braking with a directional control valve in a hydraulic

Braking with a Directional Control Valve in a Hydraulic Open-Loop Transmission

Authors: Jonathan Karlborg

Emil Sten

Examiner: Senior Lecturer Magnus Sethson

Linköping University, IEI

Supervisors: Dr. Björn Eriksson

Parker Hannifin, GMS

Ing. Andries Broekx

Parker Hannifin, GMS

Senior Lecturer Robert Braun

Linköping University, IEI

Linköping University

SE-581 83 Linköping, Sverige 013-28 10 00, www.liu.se

Linköping University | Department of Fluid and Mechatronic Systems Master Thesis 30 HP | Education - Mechanical Engineering

Spring term 2021 | LIU-IEI-TEK-A--21/04008—SE

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[Intentionally left blank]

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Abstract This project presents an investigation if natural braking characteristics could be achieved on an open-

loop hydraulic transmission without a brake valve. The goal with the simplified system was to utilize the

directional control valve to achieve similar functionality as the brake valve does in the conventional

system. If the solution functions properly, it will reduce costs, save time and simplify the conventional

system which uses a dedicated brake valve.

With a simulation model and practical experiments, the simplified system was thoroughly studied and

tested. Two different concepts of how to control the directional control valve were developed, named

Fixed Control and Torque Control. The Fixed Control concept has a predetermined de-stroking profile

which is not affected by other system signals other than the gas pedal signal. The Torque Control

concept uses in addition to the gas pedal signal, pressure sensors over the motors to maintain a

constant braking torque.

Both developed concepts were able to produce natural braking characteristics. However, the Torque

Control concept performed better at different circumstances. Respective concept can be tuned further

for improvements, but in the scope of this project the goal was accomplished.

Utilizing the directional control valve to achieve a hydraulic brake function, has potential to be a solution

for heavy mobile machinery in near future. However, further research and testing are required to be

conducted on other heavy mobile machinery which have greater top speeds and load capabilities than

the machinery used in this project.

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Acknowledgements This report presents a Master thesis which was conducted as a last assignment for the Mechanical

Engineering program at the Division of Fluid and Mechatronic Systems (Flumes) at Linköping University.

The thesis was conducted during the spring of 2021 and the project was conducted on behalf of the

corporation Parker Hannifin.

From Parker Hannifin we would like to thank our manager Erik Forsberg and our supervisors Andries

Broekx and Björn Eriksson for the tremendous amount of support, valuable knowledge and discussions

along the project. Also, a big thank you to all the co-workers at Parker Hannifin for the warm welcome

and support. From Linköping University, we thank our supervisor Robert Braun and for also providing

us with help along the project and our examiner Magnus Sethson.

Borås, June 2021

Jonathan Karlborg & Emil Sten

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Nomenclature Symbol Description Unit

𝑎 Acceleration [m/s2]

𝛼 Inclination angle [°]

𝐴 Spool area opening [m2]

𝐵𝐹𝐷 Friction parameter in gearbox and drivetrain (FD) [Nms/rad]

𝐵𝑚 Internal viscous friction parameter in the hydraulic motor [Nms/rad]

𝐶𝑚 Motor leakage coefficient [m3/(sPa)]

𝑐𝐷 Drag coefficient [-]

𝐶𝑞 Flow coefficient [-]

𝑐𝑅 Rolling coefficient [-]

𝑐𝑤 Wheel circumference [m]

𝐷𝑚 Displacement setting of the motor [m3/rad]

𝐷𝑝 Displacement setting of the pump [m3/rad]

𝛿 Damping coefficient (Second order low pass filter) [-]

𝐹𝐷 Aerodynamic drag [N]

𝐹𝑖𝑛𝑐 Inclination force [N]

𝐹𝑅 Rolling resistance [N]

𝐹𝑡 Traction force [N]

𝑔 Gravitational constant (~9.81) [m/s2]

𝐼 Inertia [kgm2]

𝑗 Jerk [m/s3] 𝑚 Wheel loader mass [kg]

𝑛𝑚 Rotational speed of the hydraulic motor [rad/s]

𝜂ℎ𝑚 Hydro-mechanical efficiency [-]

𝜂𝑣 Volumetric efficiency [-]

𝜌𝑜𝑖𝑙 Oil density [kg/m3]

𝑝 Pressure [Pa]

𝑞 Volumetric flow rate [m3/s]

𝜏 Torque [Nm]

𝑣 Velocity [km/h]

𝜔 Angular velocity [rad/s]

�̇� Angular acceleration [rad/s2]

𝜔𝑏 Break frequency (Second order low pass filter) [rad/s]

𝑥𝑎𝑐𝑡𝑢𝑎𝑙 Spool’s actual axial position [m]

𝑥𝑟𝑒𝑓 Spool’s reference value [m]

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Acronyms CAN Controller Area Network

CVT Continuous Variable Transmission

DCV Directional Control Valve

ECU Electronic Control Unit

FBD Free Body Diagram

FD Final Drive

GMS Global Mobile Systems

ICE Internal Combustion Engine

K220LS Load Sensing DCV installed in the Zettelmeyer.

LS Load Sensing

P2075 Load Sensing pump installed in the Zettelmeyer.

PPRV Proportional Pressure Reducing Valve

TLM Transmission Line Modelling

V12-80 Hydraulic motor installed in the Zettelmeyer.

V14-110 Hydraulic motor installed in the Zettelmeyer.

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Table of Contents 1 Introduction ................................................................................................................................... 1

1.1 Background ........................................................................................................................... 1 1.2 Goal ....................................................................................................................................... 2 1.3 Purpose ................................................................................................................................. 2 1.4 Delimitations .......................................................................................................................... 2 1.5 Methodology .......................................................................................................................... 3 1.6 Corporation Description ......................................................................................................... 3

2 Theory .......................................................................................................................................... 5

2.1 Wheel Loader ........................................................................................................................ 5

2.1.1 Brake Characteristics ...................................................................................................... 5

2.2 System Description ................................................................................................................ 5 2.3 Directional Control Valve (DCV) ............................................................................................. 6

2.3.1 DCV Spool ...................................................................................................................... 7 2.3.2 Solenoid ......................................................................................................................... 9

2.4 Pressure Compensator .......................................................................................................... 9 2.5 Hydraulic Continuous Variable Transmission (CVT) ............................................................ 10

2.5.1 Open- and Closed-Loop ................................................................................................ 11

2.6 Pump (Axial In-line Piston Machine) .................................................................................... 11 2.7 Motor (Bent-axis piston machine)......................................................................................... 13 2.8 Pressure Relief Valve .......................................................................................................... 15 2.9 Cavitation ............................................................................................................................. 15 2.10 Check Valve (Anti-Cavitation Valve) .................................................................................... 15 2.11 Brake Valve ......................................................................................................................... 16 2.12 Road Vehicle Dynamics ....................................................................................................... 18

2.12.1 Resistive Forces ........................................................................................................... 19 2.12.2 Moment of Inertia .......................................................................................................... 20

2.13 Hopsan ................................................................................................................................ 21 2.14 IQAN .................................................................................................................................... 21

3 Implementation ........................................................................................................................... 23

3.1 Simulation Model ................................................................................................................. 23

3.1.1 Validation of the Pump .................................................................................................. 25 3.1.2 Validation of Pressure Compensator ............................................................................. 25 3.1.3 Validation of the DCV.................................................................................................... 25 3.1.4 Validation of the Hydraulic Motor .................................................................................. 26 3.1.5 Validation of the Rotational Gears................................................................................. 27 3.1.6 Validation of the Wheel Loader ..................................................................................... 28 3.1.7 Validation of Hydraulic Volumes ................................................................................... 29 3.1.8 DCV Spool De-stroking Logic ....................................................................................... 30

3.2 Practical Experiments .......................................................................................................... 32

3.2.1 Test Approach .............................................................................................................. 32 3.2.2 Test Description ............................................................................................................ 32 3.2.3 Cavitation...................................................................................................................... 33 3.2.4 IQAN Implementation and Modelling of ECUs .............................................................. 33

3.3 Evaluation of Braking Characteristics ................................................................................... 35

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4 Results ........................................................................................................................................ 37

4.1 Simulation Results ............................................................................................................... 37 4.2 Practical Results .................................................................................................................. 40

4.2.1 Braking Time Delay ...................................................................................................... 43

4.3 Cost Analysis ....................................................................................................................... 43

5 Discussion .................................................................................................................................. 45

5.1 Results Discussion .............................................................................................................. 45 5.2 Cost Analysis Discussion ..................................................................................................... 46 5.3 Research Questions ............................................................................................................ 46 5.4 Validation of Simulation Model ............................................................................................. 47 5.5 Fixed Control vs. Torque Control ......................................................................................... 48 5.6 Safety .................................................................................................................................. 48 5.7 Sources of Error ................................................................................................................... 48 5.8 Future Work ......................................................................................................................... 49

5.8.1 Displacement Control.................................................................................................... 49 5.8.2 Torque Control for Different Speeds ............................................................................. 49 5.8.3 Fixed Control for Different Speeds ................................................................................ 49 5.8.4 Extreme Cases ............................................................................................................. 49 5.8.5 Heavy Mobile Machinery ............................................................................................... 49 5.8.6 Components and Configurations ................................................................................... 50

6 Conclusion .................................................................................................................................. 51 References ........................................................................................................................................ 52 Appendices ........................................................................................................................................ 54 Appendix A – Symbols ....................................................................................................................... 55 Appendix B – Recalculated Meter-in Area Opening ........................................................................... 57 Appendix C – Derivation of the Wheel’s Moment of Inertia, 𝐼𝑤 ........................................................... 58 Appendix C – Derivation of Motor Internal Viscous Fricton ................................................................. 60 Appendix D – Derivation of Motor Internal Leakage Coefficient.......................................................... 62

1 INTRODUCTION

1

1 Introduction Systems with motion control technologies are today generally operated by hydraulic or electric

solutions. Whether how the motion is controlled there is always a strive to manufacture systems

more simplified and efficient with new solutions. Cost savings and environmental sustainability

are two relevant reasons for this, especially the latter. This report presents an investigation of a

new hydraulic solution in a heavy mobile machinery system.

1.1 Background

Illustrated in Figure 1.1 is a Zettelmeyer wheel loader which was modified for this project to

operate in an unconventional manner. Initially the wheel loader had a closed-loop hydraulic

transmission, but with the new modification the hydraulic transmission is open-loop. The main

difference as well as the pros and cons between open- and closed-loop systems are further

explained in chapter 2 Theory.

Figure 1.1: Zettelmeyer ZL 802 Si, shortly referred as Zettelmeyer. Figure source: [1].

The main components of a conventional open-loop transmission are a pump, motor, a directional

control valve (DCV) and a brake valve. What made the modification on the Zettelmeyer

unconventional in this project, was the removal of the brake valve. The main purpose of the brake

valve is to hydraulically brake the vehicle if the gas pedal is released by the operator. An analogy

of the hydraulic brake function is engine braking on a conventional car. A brake pedal also exist

which uses mechanical friction to decelerate the vehicle but only the hydraulic braking function

was studied in this project. The drawback of a brake valve is that it requires a lot of tuning before

it performs as required. The tuning process consists of a lot of iterations and is a time demanding

process which is one reason why the brake valve was wished to be removed [2]. Another reason

for removing the brake valve is cost. Removing the brake valve would mean one less component

in the system which would be cheaper. The challenge though, is to achieve the same performance

as prior to the removal. There is also a risk that cavitation may occur in the hydraulic circuit if the

brake valve is removed. If required performance can be achieved without the brake valve, there

is money to be saved. Weight reduction, space savings and system simplification are additional

minor benefits with the removal process.

1 INTRODUCTION

2

To achieve the hydraulic braking without the brake valve, the DCV will be utilized instead. It will

be electronically controlled to regulate the braking. The challenge with the project is to model the

software which will result in a comfortable braking motion while also avoiding cavitation.

1.2 Goal

The goal with the project was to achieve braking characteristics with a natural feeling when the

wheel loader hydraulically brakes, without the brake valve mounted in the system. This was to be

achieved by implementing software code which would regulate the DCV which consequently

would regulate the speed of the wheel loader. Cavitation in the hydraulic circuit was also to be

avoided. The only quantification goal which was set was that the wheel loader should decelerate

from maximum speed to 0 km/h in approximately five seconds.

1.3 Purpose

The purpose with the thesis was to investigate the possibility to hydraulically brake the wheel

loader with the directional control valve and to test if desired performance could be acquired. The

project can be translated into the following research questions with the removal of the brake valve:

• RQ1: Can braking characteristics with a natural feeling be achieved with the DCV?

• RQ2: Is it possible to avoid cavitation when braking with the DCV?

• RQ3: Can the same performance be acquired using the DCV as with the brake valve?

• RQ4: Can a realistic simulation model of the system be created?

1.4 Delimitations

In this project several delimitations were established, which are listed below. This was because

the project had a limited time span of approximately five months to be completed in.

• Only braking characteristics on flat planes has been considered. Decent braking

characteristics on inclined planes is only considered to be an additional benefit.

• Hydraulic working attachment on the wheel loader was assumed to be idle during braking.

The installed pump was undersized to provide sufficient flow to both the transmission and

the working attachment concurrently, which resulted in instabilities.

• Regulation of the hydraulic motor displacement affects the braking characteristics and will

require a more complicated solution. It was therefore delimited to keep the displacement

constant during braking.

• No investigations were conducted on the conventional open-loop system with a brake

valve. This was because no brake valve was available at the location.

• The aerodynamic drag was not considered in the project. This was because the wheel

loader only reached a top speed of ~12 km/h which resulted in marginal effects on the

system.

1 INTRODUCTION

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1.5 Methodology

The following methods were used throughout the project.

• Literature study

• Modelling & Simulation

• Practical experiments

To obtain knowledge regarding the system and its components, a literate study was conducted.

In parallel was a simulation model of the wheel loader’s hydraulic system created in the software

Hopsan. Concepts of how to control the system was developed with aid of the simulation model.

Presented in chapter 3 Implementation is a more detailed description of the conducted work.

1.6 Corporation Description

The project was conducted at the corporation Parker Hannifin, which generally is just referred as

Parker. The corporation was founded by Arthur Parker in 1917 in Ohio, USA. Today Parker offers

expertise in motion control technologies and are active in business areas such as: hydraulics,

electromechanics, climate control, aerospace and many more. As of 2021 the corporation had

approximately 50 000 employees worldwide. [3]

Parker have been present in Sweden for almost 40 years as of 2021 and the project was

conducted at Parker’s global mobile system (GMS) division at Borås, Sweden. GMS are mainly

in charge of developing and evaluating hydraulic and electric solutions for heavy mobile

machinery systems, such as wheel loaders and excavators.

1 INTRODUCTION

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2 THEORY

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2 Theory In the following section, topics within the scope of the project will be explained closer.

2.1 Wheel Loader

Wheel loaders classifies as heavy mobile machinery which is a diverse machine with multiple

uses. One of the main uses for wheel loaders are construction work, where handling gravel, soil

and other similar mediums are managed with an attached bucket. It is also commonly used for

loading, by attaching a fork instead of a bucket the wheel loader will become a powerful and fast

asset for unloading and loading heavy pallets. In Figure 2.1 the Zettelmeyer is shown with two

different hydraulic working attachments.

Figure 2.1: Zettelmeyer with different working attachments, A) Loading fork, B) Bucket. Figure source: [1].

Wheel loaders are commonly powered by an internal combustion engine (ICE). As for the Zettelmeyer, a diesel engine powers the hydraulic pump which drives and regulates the hydraulic system. The hydraulic system in the wheel loader is used for two applications, controlling the hydraulic working attachment such as the boom and bucket movement, as well as the hydraulic transmission. This means that the hydraulic pump is running a hydraulic motor which is connected to the drive shaft of the wheel loader. [2]

2.1.1 Brake Characteristics

Wheel loaders are versatile machines that are used in different applications. The braking

characteristics can therefore not have a universal solution and requires to be modeled by

preference. For example, a wheel loader working with cargo at a loading dock may need different

characteristics than a wheel loader working at an excavation site or at other locations with steep

slopes. The general solution to the problem is to switch between different modes implemented in

the wheel loader. For example, a mode for freewheeling when no braking is required, another

mode for aggressive braking and also a gentle braking mode. These braking modes are to be

achieved with the hydraulic brake function and not by the mechanical brake. The mechanical

brake is generally only used for emergencies. [2]

2.2 System Description

In Figure 2.2 a system overview of the hydraulic transmission is illustrated. In appendix A, a list

of symbols and names are presented for the hydraulic components. The system schematic is

simplified but for the purpose of this project it does not have to be more complicated. In the

following paragraphs components, phenomena and other relevant information will be explained

to acquire an understanding of the system.

2 THEORY

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Figure 2.2: Simplified system schematic of the hydraulic transmission on the Zettelmeyer.

2.3 Directional Control Valve (DCV)

The DCV is a central component in a hydraulic system. With the help of a spool located within the

DCV the flow can be controlled by stroking (opening) or de-stroking (closing) the spool in axial

direction. Flow paths will open and close to enable the hydraulic system to function as intended.

With the gas pedal and shifter command, software logic will translate the positioning of the spool.

In Figure 2.3 a cross section view of a K220LS is illustrated, which is the DCV installed in the

Zettelmeyer. The illustration is missing components such as springs on each side of the spool,

pressure relief valves, check valves and the pressure compensator. These components are

described in the upcoming sections.

Figure 2.3: DCV, K220LS, in closed position.

The K220LS is a closed centered 4/3 DCV which means four ports are connected and the spool

can be positioned in between three modes to direct the flow. It is proportionally controlled,

meaning that the opening can be continuously regulated in contrast to on/off valves which are

either fully opened, or fully closed. The DCV has a load sensing (LS) function which is connected

to the pump which regulates the flow. In Figure 2.3 the spool is in centered position which results

in a closed DCV, meaning no flow through the DCV would occur (except radial leakage).

2 THEORY

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In Figure 2.4 the spool is positioned so that the valve is fully opened and fluid would flow through

the DCV according to the figure. Moving the spool in opposite direction will change the flow path

and workport B will be delivered with high-pressure fluid from the pump. Fluid transmitted either

from (P–A) or (P–B) is referred to as meter-in. On the contrary, flow being transmitted from (A–T)

or (B–T) is referred to as meter-out.

Figure 2.4: DCV, K220LS, fully opened.

2.3.1 DCV Spool

The spool is the component in the DCV valve that regulates and controls the flow. In Figure 2.5 a

close up on the spool notches are shown. The notches can have various shapes to affect the

opening area depending on the spool’s axial position. The notches in this case have a cylindrical

shape which will affect the area opening curve. For this case, 50% stroke will not be equivalent

to 50% opening area. The notches are designed differently between applications to enable

different flow characteristics. [4]

Figure 2.5: Close up illustration of the spool.

In addition to that, the number of notches circulating the spool’s outer circumference can vary [4].

In this case the spool has four notches which is illustrated in Figure 2.6A. Further are two other

spool designs shown. It is evident that the angles between the notches are symmetrically

designed, which is to prevent rotational motion to occur because of fluid forces acting on the

spool.

Notch

2 THEORY

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Figure 2.6: Different designs. A) Four notches 90° apart, B) Three notches, 120° apart, C) Two notches, 180° apart.

The flow through the notches or an orifice can theoretically be described with equation 2.1.

Varying the spool stroke will vary the opening area which is referred to as metering.

𝑞 = 𝐶𝑞 ⋅ 𝐴 ⋅ √2

𝜌𝑜𝑖𝑙⋅ 𝛥𝑝 (2.1)

where

• 𝑞 = Volumetric flow rate.

• 𝐶𝑞 = Flow coefficient (generally assumed to be 0.67).

• 𝐴 = Opening area, depending on the spool stroke.

• 𝜌𝑜𝑖𝑙 = Oil density.

• Δ𝑝 = Pressure difference from the inlet to the outlet

Spools are designed to be either overlapped, underlapped or line-to-line (zero overlapped). The

difference between the designs are shown in Figure 2.7. The spool used in this project was

overlapped which means a deadband had to be considered. Deadband is the amount of spool

stroke required before the opening area exceeds 0 mm2, which allows flow through the valve. [5]

Figure 2.7: Overlap, underlap and line-to-line spool designs.

2 THEORY

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2.3.2 Solenoid

The DCV in previous Figures are simplified. It has been depicted as it is directly electronically

piloted when in reality it is pressure piloted. Respective side of the DCV spool are connected to a

proportional pressure reducing valve (PPRV). In the PPRV an armature is controlled electronically

with a pilot operated solenoid. The solenoid consists of a tightly wounded coil that creates a

magnetic field when electric current is flowing through. The magnetic field that is created will move

the armature within the coil. The armature will in turn guide flow and pressure to the DCV spool

which will force it to desired position. In Figure 2.8 the two PPRVs are illustrated where the tank

line is denoted with a T and the pressure source with a P. Two PPRVs are mounted to control the

pressure on each side of the spool. When neither solenoid is energized and the pressure drops,

the DCV spool returns to a centered position because of springs mounted on respective side.

Figure 2.8: Schematic of the 4/3 DCV pressure piloted by two PPRVs.

2.4 Pressure Compensator

The pressure compensator is a component that assists to remain a constant rotational speed of

hydraulic motors during operation. This is achieved by sustaining a constant pressure drop over

the DCV regardless of the load magnitude. The principle of a pressure compensator is illustrated

in Figure 2.9.

Figure 2.9: Schematic of a simplified pressure compensator. Figure source: [6].

2 THEORY

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The pressure compensator is affected by the intermediate pressure and the load pressure. The

supply pressure from the pump, does not regulate the compensator. The supply only needs to be

sufficient to supply the load with enough pressure and flow. The mounted spring, the load and the

intermediate pressure will decide the positioning of the compensator spool. The load pressure is

acting on the compensator spool with an opening force. Higher load pressure will open the

compensator spool further. Meanwhile the intermediate pressure counteracts with an opposite

force. The pressure difference between the intermediate- and load pressure is determined by the

compensator spring force. The spring will keep the pressure drop over the main spool constant

at various loads. [6]

The benefits with a pressure compensator are that two different motors or cylinders can work with

different loads and still maintain constant speed. For example, if one motor requires 200 bar and

the other 100 bar the pump would have to supply at least 200 bar. Without the pressure

compensator the motor with the lower load would be supplied with 200 bar. That would increase

the pressure drop and thus the flow which would increase the speed of that motor. With a pressure

compensator that would not occur. The compensator spool would regulate the pressure drop over

the main spool and keep the motor at a stable level. [6] The speed would instead be regulated by

increasing or decreasing the opening area of the variable orifice (DCV spool).

2.5 Hydraulic Continuous Variable Transmission (CVT)

A transmission is a system that transfers mechanical rotating power from the engine to the wheels.

In cars the transmission is generally achieved with a gearbox which has multiple gears with

different gear ratios to accommodate the required torque and vehicle speed. Wheel loaders on

the other hand, uses continuously variable transmission (CVT). In Figure 2.10 the characteristics

of a geared and a continuous variable transmission are shown.

Figure 2.10: Geared transmission vs. CVT.

The gear ratio regulates torque and speed, at low speeds greater torque is required to accelerate

the vehicle. For a geared transmission the driving gear is smaller than the driven gear. This will

generate a slower rotation on the wheel but with a greater torque [7]. For CVT the driving gear

and driven gear is switched out to a variable displacement pump and motor. The motor which is

the driven component in this case can vary the displacement. Increased displacement will

increase the torque output and decreased displacement will reduce torque but enable increased

speed. Since the displacement is continuous variable for the pump and motor the transmission

will be continuous.

2 THEORY

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2.5.1 Open- and Closed-Loop

For hydraulic systems there are two different categories, namely the open-loop system and the

closed-loop system. In the open-loop system, the pump’s inlet flow is supplied from the tank and

the motor outlet flow is discharged back to tank. The closed-loop however is not connected to

tank. The pump’s inlet flow is directly connected from the motor outlet making it a closed circuit

between the two components. In Figure 2.11 the difference between the two systems are

illustrated. [8]

Figure 2.11: Simplified schematics of open- and closed-loop systems where main components only are considered.

The two different system types come with different advantages and disadvantages. The open-

loop is a simpler, cheaper, easier to maintain and more compact system than the closed-loop.

With a valve controlled open-loop system, which is the most common one, there is also the

possibility to control multiple loads, motors, or cylinders at the same time. The closed-loop on the

other hand has direct feedback and is more accurate than the open-loop.

2.6 Pump (Axial In-line Piston Machine)

A hydraulic pump is a component that converts mechanical rotational power into hydraulic energy.

Hydraulic energy is stored in the form of pressurized fluid. The axial in-line piston machine which

is used in the Zettelmeyer transmission will be the focus in this section. The pump machine is

illustrated in Figure 2.12 and it operates by the help of multiple pistons that rotate to generate a

pressure.

Figure 2.12: P2075, axial in-line piston pump. Figure source: [9].

2 THEORY

12

In Figure 2.13 is a simplified schematic of an axial in-line piston pump. The drive shaft is rotated

by an ICE which supplies mechanical rotational power to the pump. The drive shaft rotates the

cylinder barrel which consequently rotate the pistons. The rotation will generate hydraulic energy

with the assistance of the swash plate. The swash plate is a static plate that is connected to the

pistons via slippers. The slippers enable the pistons to rotate along the swash plate plane through

sliding on a hydrostatic thin oil film without any mechanical connections. The swash plate is tilted

at an arbitrary angle which will cause the piston to be in a retracted position at the inlet kidney

and in an extended position while at the outlet kidney. The angle of the swash plate determines

the displacement setting of the pump which is associated how much flow will be delivered to the

hydraulic circuit.

Figure 2.13: Simplified schematic of an axial in-line piston pump.

In Figure 2.14 is an example of a pump with maximum (A) and zero displacement (B) setting

illustrated. Maximum displacement will deliver maximum flow while zero displacement will deliver

zero flow to the hydraulic circuit.

Figure 2.14: Various displacement settings, A) maximum displacement, B) zero displacement.

Pumps can either have a fixed or a variable displacement. To change the displacement of an

axial piston pump the swash plate angle can be adjusted. Increasing the angle from a vertical to

a tilted setting, such as Figure A, will in turn increase the displacement. This is due to the

increased distance the piston will have to travel within the cylinder. Pumps with variable

displacement are in many cases load sensing pump. This means that the pump can sense the

load that is being applied to the system and vary the output flow and pressure to accommodate

the demand. This helps the pump save energy since it only performs work on demand. [10]

2 THEORY

13

When a piston retracts in the cylinder suction will occur and fluid will flow from the tank into the

piston cavity inside the cylinder barrel. As the cylinder barrel rotates the piston extends due to the

swash plate angle which will increase the pressure of the fluid in the cavity. When the cavity

reaches the outlet kidney fluid will be discharged out in the high-pressure hydraulic system. The

suction and discharge functions happen simultaneously because of multiple pistons rotating in

the cylinder barrel. [8]

The valve plate is a component that separates the inlet and outlet port. The two kidneys in the

valve plate connects the pistons with the two ports and enable flow. The pressure relief grooves

are a feature that reduce the noise from the pump. The grooves equalize the pressure between

the cylinder and the port. [8]

The flow delivered by a pump can theoretically be calculated with equation 2.2.

𝑞𝑝 = 𝐷𝑝 ⋅ 𝑛𝑝 ⋅

1

𝜂𝑣 (2.2)

where 𝐷𝑝 is the displacement of the pump, 𝑛𝑝 is the rotational speed of the pump and 𝜂𝑣 is the

volumetric efficiency.

2.7 Motor (Bent-axis piston machine)

Hydraulic motors work the same way as pumps .The difference between the two however is that

the motor transfers hydraulic energy into mechanical power and a pump the opposite as explained

in the section above. For motor applications it is more common to use a bent-axis machine. The

two bent-axis motors with variable displacement shown in Figure 2.15 is what was installed in the

Zettelmeyer during the project. [8]

Figure 2.15: V12-80 and V14-110 motors installed in the Zettelmeyer. Figure source: [11].

The difference between the in-line and bent-axis is that the cylinder barrel is at an angle to the

output shaft. The angle can be varied to increase or decrease the displacement. In Figure 2.16 is

a simplified schematic of the machine with maximum displacement illustrated. If the cylinder barrel

would be lowered it would result in a decreased displacement.

2 THEORY

14

Figure 2.16: Simplified schematic of a bent-axis variable displacement machine.

A bent-axis machine can achieve greater angles, up to ~45°compared to a ~21° swash plate angle

in an in-line machine. Greater angles benefit the machine in terms of efficiency and power density.

It allows the machine to have a higher torque due to the high displacement which is good in motor

operations. However, the bent-axis machine has a few downsides which is cost and complexity.

The shaft bearings in a bent-axis machine experience high loading and are therefore more prone

to premature failure compared to the axial in-line machine [12]. [8]

During operation, the ICE provides a torque to the pump shaft, 𝜏𝑝, which consequently makes the

fluid flow according to Figure 2.17. The fluid flows through the parallel motors which consequently

converts hydraulic energy into torque, 𝜏𝑚, on a rotating shaft. The shaft is connected to a final

drive (FD) where a gearbox will split and amplify the torque to the wheels. In addition, the

longitudinal rotation is translated into a transverse rotation. On the wheels there is a resistive

torque present, 𝜏𝑟, which is due to resistive forces such as rolling resistance and the force induced

by the inclination. This will be presented further in section 2.12 Road Vehicle Dynamics.

Figure 2.17: Torque sources acting on different shafts during operation.

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2.8 Pressure Relief Valve

The pressure relief valve is the component that sets the maximum pressure for a system or

section. During regular operation the pressure relief valve is closed and has no function. However,

when the system experiences high exceeding pressures the pressure relief valve will open. It

opens at a pre-set value and directs the high-pressure fluid to tank. The valve will protect

components from breaking and most importantly hoses from busting which could endanger

operator safety. [13] The two pressure relief valves referred to in Figure 2.18 were installed to

open whenever the pressure in the hydraulic circuit exceeded 330 bar. [2]

Figure 2.18: Pressure relief valves installed in the Zettelmeyer.

2.9 Cavitation

Cavitation is the phenomenon where vapors are formed and collapsed due to reduced pressure.

It most commonly occurs at the pump inlet in a hydraulic system. From section 2.6 Pump (Axial

In-line Piston Machine), it is explained that the pump sucks fluid and then discharges it out into

the system. It is critical that the flow is sufficient and not restricted from the tank to the pump. If

the flow is restricted or the pump is not receiving enough flow the risk for cavitation increases.

Incorrect hose dimensions, high oil viscosity, poor fittings and valves that are not fully opened are

some examples of flow restricting elements. [14]

The reason cavitation will appear with insufficient flow is because the pump will continue to

operate and try to suck fluid. When the flow is limited, however, the pressure will rapidly decrease

causing the fluid to vaporize. The vaporization of the fluid will result in small vapor bubbles being

formed. Shortly after when the vapor bubbles are discharged into the high-pressure system they

will collapse due to the increased pressure. The collapse of the bubbles can be explained as a

tiny explosion. These explosions create vibrations, noise and increases the wear on the pump. It

can also affect the efficiency of the pump in a negative way.

2.10 Check Valve (Anti-Cavitation Valve)

Check valves are simple hydraulic components with one function but many applications. Check

valves are designed to let fluid flow freely in one direction and be stopped in the other direction.

It is commonly implemented in pipelines to prevent backflow to pump or to keep the system from

draining while turned off. In the hydraulic transmission studied in this project, the check valves

functioned as anti-cavitation valves. In Figure 2.19 is an illustration how the check valves were

implemented and how they operate during a braking sequence. [15]

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Figure 2.19: Flow path through check valves during braking.

During acceleration and cruise, the DCV is fully opened and the risk for cavitation is minimal.

However, during braking the DCV spool will begin to de-stroke (close) to generate a braking

torque on the motors. This will limit the flow to the motor (which is acting as a pump during braking)

and would cause cavitation if the anti-cavitation valves were not installed. As Figure 2.19 illustrate

the high-pressure fluid from the meter-out side of the motor will take the path through the check

valves and back to the meter-in side of the motors. This will help to keep sufficient flow to the

meter-in side of the motors during braking. A pressure relief valve which opens at 4 bar is also

mounted prior to the tank which is to prevent the fluid from taking the path directly to tank.

2.11 Brake Valve

In Figure 2.20 is a simplified schematic of a brake valve illustrating its function. When the system

is idle the brake valve is closed due to the adjustable spring mounted in the brake valve. The

spring is always acting with a closing force on the valve. When the system is operating,

accelerating or keeping a constant speed, the valve is open. It opens because of the external pilot

pressure line which is supplied from the motor meter-in side (motor inlet), which in turn overcomes

the spring force and opens the valve. As long as the motor meter-in side pressure is kept at a

level where it overcomes the spring force the valve stays open. [16]

Figure 2.20: Simplified schematic of a brake valve. Figure source: [16].

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However, when the motor meter-in pressure drops the valve will begin to close. When the brake

valve closes, the pressure will increase at the motor meter-out(motor outlet). This means that the

pressure at the meter-out side will become greater than the meter-in, which in turn will generate

a negative torque, braking motion, on the motor. However to keep the pressure from increasing

too much there is an internal pilot pressure line. This works the same way as the external pilot

pressure, but the piston area for this line is much smaller. For example, if the brake valve opens

at 10 bar from the external pilot pressure and has a ratio of 8:1 there needs to be 80 bar at the

internal pilot pressure for the valve to open. The ratio can vary between different applications. [16]

In Figure 2.21 is an example shown with the conventional system with a brake valve. The figure

also displays how the fluid flows in the system during acceleration.

Figure 2.21: Simplified schematic of the conventional solution with a brake valve, accelerating phase.

In Figure 2.22 is an illustration of the flow shown during the braking phase. The operator has

released the gas pedal which cuts off the pump supply pressure. The brake valve then reroutes

the flow from the meter-out side to meter-in side of the motor (which acts as a pump during

braking). As a result, cavitation is avoided. The fluid will flow according to Figure 2.22 until the

pressure has decreased sufficiently. Once the pressure is sufficiently low it will take the path to

the tank.

Figure 2.22: Simplified schematic of the conventional solution with a brake valve, braking phase.

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2.12 Road Vehicle Dynamics

In Figure 2.23 is an illustration of the torque sources present in the hydraulic transmission during

operation.

Figure 2.23: Torque sources in the system during operation.

The motor torque is theoretically expressed with equation 2.3 where 𝐷𝑚 is the displacement

setting, Δ𝑝 is the pressure difference from the motor inlet to the outlet and 𝜂ℎ𝑚 is the

hydro-mechanical efficiency of the motor.

𝜏𝑚 = 𝐷𝑚 ⋅ Δ𝑝 ⋅ 𝜂ℎ𝑚 (2.3)

The resistive torque, 𝜏𝑟, is derived with the free body diagram (FBD) in Figure 2.24. From the FBD

the resistive torque is defined as

𝜏𝑟 =

𝐹𝑅 + 𝐹𝑖𝑛𝑐

𝑟𝑤 (2.4)

𝐹𝑅 is the rolling resistance and 𝐹𝑖𝑛𝑐 is an induced force due to inclination, and 𝑟𝑤 is the radius of

the wheels. If the radius of the front and the rear wheels are equal the torque on all four wheels

will be approximately equal. If not, the torque will differ. The aerodynamic drag, 𝐹𝐷, could be

included in the numerator but was neglected in this project. The neglection was made because of

the marginal impact it had on the wheel loader which had a top speed of ~12 km/h.

Figure 2.24: FBD of a wheel loader with resistive forces and torque.

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The net torque, 𝜏𝑛𝑒𝑡, acting on the wheels is described with equation 2.5. Evidently will a positive

net torque result in an accelerating vehicle and a negative net torque with a decelerating vehicle.

𝜏𝑛𝑒𝑡 = 𝜏𝑚 ⋅ 𝑖𝑓 − 𝜏𝑟 − 𝜏𝑓 (2.5)

where 𝑖𝑓 is the gear ratio in the final drive and 𝜏𝑓 is the Coulomb and viscous frictional torque in

the drivetrain and the gearbox (FD). With the net torque the traction force, 𝐹𝑡 , can be calculated

according to equation 2.6.

𝐹𝑡 =𝜏𝑛𝑒𝑡

𝑟𝑤 (2.6)

2.12.1 Resistive Forces

The resistive forces acting on a vehicle were previously introduced in equation 2.4, where 𝐹𝑅 is

the rolling resistance and 𝐹𝑖𝑛𝑐 is the forces due to inclination.

Rolling resistance, 𝑭𝑹, is defined as

𝐹𝑅 = 𝑚 ⋅ 𝑔 ⋅ 𝑐𝑅 (2.9)

where 𝑚 is the mass of the vehicle, 𝑔 is the gravitational constant (~9.81 m/s2) and 𝑐𝑅 is the rolling

coefficient. The rolling coefficient is a dimensionless quantity that can either be calculated with

experimental data using equation 2.9. It can also be assumed with the knowledge of what material

the wheels are made of and which surface is driven on. Experiments conducted in the past has

established the rolling coefficients under different driving conditions, see Table 2.1Error!

Reference source not found.. [17]

Table 2.1: Rolling coefficient under different conditions.

Rolling coefficient [-]

Condition

0.01 – 0.015 Car tires on concrete, new asphalt

0.02 Car tires on gravel – rolled new

0.04 – 0.08 Car tires on solid sand, gravel loose worn, soil medium hard

0.2 – 0.4 Car tires on loose sand

Inclination force, 𝑭𝒊𝒏𝒄, is defined as

𝐹𝑖𝑛𝑐 = 𝑚 ⋅ 𝑔 ⋅ sin (𝛼) (2.10)

where 𝛼 is the inclination angle. Evidently will the inclination force assist the vehicle to accelerate

if driving downhill rather than to decelerate the vehicle.

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2.12.2 Moment of Inertia

The moment of inertia is a quantity that determines the torque required to angular accelerate an

object about a rotational axis. In addition, the quantity is also used to determine the force required

to accelerate an object in a linear motion. [18] In vehicle dynamics the moment of inertia has a

significant impact on the drivetrain during acceleration/deceleration. The total moment of inertia

of the rotational shafts connected to the wheels, 𝐼𝑡, is defined by equation 2.11.

𝐼𝑡 =𝜏𝑛𝑒𝑡

�̇� (2.11)

where �̇� is the rate of angular velocity of the wheels (angular acceleration). 𝐼𝑡 can further be

described with equation 2.12 where it is defined as the sum of the rotational moment of inertia of

the wheels 𝐼𝑤, and linear moment inertia of the vehicle, 𝐼𝑣𝑒ℎ.

𝐼𝑡 = ∑ 𝐼𝑖𝑤

4

𝑖=1

+ 𝐼𝑣𝑒ℎ (2.12)

The vehicle’s traction force be defined by Newton’s second law (equation 2.13), which can be

also be observed in Figure 2.25.

𝐹𝑡 = 𝑚 ⋅ 𝑎 (2.13)

Where 𝑚 is the mass of the vehicle and 𝑎 is the rate of velocity (acceleration).

Figure 2.25: Traction force created by the acceleration.

The acceleration of the vehicle can further be defined as

𝑎 = �̇� ⋅ 𝑟𝑤 (2.14)

Equation 2.13 and 2.14 inserted into equation 2.6 results in

𝑚 ⋅ 𝑟𝑤2 =

𝜏𝑛𝑒𝑡

�̇� (2.15)

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By comparing equations 2.11 and 2.15 it is evident that the linear moment of inertia is

𝐼𝑣𝑒ℎ = 𝑚 ⋅ 𝑟𝑤2 (2.16)

The linear moment of inertia can vary depending if the bucket is empty or loaded. In contrast, the

rotational inertia of the wheels, 𝐼𝑤 is a fixed value and is calculated by obtaining data of the wheels

dimension and their weight. The derivation of 𝐼𝑤 is presented in Appendix B.

2.13 Hopsan

In this project Hopsan was used to model and simulate the system. Hopsan is a simulation tool

mainly used for simulating fluid mechatronic systems. It has been developed at the division of

Fluid and Mechatronic System at Linköping University since the late 1970´s. Since then it has

played an important role for education and research projects at the University. [19]

Hopsan is based on transmission line modelling method (TLM), which is used to model the wave

propagation by using impedances and wave variables with delayed information. Most power-

transmitting systems can be simulated with the TLM method, such as hydraulic, pneumatic,

electric, magnetic and heat transfer systems. [20]

2.14 IQAN

IQAN is a software developed and owned by Parker Hannifin that is used for creating electronic

controls, analyzing and maintenance for mobile systems. The software uses CAN-bus

communication (Controller Area Network) which enables fast data transfer between electronic

control units (ECUs). The IQAN software is divided into multiple different tools for different uses.

IQANdesign and IQANsimulate were the tools used throughout the project. IQANdesign is where

the logic for the ECUs are written and in IQANsimulate the ECUs can be simulated to detect errors

and avoid damage in the system. [2]

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3 Implementation In the following sections the implementation of the simulation model and the practical experiments

that was conducted on the actual wheel loader will be explained. Each step was also continuously

evaluated and analyzed in the project. In Figure 3.1 is an implementation overview presented.

Figure 3.1: Implementation overview.

3.1 Simulation Model

Before initiating any practical experiments, a simulation model of the system was created in

Hopsan. The simulation model resulted in a better understanding of the system. Each component

and its associated parameters were tested and evaluated to interpret the system. The simulation

model allowed for fast and risk-free testing which eliminated factors that could cause problem or

damage to the actual system. General disadvantages and advantages of creating simulation

models have been listed below.

Disadvantages:

• Simulation results may be inaccurate.

Practical Experiments • Implement IQAN

• Practical testing

• Validate simulation model

• Improve IQAN software

Modelling & Simulation • Hopsan model

• Simulation testing

• IQAN logic/model

Implement IQAN software to the wheel

loader and initiate practical testing.

Validate the simulation model with data

obtained from practical tests. Tune IQAN

software to achieve improved braking

characteristics.

Evaluate if the acquired results are

sufficient or if more tests must be

conducted. Analyze and discuss the

findings. Investigate potential future work.

Create an estimated simulation model and

initiate simulation testing. Improve and

validate the simulation model later. Initiate

coding of IQAN software.

Evaluation • Evaluate results

• Analyze/Discuss

• Future work

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Advantages:

• Avoid damage on actual system.

• Easy to test different cases.

• Can be time saving.

• Can obtain a better understanding of the system.

• Enables testing of configurations which are not possible in the hardware.

Considering both the advantages and disadvantages, creating a simulation model was a favorable

approach.

The graphical user interface of the simulation model is displayed in Figure 3.2. The two parallel

motors were modelled as one motor which is applicable method if the displacement of the two

motors are summed together. The simulation model consists of both tailor-made components for

this project and components from the Hopsan default library. The default components generally

allow for multiple parameters to be adjusted which enables simulating different characteristics.

The software also allows for custom coded components written in C++. In the upcoming sections

the parameterization and validation of the main components are presented.

Figure 3.2: Hopsan simulation model with Concept I – Fixed Control.

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3.1.1 Validation of the Pump

The pump used in the simulation model was a default component named Q-Type Pump With

Pressure Control. The four parameters illustrated in Table 3.1 were required to be adjusted in the

component to achieve similar dynamics of the actual system. Other parameters in the component

remained at default values.

Table 3.1: Configurations to achieve a validated pump model.

# Parameter Validated Value

Unit

1 Reference pressure difference

50 [bar]

2 Nominal maximal flow (at 125 rad/s)

0 – 80 [l/min]

3 Regulator inductance at nominal pressure

7E+5 [-]

4 Leakage coefficient 1E-11 [m3/(sPa)]

Parameter 1 was adjusted to the identical value of the actual pump by inspecting the pressure

values in the system at an idle state. To simulate different flow rates, parameter 2 was varied

between 0-80 l/min. The value of 80 was validated with a flow meter installed in the actual system.

The other parameters were iterated until the simulation results converged with the practical

results.

3.1.2 Validation of Pressure Compensator

A graph of the pressure compensator’s characteristics was provided by Parker. The graph

illustrated how Δ𝑝 over the pressure compensator depended on the flow. The pressure

compensator dynamics was modelled accordingly. It was stated that the graph included sensitive

information and therefore the validated parameters will not be presented.

3.1.3 Validation of the DCV

The DCV component was tailor-made for this project. To achieve realistic simulation results,

dynamics of the DCV spool were required to be implemented. The dynamics of the spool were

implemented with a second order low pass filter according to Figure 3.3. In the filter is 𝜔𝑏 the

break frequency, 𝛿 is the damping coefficient and 𝑠 is the complex frequency.

Figure 3.3: The DCV spool’s axial position modelled with a second order low pass filter.

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The function of the second order low pass filter is to filter out high frequencies and to damp

commands given by the operator, 𝑥𝑟𝑒𝑓. As a result, the spool’s actual position, 𝑥𝑎𝑐𝑡𝑢𝑎𝑙, will be

slightly delayed when 𝑥𝑟𝑒𝑓 is altered. For clarification is an example illustrated in Figure 3.4. The

values 𝜔 = 5 Hz and 𝛿 = 1 simulated realistic spool dynamics. [2] These values were not

validated because a spool sensor was not installed in the system.

Figure 3.4: The difference between the commanded position, 𝑥𝑟𝑒𝑓, and the actual position, 𝑥𝑎𝑐𝑡𝑢𝑎𝑙.

To compute the flow through the DCV, the opening area is required observable in equation 3.1.

The equation was also introduced in section 2.3 Directional Control Valve (DCV). 𝐶𝑞 was assumed

constant at the value 0.67 and 𝜌𝑜𝑖𝑙 was constant at 890 kg/m3, which is a standard oil density. [2]

𝑞(𝑡) = 𝐶𝑞 ⋅ 𝐴 ⋅ √2

𝜌𝑜𝑖𝑙⋅ 𝛥𝑝(𝑡) (3.1)

A data table of the opening area, 𝐴, is coded within the component with a resolution of 0.050 mm

from 0-8 mm (i.e.160 data points). Depending on the spool stroke, the opening area is linearly

interpolated with a lookup table function. The data table was provided by Parker and is validated,

at least the meter-out opening area. The data points of the meter-in opening area were

theoretically recalculated because the notches were removed with turning. Initially the meter-in

opening area was too small and restricted flow to the hydraulic motors. The restriction of flow

resulted in cavitation, which is the reason for the removal. The recalculated opening area is

presented in Appendix B.

3.1.4 Validation of the Hydraulic Motor

To model the hydraulic motor a default component named Q Type Variable Machine was used.

The parameterization of the model are shown in Table 3.2.

The two parallel hydraulic motors were modelled as one by summing their displacement

properties and the motor inertias. The low motor inertias had marginal effects on the system but

was included since it was specified in the catalogue. [11]

To model the motor internal viscous, 𝐵𝑚, and the motor leakage coefficient, 𝐶𝑚, the average value

of both motors were derived. Derivation of respective parameter is illustrated in Appendix C and

Appendix D.

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Table 3.2: Configuration of the validated hydraulic motor model.

# Parameter Validated Value

Motor 1 (V14-110)

Motor 2 (V12-80)

Unit

1 Displacement 38 – 190 22 – 110 16 – 80 [cm3/rev]

2 Motor internal viscous friction, 𝐵𝑚

0.047 0.0178 0.0766 [Nms/rad]

3 Motor leakage coefficient, 𝐶𝑚

1.39E-12 1.59E-12 1.19E-12 [m3/(sPa)]

4 Motor inertia 0.0126 0.0082 0.0044 [kgm2]

3.1.5 Validation of the Rotational Gears

The gear ratio, frictional parameter, and the total moment of inertia of the wheel loader was

modelled in the default component illustrated in Figure 3.5.

Figure 3.5: Hopsan default component – Rotational Gear with Inertia.

The gear ratio was found by drawing a line with a distinct color on a wheel and on the rotational

shaft connected to the hydraulic motors. Followed by this an operator drove the vehicle while the

amount of rotations on the shaft and the wheel was counted. When one rotation was counted of

the wheel, 24 rotations were counted on the rotational shaft. Therefore, it was validated that the

gear ratio in the final drive, 𝑖𝑓, was 24.

The total moment of inertia of the wheel loader, 𝐼𝑡 , was calculated to ~2200 kgm2 with equation

3.1 and inserted to the component. Equation 3.2 was previously presented in section 2.12 Road

Vehicle Dynamics.

𝐼𝑡 = ∑ 𝐼𝑖𝑤

4

𝑖=1

+ 𝐼𝑣𝑒ℎ = ∑ 𝐼𝑖𝑤

4

𝑖=1

+ 𝑚 ⋅ 𝑟𝑤2 (3.2)

The inertia of respective wheel 𝐼𝑤, are derived in Appendix B. The mass of the wheel loader was

specified to be 6900 kg without payload in [1].

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In addition, was the frictional torque, 𝜏𝑓, computed in the component according to equation 3.3.

𝜏𝑓(𝑡) = 𝜔(𝑡) ⋅ 𝐵𝐹𝐷 (3.3)

Where 𝜔 is the angular velocity of the rotational shaft. The frictional parameter, 𝐵𝐹𝐷, was used to

compute the Coulomb and viscous friction present in the final drive (gearbox and drivetrain). The

frictional torque wants to decelerate the vehicle to 0 km/h whenever the vehicle is in motion. The

value of 𝐵𝐹𝐷 = 400 Nms/rad was found by observing what the motor torque, 𝜏𝑚, was during cruise

in practical experiments. 𝐵𝐹𝐷 was then iterated in the simulation model until the simulated motor

torque converged with the practical experimental motor torque during cruise.

3.1.6 Validation of the Wheel Loader

The wheel loader component illustrated in Figure 3.6, was tailor-made for the project. It was

utilized to compute the velocity, jerk, and the resistive torque induced by the rolling resistance

and inclination force.

Figure 3.6: Tailor-made Hopsan component.

The velocity of the wheel loader, 𝑣, is computed by equation 3.4.

𝑣(𝑡) = 𝜔(𝑡) ⋅ 𝐶𝑤 (3.4)

where 𝐶𝑤 is the circumference of the wheels and 𝜔 is the angular velocity of the rotational velocity.

The angular velocity is obtained by a transducer and is inherently computed in the component

Rotational Gears with Inertia. 𝐶𝑤 was calculated with equation 3.5, where 𝑟𝑤 was measured to

0.55 m.

𝐶𝑤 = 2 ⋅ 𝜋 ∗ 𝑟𝑤 (3.5)

Jerk, 𝑗, was calculated with the second derivate of the velocity.

𝑗(𝑡) =

𝑑2𝑣(𝑡)

𝑑𝑡2 (3.6)

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The resistive torque acting on the wheels were computed by equation 3.6 which was derived in

section 2.12 Road Vehicle Dynamics. 𝐹𝑖𝑛𝑐 is the force induced by the inclination, 𝐹𝑅 is the rolling

resistance and 𝑟𝑤 is the wheel radius.

𝜏𝑟 =

𝐹𝑅 + 𝐹𝑖𝑛𝑐

𝑟𝑤=

𝑚 ⋅ 𝑔 ⋅ (𝑐𝑅 + sin 𝛼)

𝑟𝑤 (3.7)

where 𝑐𝑅 is the rolling resistance coefficient and 𝛼 is the inclination angle. 𝑐𝑅 was estimated to

0.02 from Table 2.1.

3.1.7 Validation of Hydraulic Volumes

The noted hydraulic volumes illustrated in Figure 3.7, were modelled with the default component

Hydraulic Volume Multi Port.

Figure 3.7: Noted hydraulic volumes.

Illustrated in Table 3.3 are the implemented values for respective hydraulic volume. Volumes 2,

3 and 6 were approximated to be small volumes by visual investigation. Volumes 1,4 and 5

represents hoses which all had an inner diameter of 25 mm.

Table 3.3: Hydraulic volumes.

# Hydraulic Volume [l]

Hose length [m]

1 0.59 1.2

2 0.1 -

3 0.1 -

4 1.23 2.5

5 1.23 2.5

6 10.1 -

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3.1.8 DCV Spool De-stroking Logic

This section presents how the DCV spool was controlled to simulate the velocity of the wheel

loader. The goal was to develop code logic that would brake the wheel loader from top velocity to

0 km/h in approximately five seconds, while also avoiding cavitation and flow through the 330 bar

pressure relief valves.

Simple code logic was initially implemented to the simulation model to observe how the system

reacted. For example, de-stroking the spool linearly in five seconds when the gas pedal command

went to 0 %. This logic would not go on to be the final product, but it resulted in a better

understanding over the transmission dynamics and characteristics. One of the discoveries from

this was that no braking occurred until the spool was ~40 % de-stroked which was useful

information. Another discovery was that de-stroking the spool too fast would simulate cavitation

behavior and pressures exceeding 330 bar.

Concept I – Fixed Control

From the conclusion explained in the section above, a concept was developed which was named

Fixed Control. The concept is illustrated in Figure 3.8. The concept initiates with a fast de-stroke

and proceeds with a slower de-stroke until it is fully closed. 0 % in the figures is equal to 2.35 mm

which is due to the overlap present on the DCV spool. This means that no flow will occur through

the DCV once the spool is at 0 %. This concept always de-stroke the spool in the same manner.

Figure 3.8: Simulation results: logic of Concept I – Fixed Control.

Concept II – Torque Control

An additional concept was also developed. This concept was named Torque Control and it

controlled the spool with the help of pressure sensors on respective side of the motors. This

method resulted in an approximately constant braking torque, hence the name Torque Control.

The concept also initiates with a fast de-stroke as well, but controls the spool different in the

intermediate area depending on the pressure difference, Δ𝑝, over the motor. In Table 3.4 the

principle of the concept is further described. More than three de-stroke speeds were implemented,

but the table illustrates the concept.

Table 3.4: Principle of Concept II – Torque Control.

Pressure difference over the motor, 𝚫𝒑

De-stroking speed of the DCV spool

Δ𝑝 < 50 bar Fast

50 ≤ Δ𝑝 < 150 bar Intermediate

150 ≤ Δ𝑝 < 330 bar Slow

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The concept was inspired by how the brake valves functions, which mechanically regulates the

flow depending on the pressure difference, Δ𝑝, over the motor. In Figure 3.9 is an example of the

concept presented how the spool is controlled during braking.

Figure 3.9: Simulation results: logic of Concept II – Torque Control.

In Figure 3.10 is also an illustration of the simulation model with Concept II, in contrast to

Concept I shown in Figure 3.2.

Figure 3.10 Hopsan simulation model with Concept II – Torque Control.

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3.2 Practical Experiments

The practical experiments were a follow up to the simulation modelling and testing. Software was

developed and implemented in the actual system and further on tested to investigate solutions.

3.2.1 Test Approach

The work process of achieving natural braking characteristics can be illustrated with the flow chart

in Figure 3.11. Hopsan simulation was necessary in the beginning of the project to obtain a better

understanding of the system or when developing new concepts. Once a better understanding of

the system was established, iteration of the IQAN logic would be the method of work. More

information on the IQAN implementation will be presented in an upcoming section, 3.2.4 IQAN

Implementation and Modelling of ECUs.

Figure 3.11: Flow chart of the work process to achieve natural braking characteristics.

Obtained data from practical tests were also used to validate the Hopsan simulation model. The

results from the practical tests are presented in chapter 4 Results.

3.2.2 Test Description

The tests were conducted in the same manner each time to obtain consistent collections of data.

Tests were conducted by accelerating the wheel loader to a certain speed and thereafter releasing

the gas pedal. Once the gas pedal was released the developed IQAN logic would de-stroke the

DCV spool which consequently braked the wheel loader. No mechanical brakes are used on the

data that is presented.

3 Implementation

33

3.2.3 Cavitation

The previously small meter-in opening area was restricting the flow too much during braking. The

spool was modified to increase the opening area. This resolved the cavitation tendencies in most

cases. Cavitation could still occur if the DCV spool was de-stroked too fast. But in the scope of

the project and braking the wheel loader in approximately five seconds, cavitation was no longer

an issue. The meter-in area opening was changed accordingly in the simulation model even

though no signs of cavitation had been present in the simulation prior to the change.

During the practical experiments, cavitation was a problem and different solutions were applied

to eliminate it. The underlying factor was that the hydraulic motors did not receive enough fluid

during braking and therefore experienced cavitation. An idea to resolve the problem was to have

a non-zero gas pedal command during braking. When the operator releases the gas pedal, in the

original model, the command instantly goes to 0 % and the pump displacement instantly goes to

0 %. Instead, the software was modeled in a way that a gas pedal command was kept during

braking when no command was given from the operator. The idea was that the pump

displacement would not be 0 % and help supply flow to the motors and avoid cavitation.

Unfortunately, this had no impact on the result and cavitation was still occurring.

3.2.4 IQAN Implementation and Modelling of ECUs

The Zettelmeyer system consisted of several electronic control units (ECUs) with inputs and

outputs that required modelling for the system to function as intended. Presented below are the

modelling of relevant ECUs.

Solenoid ECU

The solenoid ECU which translated electric current to DCV spool position was modular to accept

current from 0-700 mA. However, it was presented previously that the spool was overlapped and

a deadband had to be considered. This implies that a certain value of current was required to be

transmitted before any flow occurred through the DCV, see Figure 3.12. Likewise, would the flow

also saturate at a certain current, meaning that the spool is fully opened and any increase in

current is only excessive.

Figure 3.12: How transmitted current to the ECU affected the flow and the DCV spool.

It was found that the required current to overcome the overlap was ~330 mA and that the DCV

spool was at maximum position at ~550 mA. These values were implemented in IQAN according

to the logic illustrated in Table 3.5.

3 Implementation

34

Table 3.5: Spool position command and its translation to current.

Spool position command [%] Current [mA]

0 % 0 mA

> 0% > 330 mA

100 % 550 mA

This implementation was necessary since the ECU only accepted values from 0 – 100 % which

consequently translated the percentage value into electric current. The implementation increased

the controllability of the DCV spool position which resulted in that better braking characteristics

was able to be modelled.

Torque Data

The obtained torque data was calculated by using the displacement setting of the motors, 𝐷𝑚,

and the pressure sensors present in the machine, equation 3.7. This data was valuable to get an

interpretation of how hard the vehicle was braking.

𝜏𝑚 = 𝐷𝑚 ⋅ Δ𝑝 (3.7)

Speed Data

The obtained speed data was calculated with the help of a volumetric flow transducer installed at

the pump outlet, 𝑞𝑝. With the knowledge of the motor displacement setting, 𝐷𝑚, and the flow

through the hydraulic motors the rotational speed of the motor, 𝑛𝑚, was calculated with equation

3.8.

𝑛𝑚 =𝑞𝑚

𝐷𝑚 (3.8)

It was assumed that 𝑞𝑝 = 𝑞𝑚, which is not true since small amount leakage occur prior to the

motor. With the rotational speed of the motors it was translated to vehicle speed with the wheel

circumference and the final gear ratio value, equation 3.9.

𝑣 = 𝑛𝑚 ⋅ 𝑐𝑤 ⋅ 𝑖𝑓 (3.9)

A possible improvement to obtain the speed data from the wheel loader would be a Hall effect

sensor mounted on the drivetrain shaft, which counts magnetic pulses when rotations occur. With

the data can the rotational speed of the drivetrain shaft be translated into the speed of the vehicle.

An attempt was made to install a Hall effect sensor into the system. However, the sensor was

either defect or incorrectly installed.

3 Implementation

35

3.3 Evaluation of Braking Characteristics

The obtained data was evaluated continuously throughout the project. When evaluating the

braking characteristics both and different criteria were used for the simulation model and for the

actual system. The different criteria are presented in Table 3.6.

Table 3.6: Different criteria used to evaluate the braking characteristics.

Criterion Simulation Model

Actual System

Braking time delay. The parameter was defined as the required time for the positive motor torque to transition to a negative torque (braking torque), when the gas pedal was released. A low as possible time delay was considered advantageous.

Used Used

The magnitude of braking torque. Too high braking torque will result in the operator to accelerate towards the windshield, which is favorable to be avoided. Too low braking torque will result in less deceleration, which means the goal to decelerate the wheel loader from top speed to 0 km/h in five seconds may not accomplishable.

Used Used

Observation of significant spikes in the motor torque. Less spikes were favorable.

Used Used

Jerk, which is the rate of acceleration. Less spikes were considered advantageous.

Used Not used

Avoid cavitation and flow through the 330 bar pressure relief valves.

Used Used

Feeling, which is subjective. Not Used Used

The velocity signal was noisy and resulted in a calculated jerk which was uninterpretable.

Therefore, was this criterion not used to evaluate the braking characteristics for practical

experiments.

3 Implementation

36


4 RESULTS

37

4 Results In the upcoming sections are results from the simulation model and practical tests presented.

4.1 Simulation Results

On the upcoming pages are the simulation results presented. For respective concept two different

results are presented based on the wheel loader accelerating to either ~7 or ~12 km/h and

consequently followed by a braking phase. The gas pedal command is released from 100 – 0 %

at 20 seconds. Consequently, the developed logic will de-stroke the DCV spool and thus the

wheel loader will begin to brake. The simulation results are also displayed with DCV spool position

in percentage, where 0 % is 2.35 mm. This is due to the overlap present on the DCV spool.

The results which are presented had the following settings during the full simulation.

• The total motor displacement setting was kept constant at 76 cm3/rev (25 % of full

displacement).

• ICE rotational speed was kept constant at 1700 rpm.

• 0 % inclination during the full simulation, meaning an ideal road.

4 RESULTS

38

Simulation results from Fixed Control are illustrated in Figure 4.2. As mentioned in the chapter 3

Implementation, this concept always de-strokes the DCV spool in the same manner which can be

observed by comparing the results. The negative aspect of this, is that it will always take the same

time to decelerate the vehicle.

Fixed Control (~12 km/h)

Figure 4.1: Simulation results – Fixed Control at 12 km/h.


Figure 4.2: Simulation results – Fixed Control at 7 km/h.

4 RESULTS

39

In Figure 4.4 are the simulation results from Torque Control illustrated. By comparing the two

results it can be observed that the wheel loader is able to decelerate ~1 second faster when

beginning to brake from a lower velocity.

Torque Control (~12 km/h)

Figure 4.3: Simulation Results – Torque Control at 12 km/h.


Figure 4.4: Simulation results – Torque Control at 7 km/h.

4 RESULTS

40

4.2 Practical Results

In the following section the results from the practical experiments are presented. Respective

concept was tested at two different velocities, ~12 km/h which is the maximum velocity of the

wheel loader and ~7 km/h. The slower velocity was difficult to maintain constant at 7 km/h and it

differs marginally between the presented results. The signal from the volumetric flow transducer

used to calculate the velocity was very noisy and is the reason for the noisy data. For each of

these tests a brake sequence was conducted going forward and in reverse. The forward brake

which is always the first in respective test had a slight downhill and the reversed brake had a

slight uphill. This affected the braking characteristics.

The results which are presented had the following settings during the tests.

• The total motor displacement setting was kept constant at 76 cm3/rev (25 % of full

displacement).

• ICE rotational speed was kept constant at 1700 rpm.

4 RESULTS

41


Figure 4.5: Practical results – Fixed Control at 12 km/h.


Figure 4.6: Practical results – Fixed Control at 7 km/h.

4 RESULTS

42


Figure 4.7: Practical results – Torque Control at 12 km/h.


Figure 4.8: Practical results – Torque Control at 7 km/h.

4 RESULTS

43

4.2.1 Braking Time Delay

The required time from releasing the gas pedal to achieve a braking torque (negative torque) in

the transmission is presented in Figure 4.9. The quantity was a significant parameter to observe

when evaluating the braking characteristics.

Figure 4.9: Marked delay in motor torque from respective concept and velocity.

4.3 Cost Analysis

Illustrated In Table 4.1 is a cost analysis for respective concept. Only the hardware components

and work regarding the system changes has been considered.

Table 4.1: Cost for the conventional system and the difference for the two developed concepts.

Cost Conventional system

Concept II Concept I

Brake valve (~ €500) [2]

Tuning (-)

Pressure sensor (~ €50 – €100 / unit) [2]

Cabling (-)

Input on controllers (-)

4 RESULTS

44


5 DISCUSSION

45

5 Discussion In the following section thoughts are discussed of the conducted project.

5.1 Results Discussion

Fixed Control worked well during high speed and no major instabilities occurred during the braking

phase which can be observed in Figure 4.5. The braking torque steadily increases and decreases

during the braking. A torque spike can be observed immediately after the braking phase. The

spike occurred due to the wheel loader coming to a complete stop when the spool de-stroked

fully. The phenomenon is more prevalent to occur while going downhill than uphill since gravity

will either increase or decrease the pressure build up. This can be observed in Figure 4.5 where

the first braking phase has a significant spike while the second phase has a less noticeable spike.

Leakage occurred when the DCV was fully closed and is the reason why the motor torque is not

0 Nm during standstill. The performance of Fixed Control decreases with lower speeds, which

can be observed by the instabilities produced at the end of the braking phase in Figure 4.6. The

instabilities resulted in an uncomfortable brake feeling. Another downside of Fixed Control is that

it will always decelerate in the same time, regardless of what speed the braking initiates from.

This is because the concept is programmed to always de-stroke the DCV spool in the same

manner.

Torque Control compared to Fixed Control works significantly better for multiple velocities which

is presented in the results. Figure 4.7 illustrates that Torque Control was able to regulate at

approximately -100 Nm during the whole braking phase which resulted in comfortable braking

characteristics. The positive aspect of Torque Control is the ability to decelerate faster during

different circumstances, which Fixed Control cannot. The different braking times achieved with

Torque Control can be observed in Figure 4.7 and 4.8. The concept is not flawless however, and

the software can be tuned further for better performance at lower speeds. In Figure 4.8 an

overshoot occurred to almost -200 Nm before it stabilized and regulated at -100 Nm for a brief

period.

The time delay before any braking initiates is a parameter that affects the feeling and

performance. The delay is the time from when the gas pedal is released until a braking torque

(negative torque) is acquired. In Figure 4.9 the delay for each of the four tests have been marked.

The delay is significantly longer while going at lower speeds and shorter at higher speeds. The

reason for the increased delay is the lack of pressure build up on the meter-out side of the motor.

When going slower the flow is lower and it is therefore required that the DCV spool is de-stroked

faster to have a faster response time. This is the reason the delay is larger for Fixed Control at 7

km/h. The spool is de-stroked too slow and it takes almost 2 seconds until the spool has de-

stroked sufficiently to generate a braking torque. Torque Control can handle this better, but the

delay is still larger at 7 km/h compared to 12 km/h.

Respective test for each concept managed to avoid cavitation and flow through the 330 bar

pressure relief valves. Cavitation was not detected with sensors but with the unpleasant sound it

produces. The wheel loader experienced heavy cavitation early in the project and it was evident

whenever the phenomenon occurred afterwards.

5 DISCUSSION

46

5.2 Cost Analysis Discussion

Cost reduction has been the driving factor for this project. The brake valve removal represents

the greatest cost reductions with approximately €500, which can be compared to being a third of

the hydraulic motor cost. The tuning process is another significant cost for the conventional

system. The new developed concepts still require tuning, and it is uncertain of how much time

can be saved regarding the tuning process.

Torque Control decreases cost compared to the conventional system, but there are additional

costs attached to this solution. The greatest costs are the two pressure sensors with a cost of

approximately 50-100 euro per unit. Additional cabling and inputs on controllers are minor costs

that also are required. Fixed Control is the most cost-efficient solution that has been developed

during this project and no additional costs apart from tuning are required. However, as discussed

earlier the performance is lacking in comparison to Torque Control. Benefits and drawbacks must

be considered to determine which concept should be used.

5.3 Research Questions

RQ1: Can braking characteristics with a natural feeling be achieved with the DCV?

According to several people were natural braking characteristics achieved with DCV. It was

unanimous that Torque Control was able to achieve a better brake feel during different

circumstances compared to Fixed Control. However, the question of feel is subjective and some

may claim the developed concepts produced unnatural braking characteristics. No operator with

expertise within the area drove the vehicle. Feedback from an operator with expertise would most

likely confirm if respective concept can produce a natural brake feel. Another aspect to consider

it that different applications also require different characteristics.

RQ2: Is it possible to avoid cavitation when braking with the DCV?

Avoiding cavitation is crucial since it can cause wear and damage to components. If it cannot be

avoided, then it is a great incentive not to continue with the approach and instead investigate

other solutions. During the first stages of practical experiments cavitation was a problem and

different solutions were applied to eliminate it. The underlying factor was that the hydraulic motors

did not receive enough fluid during braking and therefore experienced cavitation. An idea to

resolve the problem was to have a non-zero gas pedal command during braking. When the

operator releases the gas pedal, in the original model, the command instantly goes to 0 % and

the pump displacement instantly goes to 0 %. Instead, the software was modeled in a way that a

gas pedal command was kept during braking when no command was given from the operator.

The idea was that the pump displacement would not be 0 % and help supply flow to the motors

and avoid cavitation. Unfortunately, this had no impact on the result and cavitation was still

occurring.

Cavitation was avoided when DCV spool meter-in opening area underwent turning. Initially the

spool meter-in area was too small and restricted flow from the load sensing pump to enter the

hydraulic motors during braking which caused cavitation. When braking slowly cavitation could

be avoided, but when braking more aggressively the phenomenon occurred. Once the meter-in

side area of the DCV spool was increased significantly, cavitation tendencies were avoided. But

to keep in mind the tests that was carried out was at low speeds. If the speed increases or the

mass of the wheel loader increases, with a loaded bucket or similar, then perhaps cavitation would

be a reoccurring problem.

5 DISCUSSION

47

As mentioned earlier in the Hopsan simulation model cavitation was not a problem prior to the

meter-in area change. The anti-cavitation valve provided the motor meter-in side with flow and

therefore the pressure never dropped under 1 bar, which suggest that cavitation is avoided. This

was a learning experience where the difference between simulation and reality was noticeable.

RQ3: Can the same performance be acquired using the DCV as with the brake valve?

From the project it is uncertain if the same performance can be acquired by utilizing the DCV as

with the brake valve. No practical experiments were conducted on the conventional system with

a brake valve and it is therefore difficult to answer. It was stated from multiple sources that the

brake valve has a fast response time [2]. When the gas pedal command is released from 100 – 0

% in the conventional system, braking will begin before the command reaches 0 %. This means

the braking valve has a low braking time delay, which is one of the advantages of the brake valve.

To achieve similar functionality when utilizing the DCV, logic was implemented which observed

the derivate of the gas pedal command. Thus, the DCV spool would begin to de-stroke to produce

a braking torque before the gas pedal command was 0 %.

RQ4: Can a realistic simulation model of the system be created?

A realistic simulation model of the system was created to a certain degree. By comparing the

simulation and practical results the accuracy of the simulation model seems decent. The motor

torque during cruise deviates marginally between the simulation and practical results. The

deviations are likely caused of the inclination of the road. Practical results were conducted on an

inclined road and the simulation were simulated with 0 % inclination. Excluding those deviations,

the simulation dynamics and magnitude of values seem identical to the practical results.

To achieve accurate simulation models of realistic systems is challenging, especially when it

comes to complex systems where a lot of parameters influence each other. The demand for

accuracy can vary on applications. Evidently is a model with greater accuracy better, but trade-

offs exist where accuracy versus the time consumption is not worth it for some applications. The

modelling process of the system and the aim to achieve as high accuracy as possible, resulted in

a better understanding of the system.

5.4 Validation of Simulation Model

It was specified in [1] that the weight of the wheel loader was 6900 kg without payload. There is

a risk that the value deviates from the true value. Many adjustments have been conducted on the

system since it was produced. Namely the removal of the counterweigh originally mounted at the

rear. The original hydraulic components have also been replaced with new. As a result, the

calculated total moment of inertia may deviate from the true value.

The dynamics of the DCV spool is not validated. The parameters used to simulate the dynamics

of the spool were only recommended values to use by the supervisors. To further validate the

model, a spool position sensor is required to be installed in the DCV and consequently iterate the

break frequency, 𝜔𝑏, and the damping, 𝛿, until the simulation results converge with the practical

results. The dynamics of the pressure compensator and the meter-out opening area of the DCV

spool is validated since data tables were provided by Parker. With the recalculation of the meter-

in opening area, the parameter should be validated as well. The frictional parameter, 𝐵𝐹𝐷, should

be relative accurate since it was iterated until simulated results of the motor torque during cruise

converged with the practical results.

5 DISCUSSION

48

5.5 Fixed Control vs. Torque Control

Based on the obtained results the Torque Control concept seems to be the evident winner.

However, to completely disregard Fixed Control would be unwise. Fixed Control has potential and

it received significantly less hours of development than Torque Control. Fixed Control worked well

at high speed and worse at lower speeds. This is because it has a predetermined de-stroking

profile which is identical for every speed. The de-stroking profile was optimized to function at high

rather than low speed. To fully utilize Fixed Control, multiple predetermined de-stroking profiles

are required. Depending on the speed the DCV spool would de-stroke differently to suit the

specific speed. As a result, can a control be acquired that could operate well during varying

speeds.

An additional benefit with Fixed Control is that no pressure sensors over the motor are required,

thus a cheaper option. It will also make the system more robust and reliable since there will be no

problems caused by faulty pressure sensors. However, Torque Control is able to avoid cavitation

during different circumstances with the aid of the pressure sensors. The different circumstances

may be a loaded bucket or going downhill which increases the pressure in the system during

braking. When the pressure increases the DCV spool must de-stroke slower to prevent cavitation

from occurring. Torque Control can regulate the spool de-stroke speed according to these

circumstances and avoid cavitation which Fixed Control cannot.

5.6 Safety

Safety is crucial when operating heavy mobile since an error can have severe consequences.

Therefore, it is crucial to keep in mind of possible risks and how to avoid them in the development

phase. Switching the braking control from a mechanically controlled brake valve to an electronic

control can cause new safety concerns. The Torque Control for example uses two pressure

sensors that regulates the braking. A faulty sensor could result in that the wheel loader comes to

an abrupt stop or experiencing no braking at all. Respective case could result in damage and

safety concerns for both the machine operator and surrounding people. Avoiding these kinds of

accidents can be eliminated by using multiple sensors to have redundancy. Code that can detect

faulty sensors which would switch the Torque Control to Fixed Control which does not rely on

pressure sensors could also be a potential solution.

5.7 Sources of Error

Although experiments were conducted in the same manner, the obtained data were marginally

inconsistent. The main cause was likely the surface and the difference in incline at the time for

braking. The location where the experiments were conducted had a general slope in one direction

which was considered and noted in the data. But the slope in each direction could still differ.

Depending on where the braking occurred the magnitude of the slope could be different and affect

the data. Another parameter that potentially affected the system were oil temperature. At the start

of each experiment when the wheel loader had cooled down, it behaved differently than at the

end of the previous experiment. The logical conclusion was that temperature affected the system

behavior.

5 DISCUSSION

49

5.8 Future Work

Below are recommendations of what topics should be further worked on.

5.8.1 Displacement Control

It could be of interest to develop software which regulates the motor displacement during

acceleration, cruise and braking. The motor displacement was kept constant at 25 % of its full

displacement during all tests. Regulation of the hydraulic motor displacement affects the braking

characteristics and will require a more complicated solution which was the reason for the constant

value. The reasoning for 25 %, was because it offered the vehicle to accelerate quickly and reach

a reasonable top speed. But keeping the motor displacement constant is not optimal, since a

regulated motor displacement could offer faster acceleration and higher top speeds.

5.8.2 Torque Control for Different Speeds

To reduce the brake time delay for Torque Control at lower speeds, the software logic could be

further developed. The concept is programmed to initiate with a fast de-stroke until a position of

approximately 40 %, and in the remaining range 40 – 0 %, it relies on pressure regulation to keep

a constant braking torque. To reduce the delay for lower speeds the initial de-stroke could be

decreased to a lower position to generate a braking torque faster. This was implemented in the

last week of the project which reduced the delay significantly. However, there was no time to

document these findings.

5.8.3 Fixed Control for Different Speeds

The full potential for Fixed Control has not yet been achieved, as mentioned earlier. Continuing

the work on the current fixed controller to optimize it further as well as expanding and create

braking profiles for multiple different speed ranges would be interesting. This would allow for a

system that is not relying on pressure sensors which would increase the reliability and create a

more cost-efficient system.

5.8.4 Extreme Cases

The wheel loader has thus far not been tested at extreme cases. For example, braking in a

downhill with a fully loaded bucket. The project has mainly focused on even ground with no

significant incline in either direction. It is recommended to investigate the performance at extreme

cases.

5.8.5 Heavy Mobile Machinery

Further investigations are required to be conducted on other heavy mobile machines that has a

greater overall mass than the Zettelmeyer. The Zettelmeyer had a total weight of approximately

6900 kg, while unloaded. Increment in weight will increase the total moment of inertia which has

a significant impact on how fast the DCV spool can be de-stroked and avoid cavitation during

braking.

In addition to that, should also vehicles with greater top speeds be investigated. The Zettelmeyer

achieved a top speed of ~12 km/h and it was found in the simulation model that greater top speeds

affects how fast the DCV spool can be de-stroked to avoid cavitation during braking.

5 DISCUSSION

50

5.8.6 Components and Configurations

In this project the focus was to electronically control the DCV spool to achieve natural braking

characteristics. No investigations were conducted on finding the best suited components or

configurations for this system. At the end of the project the system consisted of standardized

components and not components tailored for this application. Test with better suited hydraulic,

mechanical or electronic components in the system could potentially result in better braking

characteristics and a more stable system. Further investigations are recommended to be

conducted on the following topics:

• DCV dimensions.

• Meter-in and meter-out area of the DCV spool.

• Hose lengths and hydraulic volume between components.

• Oil type and temperature ranges.

• Pump, motor and pressure compensator configurations.

6 CONCLUSION

51

6 Conclusion Utilizing the directional control valve to achieve a hydraulic brake function, has potential to be a

solution for heavy mobile machinery with open-loop transmission in near future. The initial

established goals for the project was achieved and two well-functioning brake logics were

developed. Respective concept can be tuned further for improvements, but in the scope of this

project the goals were accomplished.

The benefits with electronic controls have been proven to have great freedom and can easily be

tuned with software code. Multiple software configurations can be stored in the system which

offers the operator to alter between different braking profiles. For this to function properly

however, the correct software is required which regulates the DCV de-stroke accordingly to

accommodate the desired braking torque.

However, extensive testing and investigations are required to be conducted on systems with

greater loads, higher speeds, and other more challenging conditions. It needs to be investigated

if cavitation or other system instabilities can be avoided when these parameters increase in

magnitude. The two developed concepts in this project will be a baseline to continue developing

from and to find the optimal solutions to suit the application. It must also be investigated if the

solution is legal. Laws may prevent the solution to become a standardized solution because of

safety concerns.

REFERENCES

52

References

[1] Volvo Construction Equipment, "ZETTELMEYER ZL 802 SI / L40".

[Accessed February 2021].

[2] Parker Hannifin, Oral reference, 2021.

[3] Parker Hannifin, www.parker.com. [Accessed February 2021].

[4] M. Borghi, M. Milani and R. Paoluzzi, "Influence of notch shape and number of notches

on the metering characteristics of hydraulic spool valves" August 2005.

[Accessed March 2021].

[5] Sivaranjith, "Hydraulic Servo Valves - Basic Types and Operation".

[Accessed April 2021].

[6] E. Olson, "How does a pressure-compensated flow control valve work?" GlobalSpec,

October 2019. [Accessed March 2021].

[7] Robotics, VEX, "Using Gear Ratios with the V5 Motor". [Accessed March 2021].

[8] L. Ericson and K.-E. Rydberg, Basic Fluid Power Components and Systems, Linköping:

Linköping University, September 2017.

[9] Parker Hannifin, "P2/P3 Series Piston Pumps Varible Displacement" 2009.


[10] B. Casey, "Understanding Load-sensing Control" March 2006. [Accessed March 2021].

[11] Parker Hannifin, "Hydraulic Motors V12, V14, T12 Variable Displacement" October

2003. [Accessed April 2021].

[12] B. Casey, "Hydraulic Motor Face-Off: Bent Axis vs Axial Piston" February 2017.

[Accessed February 2021].

[13] M. Gannon, "What is a relief valve?" June 2019. [Accessed March 2021].

[14] B. Gillum, "Hydraulic pump cavitation: What is it and how can you prevent it?" October

2018. [Accessed March 2021].

[15] A. Smiley, "The Importance of Check Valves in Hydraulic Systems" April 2018.


[16] HydraulicValve, "Hydraulic Brake Valve Application". [Accessed March 2021].

[17] R. Kurtus, "Coefficient of Rolling Friction" November 2016. [Accessed April 2021].

[18] R. Nave, "Moment of Inertia Rotational-Linear Parallels". [Accessed April 2021].

https://www.volvoce.com/global/en/product-archive/wheel-loaders/zettelmeyer/zl-802-si-l40/

https://www.researchgate.net/publication/255745326_Influence_of_Notch_Shape_and_Number_of_Notches_on_the_Metering_Characteristics_of_Hydraulic_Spool_Valves

https://www.researchgate.net/publication/255745326_Influence_of_Notch_Shape_and_Number_of_Notches_on_the_Metering_Characteristics_of_Hydraulic_Spool_Valves

https://forumautomation.com/t/hydraulic-servo-valves-basic-types-and-operation/4209

https://insights.globalspec.com/article/12786/how-does-a-pressure-compensated-flow-control-valve-work?_sm_au_=iVVSR6NNtHj4PsTJ6F1pQK7GkjfLc

https://kb.vex.com/hc/en-us/articles/360035590932-Using-Gear-Ratios-with-the-V5-Motor?_sm_au_=iVVSR6NNtHj4PsTJ6F1pQK7GkjfLc

https://www.parker.com/Literature/Hydraulic%20Pump%20Division/P2_P3-Files/P2-P3%20Series%20Piston%20Pumps%20HY28-1559-01-PT.pdf

https://www.machinerylubrication.com/Read/859/load-sensing-control

https://www.lifcohydraulics.com/SpecFiles/Catalogs/Parker-(V12,V14,T12)-Hydraulic-Motors,-Variable-Displacement.pdf

https://www.hydraulicspneumatics.com/hydraulics-at-work/article/21886658/hydraulic-motor-faceoff-bent-axis-vs-axial-piston#:~:text=Bent%20axis%20motors%2Despecially%20those,design%20of%20the%20same%20displacement.

https://www.mobilehydraulictips.com/what-is-a-relief-valve/

https://www.munciepower.com/company/blog_detail/hydraulic_pump_cavitation_what_is_it_and_how_can_you_prevent_it_

https://www.machinerylubrication.com/Read/31135/hydraulic-check-valves

http://www.valvehydraulic.info/hydraulic-pressure-control/hydraulic-brake-valve-application.html?_sm_au_=iVVvNvqW1B7PVrsq6F1pQK7GkjfLc

https://www.school-for-champions.com/science/friction_rolling_coefficient.htm

http://hyperphysics.phy-astr.gsu.edu/hbase/mi.html

REFERENCES

53

[19] Linköping's University, "Hopsan" Linköping University, March 2021.


[20] P. Krus, "Transmission Line Modelling for System Simulation" 2021.


https://liu.se/en/research/hopsan?_sm_au_=iVVvNvqW1B7PVrsq6F1pQK7GkjfLc

https://hopsan.github.io/tutorials/mainTLMtutorial.pdf

APPENDICES

54

Appendices

• Appendix A – Symbols

• Appendix B

• Appendix C – Derivation of the wheel’s moment of inertia, 𝐼𝑤

• Appendix D – Derivation of Motor Internal Viscous Fricton

• Appendix E – Derivation of Motor Internal Leakage Coefficient

APPENDICES

55

Appendix A – Symbols

Symbol Name Additional information

Hydraulic flow

direction

Variability

Spring

Generally variable without the variability arrow depicted.

Electronically

controlled Generally variable without the variability arrow depicted.

Pressure controlled Generally variable without the variability arrow depicted.

Fluid reservoir/tank

Atmospheric pressure in tanks (~1 bar).

1) Hoses installed into the fluid 2) Hoses installed above the fluid

Proportional 4/3 DCV with load

sensing

Parallel lines on respective side indicates proportionality.

Check valve

Enables the fluid to only flow in one direction.

Internal

combustion engine

Load sensing pump with variable

displacement

Pump which regulates the displacement depending on the load pressure. 2) Hopsan symbol

Motor with variable displacement

Pressure compensator

Pressure relief valve

Opens at a predetermined pressure.

APPENDICES

56

Proportionally controlled pressure

relief valve

Increasing the opening pressure with increasing current.

Proportional Pressure Reducing

Valve

Rotational gear with inertia

Hopsan symbol

Mechanical rotating

shaft Hopsan Symbol

Hydraulic volume

Hopsan symbol. Used to model the volume in hoses between hydraulic components.

Pressure

transducer Hopsan symbol

Angular velocity


Flow rate


APPENDICES

57

Appendix B – Recalculated Meter-in Area Opening The limited area opening was implemented in the lookup table function which is the lowest

opening area of either radial or axial.

DCV spool data:

• 𝑑𝑖𝑛𝑛𝑒𝑟 = 12 𝑚𝑚

• 𝑑𝑜𝑢𝑡𝑒𝑟 = 20 𝑚𝑚

• Meter-in overlap: 2.55 mm

𝐴𝑎𝑥𝑖𝑎𝑙 =𝜋 ∗ (𝑑𝑜𝑢𝑡𝑒𝑟

2 − 𝑑𝑖𝑛𝑛𝑒𝑟2 )

4

𝐴𝑟𝑎𝑑𝑖𝑎𝑙 = 𝜋 ∗ 𝑑𝑖𝑛𝑛𝑒𝑟 ∗ 𝑥𝑠𝑝𝑜𝑜𝑙

APPENDICES

58

Appendix C – Derivation of the Wheel’s Moment of Inertia, 𝐼𝑤

APPENDICES

59

APPENDICES

60

Appendix C – Derivation of Motor Internal Viscous Fricton

APPENDICES

61

APPENDICES

62

Appendix D – Derivation of Motor Internal Leakage Coefficient

APPENDICES

63

braking with a directional control valve in a hydraulic

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