations.

For the difference between motor and engine, the engine has idle

speed and the launch control is the clutch engagement control, but the

motor has no idle speed so clutch can be pre-engaged and the motor

controlled from the zero speed to drive the vehicle. There is no shaking and

slippage during the launching of the motor driving system. Accordingly, as

the shifting control of the PEV system is simpler than the engine system,

only the up-shifting and down-shifting are studied in this paper.

4.1.1 Shifting Process of PEV DCT

The transmission shifting can be generally divided into the power-on

87

gearshift and power-off gearshift, in which the power-on gearshift takes

place whilst the motor drives the wheel, and the power-off gearshift takes

place whilst the wheels drive the motor.

Up-shifting can thereby be divided into the Power-on Up-shift and

Power-off Up-shift. Most of the up-shifting is the Power-on Up-shift. The

Power-off Up-shift generally happens in coasting with the gears. Down-

shifting can be divided into the Power-on Downshift and the Power-off

Downshift. The Power-on Downshift generally happens during hill

climbing with the throttle but the vehicle speed decreases. Most of the

Downshifting is in the Power-off Downshift [3]. The details are displayed

in Table.4-1.

Table.4- 1 Shifting Classification in different situations

Classification Situations Remark Power-on Upshift Acceleration with Throttle Most of the upshifting Power-off Upshift (1) Acceleration and release throttle;

(2) Down Hill coasting and acceleration;

Power-on DownShift

Uphill deceleration (with throttle);

Power-off Downshift

(1) Braking (2) Uphill (without throttle)

Most of the downshifting

Briefly, DCT shift control is split into the torque phase and inertia

phase. The purpose of the torque phase is to seamlessly hand dynamic

friction torque from the originally engaged clutch to the clutch that is the

target for engagement. Towards the end of the torque phase, the control

must perform the following tasks: [32] determine the target torque at which

the releasing clutch will transition from the stick to the slip state; determine

the required torque at the engaging clutch which is required to maintain the

acceleration of the vehicle with minimum loss of the tractive load, and

transfer the torque from the releasing clutch to the engaging clutch in a

manner that minimizes vehicle transients.

The inertia phase begins once the target torque has been met during

88

the torque phase. Control then proceeds as follows [32]: determine the

target torque for the engaging clutch; hold pressure at the desired torque,

and when the speeds are matched set the pressure to the maximum and lock

the clutch. For the adoption of a torque orientated control strategy in DCT

control, the inertia phase of control requires that the clutch torque is

maintained at a constant torque that is equivalent to the vehicle angular

acceleration and any resistance torque. Though it is possible to use higher

torques to reduce the shift times, this is likely to result in surging or more

significant power-train transients than is desirable during shifting in lightly

damped power-trains.

According to the shifting condition analysis above, we can find out

that during the power-on shifting, the power is transferred from engine to

wheel, but during the power-off shifting, the power is transferred from

wheel to engine. In the up-shifting, engine speed decreases because of the

speed ratio reduction. And in down-shifting, engine speed increases.

Under the four situations in Table.4-1, torque phase and inertia phase

classifications are as follows in Table.4-2. Power-on Up-shifting is the

same as power-off Down-Shifting, firstly torque phase and then inertial

phase. Power-on Down-shifting is the same as Power-Off Up-Shifting,

firstly inertial and then torque phase.

Table.4- 2 Shift Process Classification

Up-Shift Down-Shift Power-On 1. torque phase

2. inertia phase 1. inertia phase 2. torque phase

Power-Off 1. inertia phase 2. torque phase

1. torque phase 2. inertia phase

At the beginning of the torque phase the off-going clutch is brought to

a state where it slips. The clutch slip is then controlled to stay at a small but

constant reference value. This is achieved by a closed-loop controller

which manipulates clutch pressure at the off-going clutch. Whilst clutch

slip is controlled at the off-going clutch; the pressure at the oncoming

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clutch is ramped up in an open loop way to transfer engine torque from the

off-going to the oncoming clutch. As a consequence of the increase in

pressure at the oncoming clutch, the clutch slip controller decreases the

pressure at the off-going clutch in order to maintain the reference slip value.

At the point where the full motor torque has been transferred to the

oncoming clutch, the pressure at the off-going clutch becomes zero and the

off-going clutch disengages automatically without creating a negative

torque. The transmission output torque has dropped according to the gear

ratio and a transition to the inertia phase can take place. Also a motor-

assisted closed-loop control of the inertia phase has been indicated where

engine torque is reduced by a closed-loop control motor torque to follow a

specified motor speed reference trajectory to achieve synchronization. [15,

24]

4.1.2 Shifting Quality Criterion

Shifting quality is the extent under which transmission can complete

the shifting process quickly and stably. In the meantime, it is important to

evaluate and maintain the power-train service life, considering in particular

clutch wear. Good shifting quality requires a stable gearshift with

minimized shock, however, with the multi-state system of the power-train,

even for dual clutch transmissions, it is impossible to eliminate jerk during

the shifting and one can only reduce it [33-35]. In this chapter, we choose

the three most commonly used quality criterions: shifting time, jerk and

sliding friction power.

1) Shifting Time

The shifting time is from the moment the controller gives the order to

clutch until the clutch unites, completely finishes the shifting process, and

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the engage clutch is synchronized. A good shifting quality requires

reducing the shifting time as much as possible in order to fit the

requirements of the ride. However, the driver only observes the inertia

phase of gear change as it distinguishes the two periods of gear engagement.

2) Shifting Jerk

Jerk is the change rate of vehicle longitudinal acceleration. If the

vehicle jerk is too large, it indicates that passengers observe an obvious

forward or backward shock thereby degrading the driving comfort during

the shift. Its mathematical expression is as follows:

δ

η 4-1

Where is speed, δ is vehicle gyrating mass conversion factor, is

vehicle mass, is gear ratio and is final ratio, η is transmission

efficiency, is wheel rolling radius, is clutch friction torque.

Equation 4-1 shows that jerk is proportional to the change rate of

clutch torque. The faster the torque changes, the shorter the clutch shifting

time will be, and the larger the jerk of the driven system will be.

According to passengers’ subjective feelings and their evaluation

responses, the criterions are various, Germany recommends ,

and in China it is .

3) Shifting Sliding Friction Loss Power

Clutch sliding friction loss power is a measurement of the friction

work of driving and the driven friction plates during the clutch coupling

process. It can be used to evaluate and indicate the service life of a clutch.

Its definition is as follows:

4-2

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Where is sliding friction loss power, ω is speed, m is motor, 1 and

2 are clutch 1 and clutch 2. t1 and t2 are the clutch 1 and clutch 2 engage

or disengage time.

Equation 4-2 indicates that sliding friction work is related to the

relative angular velocity between driving and the driven disc during the

engagement time. As engagement time has an inverse correlation with the

clutch friction torque's change rate, sliding friction work and jerk represent

a pair of contradictory evaluation indicators.

If the shift time is too short, it may result in serious vibrations and

jerking which in turn decreases shift smoothness. However, when the shift

time is too long, it will increase the slipping friction power and shorten the

service life of clutch discs. Therefore, excellent shift quality requires the

shift time to be as short as possible based on smooth shift. With shorter

shift time, the torque interruption is shorter and the drivability and shift

quality is better [36].

4.2 Two Speed DCT Transient Modeling

The simplified power-train of two speed DCT is schematically shown

in Fig. 4-1 with clutches in the slip state, consistent with both clutches

being energized during general shifting conditions. The motor inertia takes

input in the form of motor torque and outputs it to the clutch drum via shaft

stiffness and the damping element. The clutches are coupled to the drum so

there are two possible torque paths available as inputs that transmit power

from the motor and clutch drum, as well as two outputs that drive the gear

set either separately or simultaneously. The transmission transfers the

clutch output torques to the propeller shaft, where it drives the vehicle

inertia which is subject to various loads such as air drag and rolling

resistance. The equations of the motion of the open model are:

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Fig.4- 1 Dynamic Model of Pure Electric DCT

4-3

4-4

4-5

4-6

where θ and its two derivatives are the rotational displacement, velocity,

and acceleration, respectively, γ presents the gear ratio, I is the inertia

element, C is damping coefficient, K is stiffness coefficient, and T is torque.

For subscripts M represents motor, D for clutch drum, T for transmission, V

is the vehicle, C is clutch, 1 and 2 represents the two clutches and

respective gears. When either of the two clutches is locked, the vehicle

reverts to a three degree of freedom model where the closed clutch merges

the inertia of the drum with the transmission via a reduction gear. This gear

ratio is the combination of both the transmission ratio and final drive ratio

for this model.

The equations of motion of the closed model are:

4-7

4-8

4-9

The use of piecewise clutch models to account for transitions between

dynamic and static frictions is performed by [38,12], where, upon clutch

synchronization an algorithm estimates the clutch torque and compares it to

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static friction to determine if the friction lock is successful. A similar model

is adopted by [8]; however the post lockup torque is not well defined. In

Goetz [11,15] a transfer function model of the hydraulic system is utilized

over a detailed model, citing issues with the discontinuous contact in the

clutch as a limiting factor. In this paper a nonlinear spring contact model is

used to overcome this issue. Accordingly, a fourth variable of piston

displacement is introduced to the piecewise model of the clutch, where the

return spring separates the plate contact and the clutch torque drops to zero,

excluding a small viscous contact component.

The piecewise model of the clutch is defined as a combination of

dynamic and static friction where the static friction is calculated as the

average torque in the clutch and is limited to the static friction of the clutch.

This is defined as follows:

(4-10)

where, n is the number of friction plates, X is piston displacement and

X0 is the minimum displacement required for contact between friction

plates, is dynamic friction, is static friction, and are the outside

and inside diameters of the clutch plates, and is the pressure load on the

clutch. A relatively simple model of the coefficient of friction of the

clutches is presented as having dynamic friction, a static friction, ,

for an absolute clutch speed of approximately zero, such that numerical

error in calculations is eliminated without negatively affecting results. This

includes the phenomena of stick-slip.

The average torque, being derived from the open clutch equations

for each of the two engaged clutches is as follows:

94

(4-11)

(4-12)

(4-13)

Eqs. (4-12) and (4-13) are realized by re-arranging Eqs. (4-8) and (4-

9), respectively. Determining the average torque for clutch 1 or clutch 2 is

achieved by using the alternate subscripts of 1 and 2 in sequential order.

The average torque is important for torque based control of dual clutch

transmissions as it is the target for engaging clutch control in torque based

control applications.

4.3 Shifting Control Strategy

General control methods fall into two categories: open loop and closed

loop controls. The benefits of the open loop clutch control are its hardware

compactness and low cost but the main challenges lie in two aspects:

consistent initial condition and optimal control process. In addition, for the

closed loop clutch control, the main challenges are to form a feedback loop

in a structurally compact, precise and robust fashion [37].

The shifting control strategies in this chapter are Power-on Up-shift,

Power-off Up-shift, Power-on Down-shift and Power-off Down-shift. The

judgment conditions are the current gear number and gear number in last

time while the throttle degree is to judge power-on or power-off. Detailed

control algorithms are displayed in Fig. 4-2.

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Fig.4- 2 Shifting Condition Judgment

4.3.1 Power-on up-shift control

According to Table 4-2, during the up-shifting process is the torque

phase and then the inertial phase. The tasks of the torque phase are [11]:

a. Determine the target torque of off-going clutch, control it from

engaged to slip status;

b. Determine the target torque of on-coming clutch, keep the vehicle

acceleration performance and reduce the power loss during the

shifting;

c. Transfer the power from the off-going clutch to the on-coming

clutch smoothly;

The tasks of inertial phase are:

a. Determine target torque of the on-coming clutch;

b. Keep the pressure for expected torque;

c. When using the speed Synchronization, increase the pressure to the

maximum value and lock the clutch;

Literature [1-3] describes the relationship of these two phases:

Torque phase - the torque switching time is a key factor in shifting

96

control. If the engagement is too early, shifting shaking will happen. If the

engagement is too late, a torque hole will appear. The ideal method is to

disengage the off-going clutch first and then keep a little bit of slip, then

engage the oncoming clutch.

Inertial phase - engine speed should decrease to synchronize the on-

coming clutch speed. The engine torque can be decreased to reduce vehicle

shaking. However, in the PEV power-train system, the motor can be

controlled more quick and smoothly, so a better shifting performance can

be expected.

(a)

(b)

(c)

Fig.4- 3 Power-On Up-Shifting Process Analysis

97

As presented in the control algorithm of Fig.4-3, at the beginning of

the power-on up-shift, clutch 1 pressure is reduced and clutch 2 is prefilled

resulting in the initiation of slip in clutch 1. After a short time delay, the

torque transfer begins with clutch 1 slip compensation control. The purpose

of this is to control the clutch 1 slip at a given value to guarantee the output

torque without generating transient shock. The slip value recommended in

literature [24] is 5rpm. As it is impossible to control the slip at a constant

value, a slip zone of 8-12rpm is selected in the simulations. To compare

and confirm what the best parameter is chosen in the control, a series of

slip values are set and the simulation differences are provided towards the

end of this chapter.

When the clutch 1 pressure decreases down to zero, the torque phase

finishes and the inertial phase begins. A lot of literature [11, 24]

recommends the speed closed loop control in this phase. This allows the

engine speed to follow a prescribed speed profile. It is a good method to

control the shifting process of speed transfer, but as the speed closed loop

will bring torque change in the clutch it may cause a transient response and

fluctuation in the output torque to the vehicle. It is therefore very difficult

to adjust the control parameters to achieve smooth shifting comfort. In this

chapter a simple sectional torque control algorithm is studied whereby the

motor speed synchronizes with the clutch 2 speed, then the clutch 2

pressure is increased to line pressure and recovers the motor torque to drive

the request value. The shifting process is then completed.

The process of speed synchronization is divided into three sections,

the first section is when |Nmotor-Nc2|<=Value I, and the motor torque is

reduced gradually. The second section is when Value I <|Nmotor-

Nc2|<Value II, and the motor torque is maintained at the desired output

value. The third section is when |Nmotor-Nc2|>=Value II, and the motor

torque begins to increase again. Here the parameters value I and value II

98

both need to be calibrated. When the motor torque reduces, the output

torque of the vehicle inevitably decreases. To avoid a huge negative jerk in

the shift transfer, an appropriate minimum torque limit should be included.

A series of minimum torque limit values are given and simulations are

conducted in the latter part of this chapter to confirm the correct value

selection.

Fig.4- 4 Control Algorithm of Power-on Up-shift

Fig.4-5 is the control graph of power-on up-shifting. The first

component (a) is the pressure of clutch 1 and clutch 2, (b) motor torque and

(c) is the speed changing process of the clutches and clutch drum during the

shifting.

99

Fig.4- 5 Power-on Up-shift Control

To demonstrate the control effects, more simulation results are

presented in Fig. 4-6 whereby Fig. 4-6 (a) is the output torque of gearbox,

and Fig. 4-6 (b) and (c) represent vehicle jerk and slip friction loss during

shifting, respectively. Owing to these curves, the shifting control is smooth

and can be finished in less than one second. The jerk of vehicle is under 103/ sm , and there are no significant transient vibrations and no large slip

friction losses occurring during the shift process.

100

Fig.4- 6 Power-on Up-shift Simulation Results

Simulation in different control parameters

As mentioned above, several control parameters are identified in the

control algorithms to simulate and confirm the control parameter selection

and calibration. To investigate the influence on shift performance, more

simulations are conducted in the control of power-on up-shifts. These

include simulations in different slip values for the torque phase and

different minimum torque limit values in the inertial phase. Results are

show in Figs.4-7 to Fig.4-9.

In Fig.4-7, different slip rotating speeds are given from the minimum

<3 rpm to maximum 30-40 rpm. Fig. 4-7 (a) shows the vehicle jerk during

gear change. Obviously, if the slip value is too small, the maximum jerk

value can be restricted but high frequency vibration can arise at the start of

the torque phase. Meanwhile, if the slip value is too large, the vehicle jerk

will increase as a consequence. So in this chapter a suggested value of 8-12

rpm has been selected. Note also that the slip friction power loss is not

significantly influenced by the parameter variation, as shown in Fig.4-7 (b).

101

Fig.4- 7 Simulation results under different clutch slip rotation speeds

Fig.4- 8 Simulation results under different motor minimum torque limits

102

Fig.4- 9 Simulation results under different motor minimum torque limits

Fig.4-8 and Fig.4-9 are simulation results for different inertial phase

minimum motor torque limitations. In Fig.4-8, minimum motor limits are

selected from 0.7*Tmotor to -0.3*Tmotor. During the changing, the

shifting time is dramatically shortened from 0.4s to less than 0.15s but in

Fig.4-9 it is shown to result in increasing the maximum vehicle jerk. If the

minimum motor torque required is to be -0.3*Tmotor, a -50 3/ sm vehicle

jerk is shown in the figure and this will negatively influence the ride

comfort of passengers to a large degree. Also notice that in Fig.4-9 (b), as

the shifting time is shortened, the slip friction loss power also decreases.

From the simulation results we find that although the motor has more

flexible torque control capability, there are more issues which arise in shift

quality. This, in turn, raises issues in the clutch-to-clutch shifting process

and leads to a serious control question which takes into account all vehicle

performance parameters, including jerk, slip friction work and shifting time,

in order to decide on the most appropriate calibrated control parameters.

103

4.3.2 Power-off up-shift control

Fig.4- 10 Power-Off Up-Shift Process Analysis

Power-Off Up-Shifting is easier than Power-On Up-Shiftingbecause

there is no power transfer at the forward direction, and the drag torque of

the motor is smaller than the engine. There will be no obvious torque

interruption during the shifting so the requirement of control can be

decreased.

104

From analysis in Table 4-2, Power-Off Up-Shifting is started from the

inertial phase and then the torque phase. Meanwhile, the speed

synchronization occurs and then the torque transfer. Detailed process

analysis is presented in Fig.4-10 and a control algorithm is presented in

Fig.4-11.

As displayed in the control algorithm in Fig.4-11, at the beginning of

the Power-Off Up-Shift, the clutch 1 pressure is reduced and the clutch 2 is

prefilled resulting in the initiation of slip in clutch 1. After a short time

delay there is the speed synchronization. The process of speed

synchronization is divided into two sections, the first section is when

|Nmotor-Nc2|>=Value I, and the C2 pressure is increased gradually. The

second section is when Value II <|Nmotor-Nc2|<Value I, and the C1 and C2

pressures are both maintained at the desired output value. When |Nmotor-

Nc2| <= Value II, the speed synchronization and inertial phase finish and

the torque phase begins. In the whole inertial phase process the motor

torque remains unchanged.

In the torque phase, torque transfer begins with clutch 1 slip

compensation control. The purpose of this is to control the clutch 1 slip at a

given value to guarantee output torque without generating transient shock.

This control is the same as the Power-On Up-Shift control. When clutch 1

pressure decreases down to zero and clutch 2 pressure is increased to line

pressure, the torque phase finishes.

105

Fig.4- 11 Control algorithm of Power-Off Up-Shift (a)

(b)

(c)

Fig.4- 12 Power-off Up-shift Control

Fig.4-12 is the control graph of Power-off up-shifting. The first

component (a) is the pressure of clutch 1 and clutch 2, (b) motor torque and

106

(c) is the speed changing process of the clutches and clutch drum during the

shifting.

(a)

(b)

(c)

Fig.4-13 Power-off Up-shift Simulation Results

Simulation results are presented in Fig.4-13. Fig. 4-13 (a) is the output

torque of the gearbox. Fig. 4-13 (b) and (c) are vehicle jerk and slip friction

loss during shifting, respectively. From these curves, the shifting control is

smooth and can be finished in less than one second. The jerk of vehicle is

near 10 3/ sm , and there are no significant transient vibrations and no large

slip friction losses occurring during the shift process.

107

4.3.3 Power-On Down-Shift Control

(a)

(b)

(c)

Fig.4- 14 Power-On Down-Shift Control Process Analysis

The power-on down-shift process is the opposite to that of power-on

up-shifting; it begins with the inertial phase (Fig.4-14). As shown in the

Fig.4-15 control algorithm, at first the clutch 2 pressure is reduced and the

clutch 1 is pre-filled, then the pressure is set to initiate the slip of clutch 2.

For down shifting from 2nd gear to 1st gear, the speed of the motor will

increase to synchronize the clutch 1 speed. If the motor torque is less than

108

maximum output value, an increasing torque requirement can be given to

shorten the inertial phase. The algorithm is the same as in the inertial phase

of power-on up-shift control; and the parameters of Value III and Value IV

are selected and calibrated. When the motor speed has synchronized with

clutch 1 speed, the inertial phase finishes and the torque phase starts.

In the torque phase, clutch 2 pressure is reduced and the ramp up of

clutch 1 pressure occurs. Also the same slip feedback compensation control

as in power-on up-shift control is adopted here, to control the clutch 2 slip

value in the given target value during torque transfer and ensure smooth

shifting. When the clutch 2 pressure is reduced to zero, there is an increase

of clutch 1 pressure to line pressure and the shifting process completes. In

addition, the motor torque is returned to the driver demand values.

Fig.4- 15 Control algorithm of Power-on Down-shift

109

Fig.4-16 Power-on Downshift Control

Fig.4-16 is the control graph of power-on down-shifting. The first

component (a) is the pressure of clutch 1 and clutch 2, (b) motor torque and

(c) is the speed changing process of the clutches and clutch drum during the

shifting.

Fig.4- 17 Power-on Downshift Simulation Results

Simulation results are presented in Fig.4-17. Fig. 4-17 (a) is output

110

torque of gearbox. Fig.4-17 (b) and (c) are vehicle jerk and slip friction loss

during shifting, respectively. From these curves, the shifting control is

smooth and can be finished in less than 700ms. The jerk of vehicle is less

than 10 3/ sm , and there are no significant transient vibrations and no large

slip friction losses occurring during the shift process. Further, we can see a

better shifting result than the Up-shifting.

4.3.4 Power-off Down-shift control

The shifting control of Power-off Down-shift is the same as the

Power-On Up-Shift, first torque phase and then inertial phase. As there is

no torque transfer during the shifting, the shifting can be much more simple

than power-on shifting. Detailed control algorithms are in Fig.4-18 and

Fig.4-19.

111

(a)

(b)

(c)

Fig.4- 18 Power-off Down-shift Control Process Analysis

112

Fig.4-19 Control algorithm of Power-off Down-shift Fig.4-20 and Fig.4-21 are the control figures and results in the

simulation. Fig.4-21 (b) and (c) are vehicle jerk and slip friction loss during

shifting, respectively. From these curves, we can see the max jerk of

vehicle is just 8 3/ sm , also the shifting time can be finished in 700ms.

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Fig.4- 20 Power-off Downshift Control

Fig.4- 21 Power-off Downshift Simulation Results

114

4.4 Shifting Control Strategy with Motor Braking Torque

Control

(a)

(b)

(c)

Fig.4-22 Power-On Up-Shift control with Motor Braking Control

For much more flexible control of the motor and faster control

reaction time, the DCT shifting control mounting with electric power-train

is the best option. In this thesis we apply the motor braking torque control

in the inertial phase of the Power-On Up-Shift control algorithm.

Control figures are in Fig.4-22. Torque phase control is the same as

the traditional control method, but during the inertial phase, motor braking

torque is added to actively control motor speed synchronization and the

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shifting time can then be significantly shortened.

Detailed control algorithms used in simulation are in Fig.4-23. Here

we can see the minimum throttle command of motor can be set to -30%, so

an active braking torque is used to actively reduce the motor speed and

shorten shifting time. From the simulation, a total shifting time of 0.5s can

be reached, and the motor speed synchronization time can be shortened to

150ms. However, from Fig.4-24 the vehicle jerk is almost 50 m/s3. It is

much too big and will introduce a bad ride in terms of comfort during the

shifting.

Fig.4-23 Power-on Up-shift Control with Motor Braking Torque

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Fig.4-24 Power-on Up-shift Simulation Results (With Motor Braking Torque)

4.5 Conclusions

Dual clutch transmissions are recognized to be suitable for the electric

drive application for multi-speed pure electric vehicles. These

transmissions provide desirable qualities for high efficiency automotive

platforms, including high driving efficiency. With appropriate control

methods, good shifting comfort can be achieved.

To investigate shift control and its calibration of a two speed DCT

electric drive power-train, this chapter analyzes the shifting process.

Detailed shifting control algorithms are developed which include power-on

and power-off methods.

The simulations conduct the modeling, control and simulation of an

EV. For clutch-to-clutch shifting control studies, a dynamic model has been

analyzed and built. A good gear shift performance is demonstrated in the

simulation results. Also, to evaluate the control parameter’s influence on

vehicle performance, a series of slip value and minimum motor output

torques are set in the power-on up-shift control. After comparison of

simulation results, correct control parameters can be selected in the control

system. Calibration is performed through simulation to demonstrate how it

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is possible to achieve high quality gear shifts and the influence of motor

throttle control on overall shift quality. There are significant trade-offs

between short, fast gearshifts and the quality of shift achieved.

In the final part of this chapter, the shifting control using active motor

braking torque is provided. A simple description is given to introduce the

control algorithm. Detailed control algorithms are also studied and

simulated. Furthermore, to fully realize the achievement of shift

performance, further calibration is studied with experimental calibration in

the proceeding chapters.

References

[1] G. Lechner, H. Naunheimer, Automotive Transmissions: Fundamentals, Selection, Design and Application, Springer, Berlin, New York, 1999. [2] M.A. Kluger, Denis M. Long, An Overview of Current Automatic, Manual and Continuously Variable Transmission Efficiencies and Their Projected Future Improvements, SAE paper 1999-01-1259, 1999. [3] Lucente, G., Montanari, M., and Rossi, C. 2007 “Modelling of an automated manual transmission system” Mechatronics 17:73:91. [4] W. Grobpietsch, T. Sudau, Dual Clutch for Power-Shift Transmissions – A Traditional Engaging Element with New Future, VDIBerichte Nr. 1565, 2000, pp. 259–273. [5] Walker, P.D., Zhang, N. and Tamba, R. “Control of gear shifts in dual clutch transmission powertrains” Mechanical Systems and Signal Processing 25 (6), 1923-1936. [6] Goetz M, Levesley, M Crolla D, A gearshift controller for twin clutch transmissions, VDI Berichte, 2003, 1786: 381-400 [7] B. Matthes, “Dual clutch transmissions - lessons learned and future potential,” SAE, Tech. Rep. 2005-01-1021, 2005. [8] M. Kulkarni, T. Shim, and Y. Zhang, “Shift dynamics and control of dual-clutch transmissions,” Mechanism and Machine Theory, vol. 42,pp. 168–182, 2007. [9] S. J. Park, W. S. Ryu, J. G. Song, H. S. Kim, and S. H. Hwang, “Development of D vehicle performance simulator to evaluate shift force and torque interruption,” Int. J. Automot. Technol., vol. 7, no. 2, pp. 161–166, 2006. [10] Y. Zhang, X. Chen, X. Zhang, H. Jiang, and W. Tobler, “Dynamic modeling and simulation of a dual-clutch automated lay-shaft transmission,”

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Trans. ASME, J. Mech. Des., vol. 127, pp. 302–307, 2005. [11] M. Goetz, M. Levesley, and D. Crolla, “Dynamic modelling of a twin clutch transmission for controller design,” Materials Science Forum, vol. 440-441, pp. 253–260, 2003. [12] Walker, P.D., Zhang, N., Zhan, W.Z., Zhu, B. “modeling and simulation of gear synchronization and shifting in dual clutch transmission equipped powertrains”, Proceedings of the Institution of Mechanical Engineers, Part C: Journal of Mechanical Engineering Science, Feb 2013; vol. 227: pp. 276-287 [13] Galvagno E, Velardocchia M and Vigliani A. Dynamic and kinematic model of a dual clutch transmission. Mechanism and Machine Theory 2011, 46(6) pp. 794-805. [14] Y. Liu, D. Qin, H. Jiang, and Y. Zhang, “A systematic model for dynamics and control of dual clutch transmissions,” Trans. ASME, J. Mech. Des., vol. 131, pp. 06 012.1–7, 2009. [15] M. Goetz, M. Levesley, D. Crolla, Dynamics and control of gearshifts on twin-clutch transmission, Proceedings of the institution of mechanical engineers, Part D: Journal of Automobile Engineering 219 (2005) 951–963. [16] G. Xuexun, F. Chang, Y. Jun, Y. Zheng, Modelling and Simulation Research of Dual Clutch Transmission Based on Fuzzy Logic Control, SAE Technical Paper 2007-01-3754, 2007. [17] AnLin Ge, “Vehicle Auto Transmission Theory and Design”, China Machine Press, 1991.10 [18] Y. Lei, J. Wang, A. Ge, Research on control stratgies of double clutch transmission based on system simulation, in: Proceedings of the FISITA World Automotive Congress, Yokohama, Japan, October 2006, Paper number F2006P041. [19] A. Crowther, N. Zhang, D.K. Liu, J. Jeyakumaran, Analysis and simulation of clutch engagement judder and stick-slip in automotive powertrain systems, Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering 218 (2004) 1427–1446. [20] M. Kulkarni, T. Shim, Y. Zhang, Shift dynamics and control of dual-clutch transmissions, Mechanism and Machine Theory 42 (2007) 168–182. [21]XieFei, SongChuanxue, LiuMingshu, ZhangYoukun,LuYanhui; Research on Co-simulation Platform Based on AMESim and Simulink for Dual State CVT; Automotive Technology; Vol.8 2008 [22]Wang Pengyu,Wang Qingnian, Hu Anping,YuYuanbin;Analysis of regenerative brake system of hybrid bus based on Simulink-AMESim co-simulation; Journal of Jilin University(Engineering and Technology Edition) Vol.38 Sup. Feb.2008 [23]Liao Linqing,ZHANG Dongfang, Qu Xiang, Ke Jingjing; Simulation on the Start-up and Shift Process of Dual Clutch Transmission Vehicle Based on AMESim; Journal of Chongqing University of Technology

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(Natural Science), Vol.25 No.1,Jan.2011 [24] Goetz M, Integrated powertrain control for twin clutch transmissions, Ph. D. University of Leeds, 2005 [25] Zhang Y, Chen X, Zhang X, Jiang H, Tobler W, Dynamic modelling and simulation of a dual-clutch automated lay-shaft transmission, Journal of Mechanical Design, 2005, 127(3): 302-307 [26] Kulkarni M, Shim T, Zhang Y, Shift dynamics and control of dual-clutch transmissions, Mechanism and Machine Theory, 42 2007: 168-182 [27] Song X, Liu J, Smedley D, Simulation study of dual clutch transmission for medium duty truck application, SAE Technical Paper: 2005-01-3590, 2005. [28] Liu Z, Dong X, Qin D, Liu Y, Analysis and control on shift quality of dual clutch transmission, Journal of Chongqing University 2010 33(5): 29-34 [29] Huang Juhua, Xu Shihua, Xie Shikun, “The Design of Automatic Transmission Control System of Electric Vehicle”, Journal of Jinggangshan University (Natural Science) Vol.32 No.1 Jan.2011 [30] Wu M, Lu T, Ni C, Zhang J, Research on feedback linearization control of dual-clutch for dual clutch transmission, Mechanical Science and Technology for Aerospace Engineering,2010, 29(10), 1285-1290 [31] Lu Z, Chang X, Feng W, Up-shift control in wet double clutch transmission, Transactions of the CSAE, 2010, 25(6): 132-136 [32] Paul David Walker; Dynamics of Powertrains Equipped with Dual Clutch Transmissions; PhD thesis 289-290, University of Technology, Sydney; March 2011 [33] ZhongHua Lu, “Research on the Control Technology of Double Clutch Transmission with Two Gears Based on Pure Electric Car”, PHD paper, Jilin University, 2010 [34] XieFei, SongChuanxue, LiuMingshu, ZhangYoukun,LuYanhui; Research on Co-simulation Platform Based on AMESim and Simulink for Dual State CVT; Automotive Technology; Vol.8 2008 [35] Wang Pengyu,Wang Qingnian, Hu Anping,YuYuanbin;Analysis of regenerative brake system of hybrid bus based on Simulink-AMESim co-simulation; Journal of Jilin University(Engineering and Technology Edition) Vol.38 Sup. Feb.2008 [36]Jianguo Zhang, Yulong Lei, Changfu Zong, Hongbo Liu; Shift Quality Evaluation System Based on Neural Network for DCT Vehicles; 2010 Sixth International Conference on Natural Computation (ICNC 2010) [37]Xingyong Song, Design Modeling and Control of Automotive Power Transmission Systems, PHD Thesis of University of Minnesota, June 2011. [38] Walker, P.D., Zhang, N. and Tamba, R. “Control of gear shifts in dual clutch transmission powertrains” Mechanical Systems and Signal Processing vol. 25 no. 6, 2011.

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Chapter 5 Rig Testing

5.1 Testing Rig Design and Analysis

5.1.1 Introduction of Testing Rig

The vehicle testing rig is a piece of equipment which is integrated

with mechanism, electricity, and hydraulics. It mainly includes the basic

stand, transmission driving device, dynamometer, motor speed adjustment

device, transmission clamping mechanism, rig moving adjustment device,

transmission oil recover device, data recording device, and rig control

system.

The key part of the rig is the dynamometer. It mainly includes the

eddy current dynamometer and electric dynamometer. He eddy current

dynamometer uses eddy current effects to produce braking torque. The

value of braking torque can be controlled by an excitation current

control so is easy to automatically control. The difference between the

electric dynamometer and common motor is the stator housing of the

electric dynamometer is supported on a pair of bearings and it can

freely swing around the axis. An arm is fixed on the stator housing

which is connected with the dynamometer mechanism for measuring

torque.

The output shaft of measured power machinery and electric

dynamometer rotor are connected together to rotate. The stator windings of

the armature winding cut magnetic field lines and the induced EMF in the

armature windings. These, in turn, produce a braking torque and steering

opposite the motor as a generator in order to achieve the purpose of the

dynamometer. In contrast, when a current passes through the armature

121

circuit it will be in a magnetic field to produce an electromagnetic force

with the same driver steering torque. The motor as a driving motor is kept

running in order to achieve the anti-drag (which measures the dynamic

mechanical friction power).

For development of testing and calibration of two speed DCT

controlling and debugging, a test rig is based on UTS original power-train

test benches in order to transform the structure of the bench schematic as

shown in Fig. 5-1:

Motor

DCT

MCU

DSPACE-VCU / TCU

radiator

12V

Coolingpump

Fuse andSoftstart

Control and display system

High voltage 380V

Low voltage 12V

CAN

Cooling waterway

ElectricResourse

Flywheels

Dynamometer

wheelsTorque sensor

Half shaft

+-

Analog singal

To CoolingTower

Fig.5- 1 Schematic of Two speed DCT Power-train Rig

In Fig.5-1, the test rig uses a single eddy current dynamometer, but

uses the inertia flywheel system to simulate vehicle inertia weight, while it

links to transmit power through four wheels to simulate the real vehicle

driving conditions. The testing parts are the DCT and motor. In the absence

of a battery, the rig uses a DC power supply systems. Power control system

uses DSPACE to develop an integrated control to drive the system.

The test rig is different from the normal single-input single-axis

dynamometer engine or motor test bench system and it is also different

122

from the conventional dual dynamometer output gearbox biaxial test bench.

Compared to the single-axis test rig, the biaxial test rig coupling structure

meets the precursor-type power system testing so it is easy to install. There

is no power system installation alignment problem. Also it solves the

problem of double outputs by the single dynamometer test structure. In

addition, the flywheel structure serves instead of vehicle inertia. The use of

four wheels simulates real car tires on the vehicle transient response

simulations make it more realistic; especially for the development of the

test bench.

5.1.2 Testing Rig Parameter Matching

The eddy current dynamometer can simulate the vehicle rolling

resistance, air resistance, and gradient resistance, to meet the most basic

needs of control system development. Acceleration resistance is simulated

by the mass flywheel. This is to simulate the vehicle body inertia and road

friction conditions. Uphill drag torque is loaded by the dynamometer.

Table.5-1 is the vehicle driving resistance analysis.

Table.5- 1 Vehicle Driving Resistance Analysis

Simulation Function Equipment Have or Not? Rolling Resistance Common dynamometer load √ Air Resistance Common dynamometer load √ Acceleration Resistance Inertia Flywheel Simulation √ Gradient Resistance (Uphill Parking)

Static loading device, Electric Dynamometer simulation

---

Gradient Resistance (Uphill Running)

Common dynamometer load √

Gradient Resistance (Downhill Drag)

Electric Dynamometer Driving ---

Static Resistance Static loading device, Electric Dynamometer simulation

---

Braking in Running Static loading device, Electric Dynamometer simulation can realise

---

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Rig characteristic field calculation

The UTS dynamometer is a Horiba-WT190 system. The basic

parameters of the rated torque are 600Nm. The rated power is 190kW. The

maximum speed is 10000 rpm and the base speed is 3030 rpm. These

characteristics are shown in Fig.5-2:

Fig.5- 2 Horiba-WT190

The rig has a motor drive system for the UQM-PowerPhase @ 125

system, peak power of 125kW, and peak torque of 300Nm (Fig.5-3).

Although the drive motor power and torque are larger than the real car’s

motor parameters in order to simulate real vehicle performance, the control

limit motor output characteristics make it consistent with the real vehicle

motor characteristics. Therefore, the following matching calculation is

according to actual vehicle motor parameters.

Fig.5- 3 Motor and Controller Used on Rig

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Parameters of rig drive train:

Rig speed box ratio: 5.1:1

DCT ratio: 8.45/5.36;

Torque of motor: 255Nm, Max Speed: 9000 r/min, Base Speed: 3000

r/min;

Wheel Diameter on the Rig: 625mm

The Calculation of rig characteristics:

100005432 8563

422.5

268

89

151

Torque (Nm)

Speed (r/min)

190kW

80kW

600

Characteristics Of Motor

Characteristics Of Dynamometer

Fig.5- 4 Characteristics Matching of Motor and Dynamometer

Fig.5-4 the driving characteristics of dynamometer input torque shows

that the dynamometer can cover the motor drive system features

characteristic curve, and therefore meet the test requirements of the project.

Rig Flywheel Inertia Calculation

The inertia flywheel of the test bench is mainly used to simulate

acceleration resistance. Due to this quality, it includes two parts: the

translation quality and the rotating mass. During acceleration, the inertia is

generated not only from the quality of the translation, but also from the

rotating mass moment of inertia. In order to facilitate the calculation, the

general rotation mass moment of inertia transfers into the translation

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quality of the inertial force.

22

22 vMJ 6-1

6-2

So the rotating inertia is: 2flywheelrMJ 6-3

Where, J- conversion to the moment of inertia at the wheel 2mkg

—Flywheel angular velocity rad/s

M—Vehicle Mass kg

v—Vehicle Speed km/h

—Flywheel radius 0.425m

Whereby the vehicle inertia under no-load and full-load conditions is

as shown in Table.5-2

Table.5- 2 Vehicle Rotating Inertia

Vehicle Mass M (kg Rotating Inertia J2mkg

No load 1780-4*30(tire)=1660 299.84 Full load1780-4*30(tire)+375=2035 367.57

The UTS existing rig flywheel setting of parameters are shown in

Table.5-3:

Table.5- 3 the Existing Rig Inertia

Diameter (mm) Mass(kg) Inertia( 2mkg ) Number Big Flywheel 940 353 44.8 4

Small Flywheel 850 234 23.3 4 Therefore, the inertia of the existing rig flywheel group is:

(44.8+23.3)*4=272.4 2mkg

The existing inertia of the flywheel on the rig is less than the vehicle

demand, therefore it needs to increase the portion of acceleration resistance

set in the dynamometer.

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Fig.5- 5 Rig Inertia Flywheel Group

Dynamometer Parameters Setting

In order to simulate the vehicle driving resistance, the dynamometer

setting items are as follows:

6-4 Wherein, a0, a1, a2, and n0 can be set (a0, a1, a2 units of Nm) to simulate

the vehicle rolling resistance, air resistance, and the acceleration resistance.

According to the vehicle dynamics equation:

6-5 =n/(i/(0.377*r)) 6-6

Where, the speed calculation performed by the dynamometer ends, so

is the front dynamometer gearbox ratio 5.1, n is the dynamometer speed, r

is the radius of the tire rig used (312.5mm);

6-7

Transform the equation to

127

6-8

=

=

=0.3125* 1870+375/2 * 0.015+0 =9.64

half load and no hill --Acceleration testing

= 1870+375 * 0.015+0.3 *0.3125=221 full load of

30% hill climbing --Grade-ability testing

Nm

=811 r/min

Motor Cooling System Selection

The motor cooling system requirements are:

The cooling water pump parameter selections are:

Type MES MR2-25-900

Voltage 12VDC

Max pressure: 0.42bar 6.09psi

Max Speed 1350L/h 22.5L/min

Max current 2.4A

Max temperature 80

Characteristic curve shown in Fig.5-6:

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Fig.5- 6 Cooling Pump (Left) and Characteristics Curve (Right)

5.2 Testing Rig Development

The two major modifications required for the development of the test

facility are the modifications of the existing dynamometer and frame to

locate an EV front wheel drive assembly, and the development of a suitable

power supply. The power-train is comprised of parts and equipment from a

Volkswagen Passat as it is of comparable size to the test vehicles, and is

compatible with the DQ250 transmission.

The primary considerations for design are adaptation of the current

dynamometer and rotating inertias to the new power-train, rigid mounting

for the transmissions and wheels, and also the development of a suitable

power supply to simulate the battery pack. This chapter details the

modification of the original power-train test facilities and the development

of the power supply in the proceeding sections.

5.2.1 Frame development

Conveniently, the prototype transmission is based on a Volkswagen

DQ250 used in the Passat sedan. This was initially chosen because the

engine has similar torque and power characteristics, and the half shafts,

wheel hubs, disk brakes, and tires can be used for the front wheel drive

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assembly. The primary design requirements are for mounting points for the

transmission and two wheel hubs. The purpose of the design is to mount the

two wheel knuckles rigidly such that wheels can drive the dynamometer,

support the transmission, and support the electric motor. Fig.5-7 presents

the subassembly which has been designed for mounting the transmission.

This shows the re-designed platform to mount the EV power-train. The

motor and transmission sub-assembly are mounted as a single unit at three

locations, these being the transmission mounts 1 and 2 and the motor

support.

Fig.5- 7 Power-train Mounting Sub-Assembly 1

Fig.5-8 presents an additional view of the frame sub-assembly, with

three mounting locations for each of the steering knuckles, including the

steer arm, suspension arm, and ball joints. This design rotates the knuckle

90° from its typical orientation and provision has been made for installation

of the disk brakes.

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Fig.5- 8 Power-train Mounting Sub-Assembly 2

Wheel hubs are mounted using the steering knuckle and are designed

to allow for adjustment of the wheel location relative to the primary inertia.

These are shown in Fig.5-9. These mounts are specifically designed to

allow for translation along all three axes to align the wheels with the

dynamometer. Primary loads are supported through the steering and

suspension arms. The ball joint is used as a third mounting point to prevent

rotation of the wheels.

Fig.5- 9 Detailed sub assemblies for knuckle and wheel mounting

The power-train assembly and mounting frame are shown below in

Fig.5-10 for the detailed assembly including rotating inertias for the

dynamometer. Excluded from this design drawing are the shrouds and

additional rotating inertias for the complete test rig. These are removed to

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provide a less cluttered view of the system.

Fig.5- 10 Final Power-train and Rotating Inertia Assembly

5.2.2 Power supply development

In developing a reliable DC power supply, two options were

considered. If a battery pack was used we would need to include high

voltage charging equipment and suitable hazard protections, such as fire

protection, to protect against catastrophic failure of the battery pack.

Alternatively, if a 400V DC power supply was developed, we would not

have to worry about charging or hazard protection in the laboratory setting.

The DC supply is therefore the obvious choice, and is shown in Fig.5-11.

The chosen power supply is capable of 157kW using an ABB DCS550 4

quadrant thyristor drive, and is therefore also capable of regenerative

braking. The installed equipment is shown in Fig.5-12, prior to wiring up

the system. The system is capable of delivering 390 Amps at 400VDC, and

is therefore capable of adequately simulating the required power supply.

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Fig.5- 11 Power Supply layout motor and controller are located after the DC filter

Fig.5-12 Power supply assembly, (left) Isolator and mains contactor, (right) Inductor, capacitors and DCS550 4Q drive

To test economic performance of the system, a simulated SOC value was calculated by the DC current and voltage in software:

dtIVVCAP

SOCSOCoutMAX 1000

***6.3

10

(6-9)

Where SOC0 is the initial value, here we set it at 95%, MAXCAP is

battery capacity, which is set to the same as real vehicle’s battery pack of

22.32kWh. outV is the battery output voltage, which is equal to the DC

power voltage. V and I are the real voltage and current value input from the

DC bus.

5.2.3 Installation

The final installation is shown below in Fig.5-13 for the EV power-

train, including completed installation and frame assembly prior to

133

integrating the power-train.

Fig.5- 13Power-train Rig at University of Technology, Sydney

5.3 Rig Testing Criterion

Since the automatic transmission technology has operated for a long

time, the USA and Europe have formed a relatively complete series of

bench testing criterions, such as the SAE standard which contains various

criterions for automatic transmission assembly, the torque converter, seals

and so on. By contrast, the automatic transmission criterion in China is

more dispersed.

Hydraulic converter performance testing criterion is the only

performance test standard (QC/T557-1999), and 1991 implementation test

methods (QC/T29033-1991) can be used as some references. Other testing

criterions of parts and assembly durability are that some test specifications

are developed by manufacturers according to their needs. There are no

comprehensive and complete standards. On the whole, the automatic

transmission standard in China is still very deficient.

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Automatic transmission test criterions can be divided into the

following two categories based on different testing methods:

(1) Rig Testing Criterions - according to test content, these can be

divided into standard performance test rig and rig reliability test standards.

(2) Vehicle Testing Criterions - this mainly refers to the automatic

transmission vehicle road test to obtain the environment, reliability and

performance associated with the vehicle test standards. According to the

content of the test, vehicle test criterions can be divided into performance

and reliability test criterions. Among them, the performance test criterions

mainly include shock, vibration and noise tests; traction test standards, and

high (low) temperature test standards. Road reliability test standards mainly

consist of high-speed and low-speed power-train standard cycle tests.

From Chinese automotive test standards, the current automatic

transmission tests mainly focus on bench testing, and have developed a

number of appropriate standards. These include hydraulic converter

performance test methods (QC/T29033-91), dynamic test evaluations of the

transmission performance; the torque converter performance test method

(QC/T 557 1999) which primarily tests to determine fluid performance

parameters and the loading force matching performance of the torque

converter; automotive automatic transmission operating device

requirements (QC/T470-1999), experimental evaluations of the shift lever

handling, crisp feel etc. In the vehicle test standards, test conditions are

subject to complex, long test times and there are, as yet, no developed

special test methods and standards [4].

International standards are divided into the following categories: ISO,

DIN, and JIS. These are mainly for the development of some test standards

for safety, appearance, and the main interface of parts. These standards do

not involve testing of the automatic transmission. SAE standards include

automatic transmission assembly and major components. For example, GM

135

(OPEL), ROVEL, ZF and other car manufacturers have a lot of testing

standards for the automatic transmission road test. In general, automatic

transmission testing mainly focuses on the road test for testing the

performance and reliability factors [5] [6] [7].

American SAE-J651 passenger cars and light passenger automatic

transmission and automatic transaxle test criterions provide a mean for

comparing the performance characteristics of the automatic transmission,

which roughly describes the power testing within the scope and provides an

appropriate expression method for the tests data. This test criterion is

essentially a standard automatic transmission assembly specification. The

main tests include: a given input torque traction test, on road simulation

tests of given output torque; driving performance tests; parasitic losses tests;

and the engine full throttle test.

Chinese DCT design and testing standards are an unknown area due to

the special nature of the DCT structure, which contains the partial

characteristics of mechanical gearbox and automatic transmission. As a

result, the testing criterions designed for the DCT may consider a

combination of both.

5.4 Control System Development Based on Rapid Control

Program

5.4.1 Rapid Control Prototyping Technology Introduction

Rapid Control Prototyping technology has been developed from

manufacturing rapid prototyping (Rapid Prototyping, referred to as RP)

technology. The main idea of RP technology is that it is possible to design

products in a virtual environment, shorten their product development cycles,

and reduce development costs. Application of RP technology significantly

136

shortens the time to market for new products, and this saves the cost of new

product development and mold manufacturing. After the introduction of

real-time testing of RP technology, electronic control system design and

control algorithms were renamed rapid control prototyping (RCP)

technology. In the early stages of system development, to quickly build a

controller model, and the entire system offline and online, several tests

were employed to verify the feasibility of the control scheme. This process

became known as rapid control prototyping. The RCP and Hrdware in

Loop (referred to as HIL) simulation system for the design of the electronic

control system provides the advantages of development speed, and

acceleration of the design and development process. It has therefore been

adopted by aerospace companies in the automotive and aerospace fields

and it significantly reduces expensive, disruptive test drive requirements.

The dSPACE real-time simulation system was developed by the

German company which developed the dSPACE control system based on

MATLAB/Simulink and semi-physical simulation of the hardware and

software work platform, and it realized the MATLAB/Simulink/RTW's

completely seamless connectivity. The dSPACE hardware system processor

has high-speed computing power and is equipped with a wealth of I/O

support. This provides users with functional software which is powerful

and easy to use. It also includes the realization of automatic code

generation/downloads and a testing/debugging package of tools. The

dSPACE real-time system control algorithms and logic code acts as a

hardware operating environment. It is connected through the I/O board and

it controls objects for research and experiments to verify the feasibility of

the control scheme. This greatly simplifies the development process and

improves development efficiency.

Fig.5-14 is typical automotive control system V development process,

whereby the rapid control prototyping system can control the system in the

137

principle stage by verification algorithms, and hardware in the loop test

system controller software and hardware. This significantly reduces

development risk as it makes the control system on the vehicle before

commissioning is simulated to verify the true condition of the assessment.

This greatly reduces the post-calibration and test times, shortens

development cycles, and reduces development costs.

Fig.5- 14 “V” Development Mode Based on Rapid Control Prototyping

5.4.2 Hardware

A control system is developed using a platform of DSPACE. Control

program and relative control circuits are also developed to realize the

integrated power-train control.

Fig. 5-15 displays the control systems used in the test, the Dspace

MicroAutoBox and Rapidpro rapid control prototyping systems. Fig.5-16

shows the UTS developed DCT control system which uses rapid

prototyping control systems.

138

Fig.5- 15 DSPACE MicroAutoBox Left /RapidPro Right

Fig.5- 16 DCT Control System base on MicroAutoBox

Fig. 5-17 presnets the DCT rig electrical schematics. There are three

kind of voltage levels in the system. The main power is 400V from the lab

to drive motor. A low power of 12V is employed to drive the control

system and cooling pump, and the torque sensors in the rig. Another 5V

system is used to power the DCT sensors.

139

Fig.5- 17 Electrical Schematics of DCT Testing Rig

140

5.4.3 Software

The control system of the rig should be simpler than the real vehicle

control system, because there are no battery pack and vehicle accessory

systems etc. Therefore the control system mainly utilizes basic logic

control, DCT shift control, and motor torque control.

Control system interfaces are defined in Fig.5-18. The input signals

are from the driver of the key, acceleration pedal, brake pedal, and shifter.

Also the signals are from MCU by CAN, and gearbox signals are the clutch

pressures, output shaft speed and temperature. Meanwhile the solenoid

valve current feedbacks from the RapidPro for clutch pressure control. To

control the shifting of dual clutch transmission, TCU controls the duty

cycle of three solenoids (i.e. main solenoid, clutch 1 solenoid and clutch

2solenoid). Some relays are also controlled by the VCU, such as the sensor

power relay, high power relay, low power relay and cooling pump relay.

141

VCU

MCU

Pedal

Enable

Accelerator

Direction

Brake

Driver

Brake

Key

Relay_low

Shifter P/R/N/D

CAN

TCU DCTK1 Pressure

K2 Pressure

OutputShaft Speed_1

5VGND

Main Soleniod ValveCurrent Main

Temperature

Ready

Analog

OutputShaft Speed_2

RapidPro

K1 Soleniod ValveK2 Soleniod Valve

Current K1

Current K2

Hydraulic Pump

Relay_highRelay_Cooling PumpOn/Off

On/Off

On/Off

dutyCycle Main

dutyCycle K1

dutyCycle K2

On/Off_Relay_Pump

On/Off_Relay_High

On/Off_Relay_low

Targ

et_G

ear(

0/1/

2)

Mot

or_S

peed

Cur

rent

_Gea

rM

otTo

q_R

eq Enable

Current K1

Current K2

MicroAutoBox

Relay_sensorOn/Off

Fig.5- 18 Signals Definition of Rig Control System

The DCT control program is shown in Fig. 5-19, and the control

algorithm is divided into the following sections: (1) Signal input: the key,

gear, throttle, brake signal acquisition and processing; (2) Power on and off

logic: this determines and controls the electric vehicle power on and off; (3)

main control logic: this is conducted to determine the operating conditions

as well as the target gear judgment; (4) Fault diagnosis: fault grading and

processing; (5) SOC estimation: because there is no battery on the rig. This

uses grid power, requiring electricity and the corresponding value of the

vehicle to be running SOC as it simulates the working conditions for the

judges and economic assessment test; (6) Working mode judgment:

according to various operating conditions, operational requirements to

142

drive system contro are divided into parking, limp, sliding, crawling,

braking, driving, reversing etc. In total there are seven conditions; (7) Shift

control logic: shifting process control; (8) Cooling system control: the

motor cooling system control.

143

Fig.5- 19 DCT Control Program

144

In order to monitor and calibrate the system developed in Dspace, the

monitoring software interface is shown in Fig.5-20 to Fig.5-23. Fig.5-20 is

the vehicle monitoring interface for the model real car driving instrument

display and driver throttle, brake and key input control signal. Fig.5-21 is

the motor control interface for the real-time status monitoring of the motor.

Fig.5-23 is the motor control software which comes with a monitoring

interface and functions in the same way as Fig.5-21. Fig.5-22 is the DCT

monitoring and calibration interface for calibration of the control

parameters of the DCT.

Fig.5- 20 Vehicle Monitor

145

Fig.5- 21 Motor Monitor

Fig.5- 22 DCT Shift Monitor and Calibration

146

Fig.5- 23 Motor Control Software

5.5 Rig Testing Results and Analysis

5.5.1 Shifting Control Testing

Power-On Up-Shifting Testing

Spee

d(r/m

in)

Fig.5- 24 Power-On Up-Shifting 1000r/min30Nm

147

Spee

d(r/m

in)

Fig.5- 25 Power-On Up-Shifting 3000r/min25Nm

From Fig.5-24 and Fig.5-25, the power-on up-shifting control can be

finished in 1.2s. During the inertial phase, motor torque drop down close to

zero and then recover. The output torque is smooth, but the shaft speed has

some vibrations in shifting of 500rpm. The up-shifting control of high

speed maybe can get a better comfort than low speed.

Power-Off Up-Shifting

Spee

d(r/m

in)

Fig.5- 26 Power-Off Up-Shifting (500r/min

148

Spee

d(r/m

in)

Fig.5- 27 Power-Off Up-Shifting 4000r/min

The power-off down-shifting can be finished in 0.7s in shifting of

500rpm, but can be longer in high speed of 4000rpm. During inertial phase

the motor torque can be negative, the braking torque deduces speed

synchronize time. And the vibration is lower in high speed shifting of

4000rpm than low speed of 500rpm.

Power-On Down-Shifting

Spee

d(r/m

in)

149

Fig.5- 28 Power-On Down-Shifting 500r/min25Nm

Spee

d(r/m

in)

Fig.5- 29 Power-On Down-Shifting 3000r/min25Nm

Power-on down-shifting can be finished in 0.8s. The motor torque is

kept almost constant. Also the speed vibration is lower in high speed

shifting of 3000rpm than 500rpm.

Power-Off Down-Shifting Testing

Spee

d(r/m

in)

150

Fig.5- 30 Power-Off Down-Shifting(500r/min)

Spee

d(r/m

in)

Fig.5- 31 Power-Off Down-Shifting(3000r/min)

Power-off down-shifting can be finished in 1.5s, but shifting time of

power-off is not important, it will not influence the shifting comfort. From

Fig.5-30 and Fig.5-31, the active motor torque control can reduce speed

synchronize time, but the pressure recover time is little bit longer, maybe

need to optimize.

5.5.2 Temperature Testing

Temperature Testing of First Gear

151

Mot

orSp

eed(

r/min

)

Fig.5- 32 DCT Temperature Testing Results 1st gear

Temperature Testing of Second Gear

Mot

orSp

eed(

r/min

)

0 10 20 30 40 50 60 70 80

0 20 40 60 80 100 120 140

DCT Temperature

DCT Bearing Temperature

Time(m)

Tem

pera

ture

() 71.6

56

0 20 40 60 80

100

0 20 40 60 80 100 120 140

Motor Temperature

MCU Temperature

Tem

pera

ture

()

Time(m)

0 20 40 60 80

100 120

0 20 40 60 80 100 120 140

DCT Temperature

DCT Bearing Temperature

Tem

pera

ture

()

Time(m)

74

104.6

152

Fig.5- 33 DCT Temperature Testing Results 2nd gear

In Fig.5-32 and Fig.5-33, it can be seen that when the motor rotates at

the speed of 5000-6000rpm, the maximum DCT oil temperature is 71.6

and 104.6 respectively in 1st and 2nd gear. This does not exceed the

maximum limit of 140 . During the process, the temperature of the motor

and MCU are 80 and 40 respectively. The DCT bearing temperature is

56 and 74 respectively. This is significantly lower than oil temperature,

and proves the transmission cooling system works well to cool the bearings.

5.5.3 Driving Cycle Testing

NEDC Driving Cycle Testing

0

20

40

60

80

100

0 20 40 60 80 100 120 140

Tem

pera

ture

()

Time(m)

153

Spee

d(r/m

in)

Fig.5- 34 NEDC Driving Cycle

UDDS Driving Cycle Testing

154

Spee

d(r/m

in)

Fig.5- 35 UDDS Driving Cycle

In Fig.5-34 and Fig.5-35, there are two lines in the driving cycles; the

solid line is the target “vehicle” speed of the running cycle, while the

dashed line is the actual running speed in the experiment. Here, the rotating

speed on the transmission output shaft is converted to the equivalent of the

linear vehicle speed by multiplying by the tire radius. In the figures, the

gear shifting between first and second gears during the drive cycle shows

the first gear being used until about 30km/h before the gear shift is initiated.

The benefit of the electric motor is realized in the infrequency of gear

shifting. Obviously, the advantages of the two speed transmission results in

the reduction of the peak motor speed and torque in the prescribed drive

cycle.

The total running range calculated from a single driving cycle

155

according to SOC consumption, as shown in Table.5-4

The battery SOC is estimated across one cycle of the NEDC and the

UDDS patterns and it decreases from 93.7% to 86.2% in a single NEDC

driving cycle. In Table.5-4 there are running range calculation results in the

total SOC scope of 124.66km under the NEDC cycle and 119.02km under

the UDDS cycle. These are closed to the simulation results of 118.68km

and 112.97km, but are a little bit higher than those values. Because the DC

voltage used in the rig to calculate SOC is constant at 380V, the battery real

voltage should be decreased during the SOC reduction.

Table.5- 4 Economic Performance

Initial SOC Final SOC Running Range of Single Cycle km

Calculated Total Running Range(km)

NEDC 93.7 86.2 11 124.66 UDDS 95 86.4 12 119.02

Motor working points in NEDC and UDDS running cycles are as

shown in Fig.5-36. Compared with the simulation results in Fig.2-4, there

are almost the same working points distributions. From that we can

conclude that the rig driving cycle testing matches the theoretical analysis

well.

156

Speed(r/min)

Speed(r/min)

Fig.5- 36 Motor Working Points NEDC/UDDS

5.5.4 Efficiency Testing

Efficiency testing is completed in BJEV. Due to the power-train being

integrated with a motor and gearbox, the assembly efficiency is measured

157

and Fig. 5-37 is the efficiency MAP. To compare the efficiency, the

efficiency MAP of a single reducer is also tested in Fig. 5-38.

0.8

0.8

0.8

0.8

0.85

0.85

0.85

0.85 0.850.85

0.88

0.880.88

0.880.88

0.90.9 0.9

0.92

Speed(rpm)

Torq

ue(N

m)

40kW

80kW

0 1000 2000 3000 4000 5000 6000 7000 8000 90000

50

100

150

200

250

Speed(r/min)

0.8

0.8 0.8

0.85

0.85

0.85 0.850.85

0.88

0.88

0.880.88

0.88

0.9

0.9 0.9

0.92

Speed(rpm)

Torq

ue(N

m)

40kW

80kW

0 1000 2000 3000 4000 5000 6000 7000 8000 90000

50

100

150

200

250

Speed(r/min)

Fig.5- 37 Efficiency MAP of Two Speed DCT Power-train includes Motor and Controller 1st gear/2nd gear

158

0.60.7 0.70.7

0.8

0.8 0.80.8

0.85

0.85

0.85 0.85

0.85

0.88

0.88

0.88

0.9

0.9

Speed(rpm)

Torq

ue(N

m)

40kW

80kW

0 1000 2000 3000 4000 5000 6000 7000 8000 90000

50

100

150

200

250

Speed(r/min)

Fig.5- 38 Efficiency MAP of Single Reducer Power-train (includes Motor and Controller)

In Fig.5-37and Fig.5-38, the high efficiency areas of the two speed

DCT are obviously greater than the single reducer gearbox. These are

employed to compare the efficiency difference directly. The statistics of the

area proportions are in Table.5-5.

Table.5- 5 Compare of Efficiency Area

Efficiency ≥80% ≥85% ≥90% Two Speed DCT(1st gear) 69.6 55.5 18.9 Two Speed DCT (2nd gear) 69.4 55.0 22.6 Single Reducer 68.3 40.5 5.0

From Table.5-5, the area proportions of two speed DCT in the

efficiency areas of >80%, >85% and >90% are all greater than that of the

single reducer. This is especially the case in the high efficiency areas (>85%

and >90%).

5.6 Conclusions

The test rig was built in UTS from the modification of a former engine

power-train rig. Some calibrations were calculated to match the electric

159

driving requirements. The frame structure was designed and made for the

new power-train installation. 400V DC power was installed to power the

motor.

For the preparation of the DCT rig testing, some testing criterions

were introduced and studied.

One of the most important part of the works is the control system

development. This was developed using the MicroAutoBox of DSPACE.

The control hardware, software and electrical schematics were developed.

Some testing was finished on the rig. This included shifting tests,

temperature tests, driving cycle tests, and efficiency tests. From the testing

results, we can reach the following conclusions:

1. The shift control program can achieve a good shifting of each shift

control. In most cases it can shift smoothly and there is no abnormal

noise;

2. The maximum transmission temperatures in a high-speed drive of 1st

gear and 2nd gear are 71.6 and 104.6 . This does not exceed the

maximum limit of 140 . The maximum temperatures of the bearing

housing after improved bearing lubrication are 56 (1st gear) and 74

(2nd gear). There is no over-temperature phenomenon which indicates an

improved lubrication system;

3. During the driving cycle tests, the shift control works well without

frequent shifting phenomenon. This indicates that the control program

can adapt well under different conditions. Further, the calculated

running range matches the simulation results well and verifies the

validity of the economy shift schedule;

4. From the efficiency testing MAP and statistics table, the efficiency of

the two speed DCT is greater than the single reducer.

160

Chapter 6 Vehicle Integration and Road Testing

6.1 Vehicle Integration

Fig.6- 1 Q60FB Prototype Car

During the vehicle integration the two speed DCT is mounted on the

Q60FB pure electric vehicle, as in Fig.6-1. The engine compartment is very

narrow. In addition, the DCT is a little bit too big and the structure is

complex, so the layout work is difficult. (Fig.6-2)

Fig.6- 2 Q60FB Compartment Left and DCT sample Right

161

The compartment structure and component layout guarantee the

minimum modification requirements in the design process:

1) Define the power-train layout position and angle;

2) Define the motor and gearbox internal and external spline matching

and interface connection scheme;

3) Define the drive shaft and suspension modification scheme;

4) Define the braking, steering, AC, high voltage and low voltage

system layout scheme in the compartment;

Fig.6- 3 Vehicle Layout Scheme

162

Fig.6- 4 Compartment Layout

5) Satisfy DCT’s lubrication, cooling, and electrical interface

requirements, define DCT hydraulic pump position;

6) The motor and motor controller is closed. To shorten the three-phase

line, make the wire layout clearly, this can decrease power loss;

7) Electrical components in the compartment satisfy the fix and

maintenance dismantle request;

Fig.6- 5 Layout of Batteries

Batteries

Fuse Box

Motor Controller

Charger DCDC

12V Battery High Voltage Control Box

163

8) Batteries mounted under the floor and between the front and rear

suspensions. The crash safety should be considered along with the

axle load distribution, maintenance requirements and modification

convenience;

Fig.6- 6 Layout of Charger Port

9) The charger port is mounted at the position of the oil filler;

Fig.6- 7 Installation of the Real Car

Fig.6-7 is the real prototype car installation and working shop;

Advanced DCT technology is integrated with the pure electric drive

system. System efficiency and vehicle performance is optimized along with

the driving ride and comfort. In the process of two speed DCT development

and layout and following the principles of minimum modification and

vehicle performance guarantees, the project development cost and time

period can be decreased.

6.2 On Road Calibration

Real vehicle calibration is an important part of development. From the

164

real road testing and relative parameters adjustment, control performances

can be optimized to adapt to on road driving. Fig. 6-8 is calibration

working on the road; this work is finished in Beijing, on the road near the

BJEV Company.

Fig.6- 8 On Road Testing and Calibration

The main purpose of this thesis is the shifting control of DCT, so

mainly the calibration focuses on the shifting control algorithms. There are

two control methods researched in this thesis, the first traditional control

algorithm is calibrated on the rig, so on road calibration is focused more on

the second motor braking torque control. Fig.6-9 is the shifting curve of the

Power-On Up-Shift.

165

Spee

d(r/m

in)

Fig.6- 9 Power-On Up-Shift Results Motor Status 100-Drive 200-Brake

In Fig.6-9 we can see, from motor active braking control, the shifting

time can be shortened significantly, and speed synchronization can be

finished in 200ms. However, the subjective feeling of driving will be a little

worse; because of the minus torque during the shifting which causes

shaking and worsens the ride comfort.

Spee

d(r/m

in)

Fig.6- 10 Power-Off Up-Shift Results

In the Power-Off Up-Shifting, there is no torque output. The shifting

time is not a key point of the control, so braking control of the motor is not

that important.

From results, we can see that there is a big gap between the shifting

control rig testing results and the simulation results. This is because the

166

simulation model is an ideal model if viewed in relation to the

characteristics of clutch, but not if viewed in relation to hydraulic modes.

The control curves of in-car testing (Fig.6-9) are close to the rig

testing results (Fig.5-25), but there is still some difference between the two.

The rig testing can get smoother results during the shifting. Also the

hydraulic pressure is more difficult to control on the vehicle. We think the

difference between the two situations is inertial. On the rig is lower than the

real vehicle and the driving conditions on the road are tougher than the rig,

so the vehicle testing should be more difficult than the rig. That is why the

rig calibration cannot be instead of in-car calibration totally, but it is a good

reference to shorten the development time.

6.3 On Road Testing

Validation of vehicle performance, dynamic performance and

economic performance has been tested, and the results are displayed in

Table 6-1 and Table 6-2. From Table 6-1 and Table 6-2 we can see that

dynamic performance and economic performance testing results can almost

match the design simulated results.

6.3.1 Dynamic Performance Testing

Table.6- 1 Dynamic Performance Results (Q60EV-DCT)

Items Test Results Simulated Results

Dynamic Performance

Max Speedkm/h

Constant 140 146 Short time 184 187

0 100km/h acceleration time(s) 14.2 14 0 50km/h acceleration time(s) 5.1 4.9

Grade ability(%) 30 31

6.3.2 Economic Performance Testing

Table.6- 2 Economic Performance Results (Q60EV-DCT)

167

Items Test Results Simulated Results Economic

Performance Running Range

km

60km/h 138 146

Energy ConsumptionkWh/ 60km/h

14.7 13.7

6.4 Conclusions

Vehicle integration is an important part of the production development.

How to mount the prototype into the limited cabin space is the key point of

the layout design. Also on road calibration is an important step of the

control system development. In this chapter, the second control method of

the active braking torque control is calibrated on road. From the results, the

shifting time can be shortened significantly, and speed synchronization can

be finished in 200ms. However, the subjective feeling of driving will be a

little worse. Finally dynamic and economic performance tests have been

done on the vehicle. The results match the design requirements well.

168

Chapter 7 Thesis Conclusions

7.1 Summary of the Thesis

This project developed the two speed DCT prototype in cooperation

with UTS and NTC. In total, 3 samples were developed, two of which have

been shipped to China and mounted on one prototype vehicle.

The prototype was developed based on 6 speed DQ250. The technical

difficulties in the course of project development are the system design and

the modification of key modules of the DCT

system (hydraulic module, control module), control program development

and calibration, and rig test methods. In the project DCT theory analysis

and matching has been used to complete the simulation platform building

and the shift transient control theory part analysis. AVL/NTC Company’s

experience of transmission design and modification has been used in the

two speed DCT prototype design and manufacture and this mainly includes

the hydraulic system and mechanical parts. Control system development

occurred in three parts. AVL/NTC developed the shifting control program,

BJEV developed the vehicle control part program and finished the on

vehicle testing control system, and UTS developed the RCP control system

used on the test rig.

Control technology is the core of project development, through the

analysis of theoretical components and the simulation of DCT control. The

DCT control process and control mechanism theory have been analyzed.

Through rig test and vehicle test calibration, the control program

development technology and calibration method have been studied,

Because the electric drive system is different from the traditional

internal combustion engine, the motor can be driven in the more flexible

169

mode of brake and drive control. In this thesis, shifting control through

the driving and braking active speed control have been employed to

synchronize the target speed.

In addition, in the process of development it has been found that the

electric hydraulic pump is one of the key points of project

development, and the electric pump has worked in big current and high

energy consumption. The requirements of noise and durability in the

electric pump selection represent the primary project difficulty. This needs

continued research.

7.2 Summary of Findings and Contributions

In terms of the research of the two speed DCT pure electric drive

system in this thesis, the main contributions are as follows:

The first one is the gear ratio selection and design method of multi-

gear electric drive system:

The traditional gear ratio design method is based on an internal

combustion engine. The engine characteristics curve from the motor

characteristics curve is different. The inertial of the motor should be

smaller than the engine. The gear ratio selection theory of the motor should

be different from the engine. In particular, the gear ratio rate matching

engine cannot be greater than 1.7-1.8, or else the shifting will be difficult.

However, this value of gear ratio match with the motor can be up to 3 or

even greater, and the maximum value needs theoretical analysis. In addition,

the gear box of the internal combustion engine is generally 5 speed or 6

speed, and it can even be up to 8 speed. The maximum speed of the motor

is significantly higher than the engine, so 2speed or 3speed can satisfy the

need of vehicle drive, with the same effect as a traditional 5-6 speed gear

box. Furthermore, in terms of the reverse rotation of the motor, the reverse

170

shaft can be canceled in the electric drive system, so the reverse gear does

not need to design again, and the 1st gear can be used in the reversing drive.

The second contribution is the multi-speed transmission shifting

schedule design method:

The electric system shifting schedule design aims to improve system

efficiency and extend running range, so the shifting schedule design is

based on the drive line efficiency MAP. This is the same as the engine’s

shifting schedule design from the oil consumption MAP. This thesis found

the electric drive system economic shifting schedule design method was

just the electric drive system shifting schedule design method. Because the

transient characteristics of the motor are not that obvious, the two

parameters shifting method can satisfy the general shifting requirement.

But in consideration of the temperature influence of the electric system,

and the instability of the gear box transmission of the electric drive system,

the introduction of the temperature into the shifting is the next step.

The third contribution is the simulation method.

Generally University and research institutions adapt Matlab/Simulink

to build the system and vehicle simulation platform, and this needs

complex model building work. Companies always like to use mature

commercial simulation software, such as Crusie, Advisor etc. to do the

matching and control system development. In this thesis, commercial

software of AMESim was introduced to build the steady state simulation

platform, meanwhile the co-simulation method with Simulink was studied

and the control algorithm was developed in the simulink environment.

However, for the transient shifting control program development, Simulink

models were built to simulate the transient characteristics of the power-

train. But these two simulation platforms need more work for integration

purposes.

The fourth contribution is the pure electric multi speed transmission

171

shifting control method research.

Two of the shifting control methods were studied in this thesis. The

first was the traditional shifting control method; and the second was the use

of the motor active braking torque control to shorten the motor speed

synchronous time. The results show that the shifting time can be shortened

significantly, but the ride comfort will be worsened because of the torque

hole during the motor negative braking torque in the inertial phase.

Through shifting control, we can find that if the motor inertial is lower

than the engine, this changes the dynamics of the speed phase of the shift.

Additional motor control and the torque handover algorithm is required for

fully optimized control. This may be a solution to the instability at the end

of the shift.

The fifth contribution is the rig testing and calibration method.

There has been no multi-speed gearbox testing criterion of electric

power-train system in recent years. In this research, through the

introduction of manual transmission and automatic transmission, and the

motor testing method, the two speed DCT was tested on the rig. This

included the shifting testing and calibration, temperature testing, and

driving cycle simulation testing of NEDC and UDDS.

The sixth contribution is the concept of multi-speed transmission

structure and exploration.

During the research of the two speed DCT in this project development,

some other kinds of multi-speed electric drive systems were also studied,

such as the two motor multi-speed system, and multi speed electric system

based on the CVT. In terms of the multi speed of the AMT system for the

torque interruption during shifting, the ride comfort will be worse. The

multi speed electric drive system based on AT or CVT for the complexity

of structure, and the system efficiency is a little bit low. The electric power-

train used in the DCT can get good system efficiency and shifting comfort

172

but the control complexity will increase. The key point is that the modules

(hydraulic module and dual clutch module) are the main challenge of the

industrialization.

In the two motor multi speed system through the means of the

introduction of the second motor into the system, the same function of the

dual clutch’s non-torque interruption can be realized through two motor

torque controls. However, the cost and control should be the research points.

Overall, the multi-speed electric power-train is much more flexible

than the multi speed engine system so a bigger design space in terms of

structure and control needs to be explored.

7.3 Limitations to Research

Up to now, the project has achieved two speed DCT prototype

development and on vehicle testing. This has been initially realized in

the control function but there are still many shortcomings which need to be

further improved. The main shortcomings are as follows:

The first one is the shifting schedule research is only considered from

the steady state, and more works concerning the transient state should be

introduced into the research. Furthermore, only two parameters have been

studied in terms of the schedule design, and more parameters such as

battery characteristics and the gearbox temperature should be considered.

The second one is the shifting control algorithm. This still needs to be

explored in more control methods, such as:

Torque instability at the end of the shift:

This is due to motor control, although further optimization torque

handover may also achieve significant improvements. If during the speed

phase of the torque handover a speed control is used this may be able to

resolve the issue.

173

Solenoid driving:

The solenoid control is highly non-optimized. Considerable

improvement can be achieved with a better driver circuit and control.

This issue created inconsistent shifting.

The third one is the transient control theory research which still needs

more exploration. In this thesis, up-shifting control has been researched,

but the shift during braking has not been considered. So how to use this

platform to extend the transient research, make comparisons in terms of the

theory and testing results, and direct the multi gear electric drive system

(and even the hybrid system design and control) are worth further research.

This will help to push the development of electric power-train and relative

industrialization.

The fourth one is the multi speed electric structure analysis is not

enough. The theoretical support is still limited and it requires more

exploration for it to work, such as different transmission being used in

electric power-train, etc.

7.4 Future Research

In Chapter Two, the two motor multi speed system was analyzed, and

the comparison with the single reducer system was outlined. This provided

a power split system design concept. It is one of the key optimization

methods. Another method is the optimization of the existing single motor

system. From the two speed DCT we can conclude that to improve the

efficiency of the electric power-train, this should be done mainly through

the improvement of the motor working points to reduce system loss and

increase system efficiency. Another method is through a decrease in the

accessory system. The following structure is from these two methods to

improve the system efficiency; and it is important for further research.

174

The two speed electric power-train based on the mechanical pump was

provided; and a detailed structure is in Fig.7-1.

1st gear

2nd gear

MechanicalPump

Differential

Input shaftOutputShaft

Main Gear

Clutch

One way bearingclutch

Torsionalvibrationdamper

Clutch Actuator

Fig.7- 1 Two Speed Electric Power-train used Mechanical Pump

The two speed electric power-train system used mechanical pump

mainly includes a torsion vibration damper, 1st gear and 2nd gear, one-way

bearing clutch, shifting clutch, mechanical pump, main gear and differential.

The output shaft of the motor connects to transmission input shaft

through the torsion vibration damper. The function of the damper is to

reduce torque vibration and guarantee ride comfort. The 1st gear connects to

the input shaft through a one-way clutch. There is a shifting clutch between

the 1st gear and 2nd gear. The engagement of the clutch is the shifting

process. The mechanical pump is fixed on the outside of the input shaft,

and it is driven by the motor.

The innovations of this structure are:

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The one way clutch and shifting clutch are used to realize the shifting

control, and it is the sample. During the shifting, we only need to control

the engagement or disengagement of the shifting clutch. The one way

clutch automatically disengages because of the difference in the shaft speed.

It is easy to control.

The one-way clutch structure of the 1st gear makes it possible to use

the mechanical pump. It can be simple and cost saving compared to the

general automatic transmission electric pump.

The shifting clutch is often open at the initial state, and this makes it

possible to start the vehicle without pump pressure. When motor speed

reaches a given value, the mechanical pump pressure can be the set value,

and the shifting can be done. This makes it possible for the system energy

saving in the general condition.