Flow Sharing Capability

Displacement Controlled Fluid Power System with

Flow Sharing Capabilities

Robert AnderssonMikael Axin

Division of Fluid and Mechanical Engineering Systems

Master ThesisDepartment of Management and Engineering

LIU-IEI-TEK-A- -09/00577- -SE

Displacement Controlled Fluid Power System with

Flow Sharing Capabilities

Master Thesis in Fluid Power

Department of Management and Engineering

Division of Fluid and Mechanical Engineering Systems

Linköping Universityby



Supervisors: Björn Eriksson

IEI, Linköping University

Daniel Sundkvist

Parker Hannifin AB

Examiner: Karl-Erik Rydberg

IEI, Linköping University

Linköping, 26 February, 2009

Avdelning, Institution

Division, Department

Institutionen för ekonomisk och industriell utvecklingFluid och mekanisk systemteknik

Department of Management and EngineeringFluid and Mechanical Engineering Systems

Datum 2009-02-26Date

Språk

Language

� Svenska/Swedish

� Engelska/English

�

�

Rapporttyp

Report category

� Licentiatavhandling

� Examensarbete

� C-uppsats

� D-uppsats

� Övrig rapport

�

�

URL för elektronisk version

http://www.ep.liu.se

ISBN

—

ISRN


Serietitel och serienummer

Title of series, numberingISSN

—

Titel

Title Displacement Controlled Fluid Power System with Flow Sharing Capabilities

Författare

Author Robert AnderssonMikael Axin

Sammanfattning

Abstract

The purpose of this master thesis is to further develop a displacement controlledfluid power system. It uses similar components as a load sensing system but thepump is controlled in a different way. Instead of a pressure closed loop controlmode the pump operates in an open control mode where the requested displace-ment is set by the operator. This principle might imply higher energy efficiency,faster response and less oscillations.

If the pump is displacement controlled and the valve is equipped with commonpre compensators the flow delivered from the pump needs to be matched by thevalve. A flow map would then be required and problems might occur if incorrectflow is delivered by the pump. A solution to the problem is to utilize pre com-pensators with anti saturation. The flow will then be shared proportionally to theactive actuators and no flow map is needed. Since the compensators will makesure that all flow will reach the actuators the main spool can be manoeuvred toits end position, which allows additional energy savings.

The displacement controlled system has been designed and simulated using thesimulation software AMESim. All components in the system have been modelledand validated using a laboratory platform. The system has also been implementedin a wheel loader application where it can be compared to a load sensing system.Measurements confirm that the energy efficiency is higher in a displacement con-trolled system compared to a load sensing system during a short duty cycle.

Nyckelord

Keywords Energy efficiency, pump pressure margin, compensation, displacement control

http://www.ep.liu.se

Abstract

The purpose of this master thesis is to further develop a displacement controlledfluid power system. It uses similar components as a load sensing system but thepump is controlled in a different way. Instead of a pressure closed loop control modethe pump operates in an open control mode where the requested displacement isset by the operator. This principle might imply higher energy efficiency, fasterresponse and less oscillations.

If the pump is displacement controlled and the valve is equipped with commonpre compensators the flow delivered from the pump needs to be matched by thevalve. A flow map would then be required and problems might occur if incorrectflow is delivered by the pump. A solution to the problem is to utilize pre com-pensators with anti saturation. The flow will then be shared proportionally to theactive actuators and no flow map is needed. Since the compensators will makesure that all flow will reach the actuators the main spool can be manoeuvred toits end position, which allows additional energy savings.

The displacement controlled system has been designed and simulated using thesimulation software AMESim. All components in the system have been modelledand validated using a laboratory platform. The system has also been implementedin a wheel loader application where it can be compared to a load sensing system.Measurements confirm that the energy efficiency is higher in a displacement con-trolled system compared to a load sensing system during a short duty cycle.

v

Acknowledgments

This master thesis has been written at Parker Mobile Systems Team in Borås. Wewould like to thank the whole department for their time and effort. Our supervisorDaniel Sundkvist can always spare a moment for discussions and he has allowedus to go our own way during this master thesis. Anders Eliasson has helped usa lot with the laboratory platform and the test rig. When it comes to technicalissues, Anders Lindström has been of great help.

We would also like to thank Per-Anders Kumlin at Parker Mobile ControlDivision. Because of his master thesis and especially the construction of the testrig, our work became a lot easier.

Our supervisor at the university has been Björn Eriksson. He has helped us alot in almost all possible ways, especially with problem regarding the report. Hehas also been involved in the simulation and the system design.

Finally, we would like to thank our examiner Karl-Erik Rydberg and our op-ponent Karl Pettersson.

Linköping, February, 2009


vii

Contents

1 Introduction 7

1.1 Background . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 71.2 Purpose . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 71.3 Delimitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81.4 Method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81.5 Report Outline . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9

2 Basic Theory 11

2.1 Mobile Fluid Power Systems . . . . . . . . . . . . . . . . . . . . . . 112.1.1 Constant Flow . . . . . . . . . . . . . . . . . . . . . . . . . 112.1.2 Constant Pressure . . . . . . . . . . . . . . . . . . . . . . . 122.1.3 Load Sensing . . . . . . . . . . . . . . . . . . . . . . . . . . 12

2.2 Pressure Compensation . . . . . . . . . . . . . . . . . . . . . . . . 132.2.1 Common Pre Compensation . . . . . . . . . . . . . . . . . . 132.2.2 Pre Compensation with Anti Saturation . . . . . . . . . . . 142.2.3 Post Compensation . . . . . . . . . . . . . . . . . . . . . . . 16

2.3 Displacement Controlled System . . . . . . . . . . . . . . . . . . . 172.3.1 Flow Mapping . . . . . . . . . . . . . . . . . . . . . . . . . 172.3.2 Compensation . . . . . . . . . . . . . . . . . . . . . . . . . 172.3.3 Energy Savings . . . . . . . . . . . . . . . . . . . . . . . . . 17

2.4 Flow Forces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18

3 Design of Simulation Models 21

3.1 Pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 213.1.1 Pressure Control . . . . . . . . . . . . . . . . . . . . . . . . 233.1.2 Displacement Control . . . . . . . . . . . . . . . . . . . . . 23

3.2 Valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 263.2.1 Cartridge Valve . . . . . . . . . . . . . . . . . . . . . . . . . 263.2.2 Main Spool . . . . . . . . . . . . . . . . . . . . . . . . . . . 283.2.3 Common Pre Compensator . . . . . . . . . . . . . . . . . . 303.2.4 Pre Compensator with Anti Saturation . . . . . . . . . . . 333.2.5 ∆pp Limiter . . . . . . . . . . . . . . . . . . . . . . . . . . . 36

3.3 Actuator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 373.4 Load Sensing Systems . . . . . . . . . . . . . . . . . . . . . . . . . 37

ix

x Contents

3.5 Displacement Controlled Systems . . . . . . . . . . . . . . . . . . . 40

4 Validation of Simulation Models 43

4.1 Laboratory Platform . . . . . . . . . . . . . . . . . . . . . . . . . . 434.1.1 ∆p/q Test . . . . . . . . . . . . . . . . . . . . . . . . . . . . 454.1.2 Flow Forces . . . . . . . . . . . . . . . . . . . . . . . . . . . 454.1.3 Step Response . . . . . . . . . . . . . . . . . . . . . . . . . 47

4.2 Test Rig . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 474.2.1 Pump . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 494.2.2 Load Pressure Feedback . . . . . . . . . . . . . . . . . . . . 494.2.3 Pressure Losses . . . . . . . . . . . . . . . . . . . . . . . . . 50

4.3 Load Sensing Systems . . . . . . . . . . . . . . . . . . . . . . . . . 514.3.1 Pump Saturation . . . . . . . . . . . . . . . . . . . . . . . . 514.3.2 Step Response . . . . . . . . . . . . . . . . . . . . . . . . . 53

5 Design of Displacement Controlled Systems 55

5.1 Flow Mapping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 565.1.1 Valve Control . . . . . . . . . . . . . . . . . . . . . . . . . . 565.1.2 Pump Control . . . . . . . . . . . . . . . . . . . . . . . . . 57

5.2 System Characteristics . . . . . . . . . . . . . . . . . . . . . . . . . 585.2.1 Response Behaviour . . . . . . . . . . . . . . . . . . . . . . 585.2.2 Dynamic Stability . . . . . . . . . . . . . . . . . . . . . . . 585.2.3 System Pressure . . . . . . . . . . . . . . . . . . . . . . . . 58

5.3 Incorrect Flow Delivery . . . . . . . . . . . . . . . . . . . . . . . . 595.3.1 Not Enough Flow is Delivered . . . . . . . . . . . . . . . . . 605.3.2 Too Much Flow is Delivered . . . . . . . . . . . . . . . . . . 61

5.4 Displacement Controlled System with Flow Sharing Capabilities . 625.4.1 One Actuator . . . . . . . . . . . . . . . . . . . . . . . . . . 635.4.2 Two Actuators . . . . . . . . . . . . . . . . . . . . . . . . . 645.4.3 Flow Forces . . . . . . . . . . . . . . . . . . . . . . . . . . . 655.4.4 Position Feedback . . . . . . . . . . . . . . . . . . . . . . . 665.4.5 Different Loads . . . . . . . . . . . . . . . . . . . . . . . . . 665.4.6 Redesign of the Flow Sharing Compensator . . . . . . . . . 685.4.7 Lowering Motion . . . . . . . . . . . . . . . . . . . . . . . . 695.4.8 Cylinder is Unable to Move . . . . . . . . . . . . . . . . . . 70

6 System Improvements - Verifying Measurements 71

6.1 Pump Pressure Margin Reduction . . . . . . . . . . . . . . . . . . 716.2 Pump Saturation . . . . . . . . . . . . . . . . . . . . . . . . . . . . 736.3 Step Response . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 746.4 Short Duty Cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . 75

7 Summary & Conclusions 81

8 Future Work 83

Bibliography 85

A Hydraulic Schematic of a L90LS Valve 87

xii Contents

Contents 1

List of Figures

2.1 Constant flow power figure . . . . . . . . . . . . . . . . . . . . . . 112.2 Constant pressure power figure . . . . . . . . . . . . . . . . . . . . 122.3 Load sensing power figure . . . . . . . . . . . . . . . . . . . . . . . 132.4 Common pre compensators [5] . . . . . . . . . . . . . . . . . . . . 142.5 Pre compensators with anti saturation [5] . . . . . . . . . . . . . . 152.6 Post compensators [5] . . . . . . . . . . . . . . . . . . . . . . . . . 162.7 Dispacement controlled system power figure . . . . . . . . . . . . . 182.8 Flow forces acting on a spool [4] . . . . . . . . . . . . . . . . . . . 18

3.1 P1 pump controller [5] . . . . . . . . . . . . . . . . . . . . . . . . . 223.2 Simulation model of a pressure controlled pump . . . . . . . . . . . 243.3 Simulation model of a displacement controlled pump . . . . . . . . 253.4 L90LS valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 263.5 Pressure reducer and cartridge valve . . . . . . . . . . . . . . . . . 273.6 Simulation model of the cartridge valve . . . . . . . . . . . . . . . 283.7 Main spool . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 293.8 Restriction area for the main spool . . . . . . . . . . . . . . . . . . 293.9 Simulation model of the main spool . . . . . . . . . . . . . . . . . 313.10 Pre compensator . . . . . . . . . . . . . . . . . . . . . . . . . . . . 323.11 Restriction area for a common pre compensator . . . . . . . . . . . 323.12 Simulation model of a pre compensator . . . . . . . . . . . . . . . 333.13 Pre compensator with anti saturation . . . . . . . . . . . . . . . . 343.14 Restriction area for a pre compensator with anti saturation . . . . 343.15 Simulation model of a pre compensator with anti saturation . . . . 363.16 Simulation model of a ∆pp limiter . . . . . . . . . . . . . . . . . . 363.17 Simulation model of an actuator . . . . . . . . . . . . . . . . . . . 373.18 A load sensing system with common pre compensators . . . . . . . 383.19 A load sensing system with pre compensators with anti saturation 393.20 A displacement controlled system with common pre compensators 403.21 A displacement controlled system with flow sharing capabilities . . 41

4.1 Setup during laboratory tests . . . . . . . . . . . . . . . . . . . . . 444.2 Plugged connection between lspb and t2b/t3b . . . . . . . . . . . 444.3 ∆p/q test using common pre compensators . . . . . . . . . . . . . 454.4 ∆p/q test using pre compensators with anti saturation . . . . . . . 464.5 The influence of flow forces . . . . . . . . . . . . . . . . . . . . . . 464.6 The affects of flow forces in laboratory and simulation . . . . . . . 474.7 Main spool step response . . . . . . . . . . . . . . . . . . . . . . . 474.8 Zettelmeyer 802 Si [5] . . . . . . . . . . . . . . . . . . . . . . . . . 484.9 P1 pump validation . . . . . . . . . . . . . . . . . . . . . . . . . . 494.10 Pressure build up in the load pressure feedback pipe . . . . . . . . 504.11 Pressure losses in the pipe connecting the pump and the valve . . . 514.12 A saturated situation with common pre compensators in the test rig 524.13 A saturated situation with common pre compensators in simulation 52

2 Contents

4.14 A saturated situation with pre compensators with anti saturation . 534.15 Step response made in the test rig . . . . . . . . . . . . . . . . . . 544.16 Step response made in the simulation model . . . . . . . . . . . . . 54

5.1 Displacement controlled system [5] . . . . . . . . . . . . . . . . . . 555.2 Flow field when using common pre compensators . . . . . . . . . . 575.3 Load sensing and displacement controlled power figure . . . . . . . 595.4 System characteristics when not enough flow is delivered . . . . . . 605.5 System characteristics when too much flow is demanded . . . . . . 615.6 Flow field when using pre compensators with anti saturation . . . 625.7 Flow field when manoeuvring the main spool to its end position . . 635.8 ∆p controls the flow . . . . . . . . . . . . . . . . . . . . . . . . . . 645.9 Two actuators with different flow demand . . . . . . . . . . . . . . 655.10 Two actuators affected by flow forces . . . . . . . . . . . . . . . . . 655.11 Two actuators controlled by a position feedback . . . . . . . . . . . 665.12 Two actuators with compensator in its end position . . . . . . . . . 675.13 Two actuators with different loads and decreased restriction area . 685.14 Redesign of the quota . . . . . . . . . . . . . . . . . . . . . . . . . 685.15 Redesign of the quota and the maximal restriction area . . . . . . 695.16 Two actuators with different loads and a redisigned compensator . 69

6.1 Pump pressure margin in a load sensing system . . . . . . . . . . . 726.2 Pump pressure margin in a displacement controlled system . . . . 726.3 Pump saturation in a load sensing system . . . . . . . . . . . . . . 736.4 Pump saturation in a displacement controlled system . . . . . . . . 736.5 Step response in a load sensing system . . . . . . . . . . . . . . . . 746.6 Step response in a displacement controlled system . . . . . . . . . 746.7 Step response in both systems . . . . . . . . . . . . . . . . . . . . . 756.8 Short duty cycle [3] . . . . . . . . . . . . . . . . . . . . . . . . . . . 756.9 Command signals using a load sensing system . . . . . . . . . . . . 766.10 Actuator positions using a load sensing system . . . . . . . . . . . 766.11 Pump- and load pressure using a load sensing system . . . . . . . . 776.12 Pump pressure margin using a load sensing system . . . . . . . . . 776.13 Command signals in a short duty cycle . . . . . . . . . . . . . . . . 786.14 Actuator positions in a short duty cycle . . . . . . . . . . . . . . . 786.15 Pump- and load pressure in a short duty cycle . . . . . . . . . . . 786.16 Pump pressure margin in a short duty cycle . . . . . . . . . . . . . 786.17 Power consumption in a short duty cycle . . . . . . . . . . . . . . . 796.18 Energy consumption in a short duty cycle . . . . . . . . . . . . . . 79

7.1 Pump pressure margin in load sensing and displacement controlled 82

Contents 3

List of Tables

4.1 Measured parameters in the laboratory tests . . . . . . . . . . . . . 444.2 Measured parameters in the test rig . . . . . . . . . . . . . . . . . 48

4 Contents

Nomenclature

Ac Compensator restriction area [m2]Ac,max Maximal compensator restriction area [m2]Ac1 Compensator area exposed to control pressure [m2]Ac2 Compensator area exposed to control pressure [m2]As Main spool restriction area [m2]As,allowed Maximal allowed main spool restriction area [m2]As,max Maximal main spool restriction area [m2]c Speed of sound in oil [m/s]Cq Flow coefficient [−]d Diameter [m]Dp Pump displacement [m3/rev]F0 Preload spring force [N ]Fs Spring force [N ]Fs Flow force [N ]k Spring stiffness [N/m]l Length [m]L Length [m]np Pump shaft speed [rev/s]p Pressure [Pa]p1 Upstream pressure [Pa]p2 Downstream pressure [Pa]pl Load pressure [Pa]pl,max Maximal load pressure [Pa]po Pressure reduced by the main orifice [Pa]pp Pump pressure [Pa]pr Pressure reduced by the compensator [Pa]P Power [W ]q Flow [m3/s]qc Flow across the compensator [m3/s]qin Flow into a volume [m3/s]qout Flow out of a volume [m3/s]qp Pump flow [m3/s]

5

6 Contents

qs Flow across the main spool [m3/s]t Time [s]twave Time for a wave to travel across a volume [s]v Velocity [m/s]V Volume [m3]w Area gradient [m]xc Compensator spool position [m]xs Main spool position [m]βe Effective bulk modulus [Pa]δ Jet angle [ ◦]∆pf Pressure losses [Pa]∆pp Pump pressure margin [Pa]∆ps Pressure difference across the main spool [Pa]εp Pump cam position [−]ηvol,p Volumetric efficiency of the pump [−]λ Friction factor [−]ρ Density [kg/m3]

Chapter 1

Introduction

1.1 Background

Mobile fluid power applications of today consume more energy than necessary toachieve useful work. Since there are future demands for lower fuel consumption,more energy efficient systems need to be developed. When energy efficiency is avital issue load sensing systems are frequently utilized. Therefore it is appropriateto compare new system proposals with a load sensing system.

Most fluid power systems still utilize mechanical control. But as electric controlbecomes more common the system design has got new possibilities. Both pumpand valve could be controlled electronically which allows new control strategies.One way is to control the displacement of the pump instead of the pressure, whichis done in a load sensing system.

An earlier master thesis at Parker Hannifin [5] has shown that it is possibleto design a displacement controlled system using similar components as in a loadsensing system and improve the energy efficiency. However, this type of systemdesign raises new control problems that need to be solved in order to guaranteethe system operability under all circumstances.

1.2 Purpose

In the previous master thesis a displacement controlled system was implementedin a wheel loader application. Although it worked reasonably well there was in-sufficient knowledge of the system characteristics. In order to further develop thedisplacement controlled system such knowledge is a necessity.

The purpose of this master thesis is therefore to gain knowledge about thesystem characteristics in order to make further developments. An investigationregarding further improvement of the energy efficiency should also be made.

7

8 Introduction

1.3 Delimitations

When discussing about energy efficiency in this master thesis, it is the workinghydraulics that is referred to. Neither the steering nor the transmission has beentaken under consideration.

When designing displacement controlled systems, it might not be necessary toutilize a load sensing pressure compensated directional valve. However, in thismaster thesis no other opportunities have been taken under consideration.

The theory about post compensated valves is included in this master thesisbut further investigations have not been made. This is partly because only a precompensated valve was available in the test rig.

In the simulation models the leakage has been neglected. The simulation mod-els are only used for comparison and in that point of view the leakage will nothave any influence.

1.4 Method

The chosen approach when developing the displacement controlled system is sim-ulation. In a simulation environment all parameters can be measured and a verygood overview of the system is obtained. AMESim is the chosen simulation soft-ware in this master thesis. The software is based on components representing realphysical models and is therefore simple to use and the models easy to overview.The program comes with standard libraries containing among other things hy-draulic components.

When designing simulation models it is necessary to compare the simulationresults with measurements made on real components. In order to validate thesimulation models a laboratory platform as well as a test rig has been used. Toget a good reference, a load sensing system is built using the validated simulationmodels. The test rig can then be used to achieve proper system characteristics inthe simulation model of the load sensing system.

Since the displacement controlled system consist of similar components as theload sensing system, it can be used when developing the displacement controlledsystem. Behaviours that hardly can be noticed otherwise can be detected whilesimulating and adjustment can be made.

When comparing the displacement controlled system with a load sensing sys-tem regarding energy efficiency it is not reliable to use the simulation models.Instead a wheel loader application is used in order to get a proper comparison.

Because the final comparison is not made in the simulation software the modelscan be made fairly simple and only the dynamics of interest are taken underconsideration. The models are in some cases made in a general way in order toget an easy comparison. Therefore the simulation models should not be seen asan exact image of the real components but more as a tool to compare differentsystem designs.

1.5 Report Outline 9

1.5 Report Outline

The second chapter consists of basic theories about common fluid power systems.Different pressure compensators used in load sensing systems are also explained.The reader is then introduced to a displacement controlled system and the conceptof flow forces.

In the third chapter the design of all simulation models will be explained, atfirst separately and then together in load sensing- and displacement controlledsystems.

In the fourth chapter all simulation models will be validated. Both a laboratoryplatform and a test rig are utilized. Some important characteristics in load sensingsystems are also validated. These systems will later be used as a reference whendesigning new types of systems.

The fifth chapter will explain how displacement controlled systems could bedesigned. Control strategies and the characteristics of the system are discussed.Simulation models are used to confirm the discussions and finally a new systemproposal is introduced.

In the sixth chapter a load sensing system and the new system proposal arecompared in a wheel loader application. The performance and the energy efficiencyare considered.

The seventh and eighth chapter consists of a summary, some conclusions andfuture work.

10 Introduction

Chapter 2

Basic Theory

2.1 Mobile Fluid Power Systems

This is a short summary of the most common fluid power systems that is usedtoday and their power characteristics. All of the systems have their advantagesand disadvantages. A system that is suitable in one application can be useless inanother. To understand the power comparison between the systems it is importantto know the relation between pressure, flow and power, see equation (2.1).

P = q · p (2.1)

2.1.1 Constant Flow

Constant flow systems are the most commonly used systems in mobile applicationstoday. It uses a pump with fixed displacement and an open centre valve. Thesystem design is therefore fairly simple.

Figure 2.1: Constant flow power figure

11

12 Basic Theory

To obtain high energy efficiency in a constant flow system all of the pump flowneeds to be used. In mobile applications this is often not the case. The unusedflow will be throttled directly to the reservoir through the open centre valve fromthe current pressure level. Depending on the application and point of operationbig energy losses might occur, see figure 2.1.

A drawback with the constant flow system is that it is sensitive to actuatorinterference. It means that the speed of the lightest load will be affected by theheaviest load [1].

2.1.2 Constant Pressure

Constant pressure systems utilize a variable pump and a pressure regulator ora fixed pump and a relief valve to maintain a constant system pressure. If theactuators operate at the same pressure level as the pump the energy efficiency willbe high. Otherwise big energy losses will occur, see figure 2.2.

Unlike the constant flow system, the constant pressure system is not sensitiveto actuator interference. This implies as long as the pump can supply the systemwith enough oil. If that is not the case the pump is saturated and the heaviestload will decrease in speed or even stop [1].

Figure 2.2: Constant pressure power figure

2.1.3 Load Sensing

In mobile applications both pressure and flow tends to vary a lot during operation.Load sensing systems are equipped with a variable pump and a load pressurefeedback. That gives the opportunity to adapt both pressure and flow to what iscurrently needed by the actuators which gives high energy efficiency.

The weaknesses that might occur in a load sensing system are oscillations andslow response. Both are due to the load pressure feedback controlling the pumppressure [6].

In a load sensing system the pump pressure is continuously adjusted to thehighest load pressure plus a constant pressure margin, ∆pp, see figure 2.3. Some

2.2 Pressure Compensation 13

pressure is lost in the pipes and the valve also needs a certain pressure drop. ∆ppis set to overcome all these losses and it is a necessary energy loss to guaranteethe system operability [1].

Figure 2.3: Load sensing power figure

2.2 Pressure Compensation

The valves in load sensing systems are often equipped with pressure compensators.There are different kinds of pressure compensators but the principle is the same:To maintain a constant flow through the main spool independent of variations inload and pump pressure.

2.2.1 Common Pre Compensation

In a common pre compensated load sensing valve the compensator is placed upstream of the main spool. It acts in the same way as a pressure reducing valvewhere the reduced pressure pr acts on one side of the compensator and the loadpressure pl together with a spring on the other, see figure 2.4 [9].

If the pump pressure increases the compensator reduces its orifice area andvice versa resulting in a constant output pressure. The principle is the same forthe load pressure, when it decreases the compensator will reduce its orifice areaand vice versa.

The spring force depends on the preload, which is constant, and accordingto Hook’s law also the position of the compensator. Since the contribution fromthe preload is much bigger, the compensator position can be neglected, see equa-tion (2.2).

Fs = F0 + kxc ≈ /F0 � kxc/ ≈ F0 (2.2)

Equation (2.2) together with the force equilibrium for the compensator, equa-tion (2.3), and the flow equation gives the flow across the main spool.

14 Basic Theory

Figure 2.4: Common pre compensators [5]

F0 +Ac1pl = Ac1pr ⇔ F0 = Ac1(pr − pl) (2.3)

qs = CqAs

√

2

ρ(pr − pl) = CqAs

√

2

ρ

(

F0

Ac1

)

(2.4)

According to equation (2.4) the flow across the main spool depends on themain spool restriction area, As.

Equation (2.4) is valid as long as the pump can supply sufficient flow. In asaturated situation the pump pressure will drop resulting in the compensator withthe heaviest load will open completely. That function will then loose speed oreven stop. Functions operated simultaneously at lower pressure levels will movenormally.

2.2.2 Pre Compensation with Anti Saturation

This type of compensators has the same features as the common pre compensatorsbut also an anti saturation function. This means that all actuators can be giventhe same flow priority independent of variations in load- and pump pressure.

In compensators with anti saturation functions there is no spring keeping thepressure drop across the main spool constant. The spring force has been replacedby two pressure signals that constitute ∆pp according to figure 2.5 and equa-tion (2.5).

2.2 Pressure Compensation 15

Figure 2.5: Pre compensators with anti saturation [5]

∆pp = pp − pl,max (2.5)

Equation (2.5) together with the force equilibrium for the compensator, equa-tion (2.6), and the flow equation gives the flow across the main spool.

Ac1pp +Ac2pl = Ac1pl,max +Ac2pr ⇔ (pr − pl) =Ac1Ac2

(pp − pl,max) (2.6)

qs = CqAs

√

2

ρ(pr − pl) = CqAs

√

2

ρ

(

Ac1Ac2

)

∆pp (2.7)

According to equation (2.7) the flow across the main spool depends on themain spool restriction area, As, and the pump pressure margin, ∆pp. When thepump can supply sufficient flow, ∆pp remains constant and the flow only dependson the main spool restriction area.

In a saturated situation ∆pp will drop and according to equation (2.7) the flowacross all main spools will decrease proportionally. This is the main differencecomparing to common pre compensated valves where the heaviest function willdecrease in speed or even stop.

16 Basic Theory

2.2.3 Post Compensation

The functionality of post compensation is the same as with pre compensation withanti saturation. Both will distribute the flow between all functions in proportionto demand in a saturated situation. The pressure reduced by the main orificepo acts on one side of the compensator and the maximum load pressure pl,maxtogether with a spring on the other, see figure 2.6 and equation (2.8). The sameassumption with the spring according to equation (2.2) is done.

Figure 2.6: Post compensators [5]

Ac1po = Ac1pl,max + F0 ⇔ F0 = Ac1(po − pl,max) (2.8)

Equation (2.5) and (2.8) together with the flow equation confirm that the flowonly depends on the main spool restriction area according to equation (2.9).

qs = CqAs

√

2

ρ(pp − po) = CqAs

√

2

ρ(∆pp −

F0

Ac1) (2.9)

In a saturated situation ∆pp will drop and according to equation (2.9) the flowacross all main spools will decrease proportionally.

2.3 Displacement Controlled System 17

2.3 Displacement Controlled System

A displacement controlled system and is a non conventional fluid power systemusing similar components as a load sensing system. The difference is that it usesa displacement controlled pump without feedback instead of a pressure controlledpump with feedback, which is the case in a load sensing system. This meansthat the operator controls both the pump and the valve with the joystick. Aconsequence is that the demanded flow must be known [5].

A displacement controlled system does not suffer from the same oscillationproblems as a load sensing system because it utilizes an open control without aload pressure feedback. The open control also increases the response time of thepump. By controlling the displacement of the pump, it allows a reduction of ∆ppwhich means energy savings [6].

2.3.1 Flow Mapping

To be able to calculate the demanded flow it is necessary to know how muchflow that can pass by the valve. There are several ways to find out but theyall need knowledge of the flow capacity in the valve. When the flow is known,the displacement of the pump can be calculated and sent to the pump controlleraccording to equation (2.10). The pump will then deliver the flow demanded fromthe operator.

qp = εpDpnpηvol,p (2.10)

2.3.2 Compensation

The same compensators used in load sensing system can be used in a displacementcontrolled system to prevent load interference. Problems might occur when usingcommon pre compensated valves. If more flow is delivered from the pump thancan pass by the valve the extra flow will build pressure and the pump pressure willhit its maximum value which will result in high energy losses. Using a valve withpost compensation or a pre compensated valve with anti saturation will eliminatethese problems. The extra flow will then be shared proportional to the activeactuators.

2.3.3 Energy Savings

Load sensing systems have a fixed ∆pp to overcome the pressure losses in the sys-tem. ∆pp is set in order to guarantee the system operability under all possiblesituations. However, in some situations parts of ∆pp is throttled in the com-pensator resulting in energy losses. Using a displacement controlled system thisproblem is avoided because the system compensates for pressure losses betweenpump and valve itself. The compensator does not need to reduce the pressure andenergy efficiency will be high, see figure 2.7.

18 Basic Theory

Figure 2.7: Dispacement controlled system power figure

2.4 Flow Forces

The velocity of the oil is constant when approaching the inlet orifice. When comingtowards the outlet orifice the velocity is increased. Because the absolute pressureis constant, the static pressure will decrease. The force acting on the spool by theoutlet orifice is therefore less than the force by the inlet orifice. The resulting forceis called the flow force and it will always act in the closing direction.

Figure 2.8: Flow forces acting on a spool [4]

2.4 Flow Forces 19

The flow force is calculated according to equation (2.11) [7].

Fs = |2Cqwxs(p1 − p2)cos(δ)|+ ρlq̇ (2.11)

Flow forces are defined positive in closing direction, explaining the absolutevalue on the static part of the equation. δ in equation (2.11) is the angel of the oilwhen passing by the outlet orifice, see figure 2.8. This angel is often approximatedto 69 ◦ for small openings of the spool.

Flow forces can be a problem in valves because the position of the spool isaffected. With special geometry of the spool and the housing, the influence of flowforces can be reduced. But there are some cases when flow forces are of benefit.For example, a valve can be pressure compensated by flow forces.

20 Basic Theory

Chapter 3

Design of Simulation Models

In order to develop the displacement controlled system more knowledge aboutthe system characteristics is necessary. A way of getting such knowledge is tosimulate the whole system. Behaviours that hardly can be noticed otherwise canbe detected while simulating and adjustment can be made.

AMESim is the chosen simulation software in this master thesis. The softwareis based on components representing real physical models and is therefore simple touse and the models easy to overview. The program comes with standard librariescontaining among other things hydraulic components.

In this chapter, all simulation models that is necessary to design load sensingsystems as well as displacement controlled systems will be explained. Geometricproperties are determined from cad drawings and area curves for the spool arecalculated using Parker´s inhouse program Veber. Unknown parameters suchas flow forces and dynamic properties are determined by lab measurements, seechapter 4.

The simulation models are designed using the hydraulic component design

library which provides detailed hydraulic components. A basic hydraulic library

is also available with standard models but the dynamics is limited. That librarycan be used when dynamics have less influence and the static behaviour is ofinterest.

Because the simulation models are made in a comparative point of view theyshould not be seen as an exact image of the real components but more as a toolto compare different system designs.

3.1 Pump

A variable axial piston pump is modelled in AMESim to simulate Parker´s P1075pump. The model is a more general pump and do not have the same componentsas a real P1075 but parameters are adjusted to strive for similar behaviour. P1075can either be controlled by pressure or by displacement, therefore two versions ofthe pump controller are modelled in AMESim.

21

22 Design of Simulation Models

P1075 uses an electric pump controller called idec. To control the pump idec

uses two pressure sensors, one displacement sensor and one rotary speed sensor.The controller and the sensors can be seen in figure 3.1. Output from the controlleris a current which is sent to a solenoid acting on the control valve. The solenoidacts against a spring located on opposite side of the control valve.

The control valve is a 4 port 2 position valve which directs oil to the controlpistons. When the controller request an increase of the pump displacement thevalve is positioned according to figure 3.1. Pump pressure is then directed to thespring loaded control piston. The other control piston is connected to reservoir,resulting in an increase of pump displacement. In the other case when a decreaseof displacement is requested, pump pressure is directed to both control pistons.Because the spring loaded piston has a smaller area the resulting force will decreasethe pump displacement.

The control pistons act on the swash plate which controls the flow. The strokeof the pump pistons is dependent on the swash plate angel. An increase of theangel means larger stroke for the pump pistons which means that more flow isdelivered from the pump.

Figure 3.1: P1 pump controller [5]

3.1 Pump 23

3.1.1 Pressure Control

In the simulation model the load pressure is connected to the pump and convertedto an electric signal, see figure 3.2. ∆pp is then added to the load pressure signal.The pump pressure is also converted to an electric signal. The signals are thencompared and the result is sent to a pid controller. The gain of the controlleris adjusted to give expected behaviour, see section 4.2.1. The signal output fromthe controller is sent to a solenoid where it is transformed to a force acting on thecontrol valve.

The spring in the control valve is modelled with a mechanical spring. Thespring force at zero displacement is set to zero to avoid static error in the controlloop. When an increase of the pump displacement is requested the controller willsend a signal to the solenoid which will move the spool to the left. Volume 1 is thenconnected to pump pressure and volume 2 to the reservoir. Each of the volumesare connected to the control pistons. If the spool is moved to the right, pumppressure is connected to volume 1 and 2. This is done by adding an underlap topiston 1. Volume 1 and 3 are therefore always connected independent of the spoolposition.

Volume 1 is connected to the spring loaded control piston and volume 2 tothe other control piston. The velocity and force from the pistons are transformedto angular velocity and torque using the transformers. The constants connectedto the transformers are the lever arm between the pistons and the swash plate.The transformers are connected to rotary nodes used to synchronize the motionbetween the two control pistons. Relations between the motions is set in orderto make the pistons move the same distance but in opposite direction when theswash plates angel changes.

A third piston receives the pump pressure and acts on the swash plate bythe transformer. The purpose is to simulate how the pump pistons on the highpressure side acts on the swash plate.

The swash plate is simulated using an inertia connected to the rotary node.The inertia is also connected to an end stop of rotary motion. This is used todefine maximum angel of the swash plate. An angel sensor receives the angel ofthe swash plate and is multiplied with a gain to calculate the displacement ofthe pump. The displacement signal is sent to an ideal pump model. Instead ofsimulating pump pistons an ideal pump is used as a flow source. Pressure ripplescaused by the pump pistons has not been taken under consideration because itwill not add any dynamics and high frequency ripple will increase the simulationtime.

As a power source a diesel engine with a certain torque and speed capacityshould be used. In the model however the engine is assumed to have enough powerto always keep a given speed. Therefore the engine is modelled as a constant speedsource connected to the pump.

3.1.2 Displacement Control

Instead of using pressures in the control loop, pump displacement and requesteddisplacement are used, see figure 3.3. Pump displacement is measured from the


displacement sensor and compared with the requested displacement which is aninput to the model and the result is sent to the controller. Since the control erroris smaller the controller gain needs to be higher. The signal output is send to thesolenoid in the same way as the pressure control mode.

Figure 3.2: Simulation model of a pressure controlled pump

3.1 Pump 25

Figure 3.3: Simulation model of a displacement controlled pump


3.2 Valve

The modelled valve is Parker´s L90LS, which is a load sensing and pressure com-pensated directional valve. It is constructed for many different applications such ascranes, construction machinery and forest machinery. The valve can be equippedwith up to 12 sections. It is constructed for 320 bar system pressure and a flow of90 l/min with pressure compensators in each section.

In figure 3.4 a cross section of the valve is shown. The main spool is controlledby a pressure reducing cartridge valve. To obtain a constant pressure drop acrossthe main spool, different types of pressure compensators might be used. In thissection common pre compensators and pre compensators with anti saturation willbe explained. L90LS can also be equipped with two port relief valves in eachsection and a ∆pp limiter.

Figure 3.4: L90LS valve

3.2.1 Cartridge Valve

The L90LS valve is controlled by a pressure reducing cartridge valve called PVC25.It is an electro hydraulic valve delivering a pilot pressure to the main spool. Whenthe machine operator moves the lever a current is sent to a solenoid. The solenoidtransforms the current into a force acting on one side of the cartridge.

The cartridge valve is supplied with pressure from the pump circuit. Becauseof varying pressure levels in the pump circuit a pressure reducer is used to providethe cartridge with a constant pressure, usually 22 or 35 bar. By changing itsrestriction area the cartridge can control the downstream pressure, which acts

3.2 Valve 27

against the solenoid. There is also a small leakage in the valve to achieve stability.The downstream pressure acts as a pilot pressure on the main spool. But first it

passes through a damping orifice between the cartridge valve and the main spool,see figure 3.5.

Figure 3.5: Pressure reducer and cartridge valve

Simulation Model

The important thing when modelling the cartridge valve is that the dynamics ofthe main spool is correct. Neither the pressure reducer nor the cartridge itself addsmuch dynamics. Most of the dynamics depends on the damping orifice. Becauseof that the cartridge can be simplified.

In the simulation model a signal source represents the lever position in percent,see figure 3.6. The signal is recalculated into a pilot pressure and the preload ofthe spring controlling the main spool is added in the function box. The signal isthen divided depending on the level position being positive or negative. Finallythe pilot pressure is damped in orifice 1. To achieve correct dynamics the diameterof orifice 1 can be adjusted, see section 4.1.3.


Figure 3.6: Simulation model of the cartridge valve

3.2.2 Main Spool

To control the main spool, the cartridge provides the valve with a pilot pressure,see section 3.2.1. It acts on either side of the spool and works against a springpackage. This spring package is placed on the right side of the spool but actsin both directions. When no pilot pressure acts on the spool, the spring packagewill place the spool in neutral position. To be able to move the spool, the pilotpressure needs to overcome the preload of the spring. To avoid leakage the spoolhas a under lap before it opens.

Only a movement of the main spool in one direction will be explained, in thiscase a movement which will result in a lifting motion. If the movement was in theopposite direction the only difference is that the flow will be directed to the othermotor port resulting in a lowering motion.

When the spool is moved the load pressure holes will be connected with motorport A. The pressure in that motor port will then represent the load pressure forits section in the valve. The load pressure is taken through a channel inside thespool to volume 4, see figure 3.7. From this volume the load pressure is connectedto a compensator, see section 3.2.3 and 3.2.4, and to the pump, see section 3.1.

The pump port is also connected to port A when the spool is moved. Atfirst the flow will only pass through the control notches and when the spool hasbeen moved some more the whole ring area will open. Figure 3.8 shows how therestriction area of the spool depends on the spool position from pump port tomotor port. The spool design can be different depending on the application. Forexample, the restriction area for port A can be different compared to port B, whichis the case here.

From port A, the flow goes to the actuator, see section 3.3. When it returnsto the valve, it comes to motor port B. This port is connected to the reservoir.The oil will be throttled across the spool to the reservoir. Figure 3.8 shows howthe restriction area of the spool depends on the spool position from motor port toreservoir.

3.2 Valve 29

Figure 3.7: Main spool

0 0.2 0.4 0.6 0.8 10

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Position/max position

Are

a/m

ax a

rea

0 0.2 0.4 0.6 0.8 10

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1


Are

a/m

ax a

rea

P−AB−T

P−BA−T

Figure 3.8: Restriction area for the main spool


A small control notch is milled in the spool, see figure 3.7. When the spoolis in neutral position the control notch will connect volume 4 with the reservoir.This means that the load pressure will drop until it reaches the pressure in thereservoir. Because of the load pressure feedback, the pump pressure can also dropand remain at standby pressure.

Simulation Model

The pilot pressure is connected to volume 5 and acts on piston 3. The springpackage is represented by one spring on each side of the main spool. When thespool is moved towards one of the springs it will be compressed while the otherspring will not be affected. By setting the correct preload and stiffness on thesprings, they will work in the same way as the spring package.

When the spool is moved to the right motor port A will be connected to volume6 via an orifice in piston 4. The pressure in motor port A is the load pressure andit will be connected with the compensator via volume 6. If the spool is movedsome more the pump port is connected to motor port A and flow can pass bythe restriction in piston 5. By converting the spool position into restriction areaand hydraulic diameter using a look up table, piston 5 represents both the controlnotches and the ring area.

The hydraulic diameter is a commonly used term when handling flow in non-circular tubes and channels. Because of the complex geometry of the spool therestriction area is assumed be circular. The hydraulic diameter is thus equal to thediameter of the assumed circular restriction. This will have almost no significancein the simulation model because the flow will be turbulent anyway.

The flow will be sent from motor port A to the actuator and move the load.It will then return to motor port B, which is connected to piston 6 where the flowpasses by the restriction to the reservoir. Piston 6 also uses a look up table toconvert spool position into restriction area and hydraulic diameter.

Volume 6 is also connected to the reservoir via a variable orifice, which representthe small control notch in figure 3.7. When the main spool is in neutral positionthe variable orifice will open and the load pressure will drop. If the main spool ismoved from neutral position the variable orifice will be closed.

The mass in the model represent the mass of the spool and the end stops. Theend stops are set to represent the length of the stroke. The model also consistsof a position sensor. Even if there is no such sensor in the real valve it might beuseful to know the spool position while simulating.

3.2.3 Common Pre Compensator

At the left end of the compensator acts the load pressure together with a spring.The load pressure is connected to volume 4, see section 3.2.2, and passes througha damping orifice before it reaches the compensator. The spring is available indifferent designs depending on the stiffness. Differences in the stiffness will resultin different pressure drops across the main spool and thereby different flows, seeequation (2.4).

3.2 Valve 31

Figure 3.9: Simulation model of the main spool


When the main spool is in neutral position the load pressure is the same asthe pressure in the reservoir, see section 3.2.2. The compensator will be in its leftposition which means that no flow can pass by the restriction. When the mainspool is moved the compensator will sense the load pressure and move to the right.Flow is then allowed to pass by the restriction and the pressure after the restrictionwill act on the right side of the compensator via a channel inside the compensator,see figure 3.10.

Figure 3.10: Pre compensator

Because of the force balance for the compensator, the reduced pressure afterthe restriction will be equal to the load pressure plus the resulting spring pressure,see equation (2.3). The compensator will find an equilibrium position by changingits restriction area resulting in a pressure drop across the compensator. How therestriction area depends on the compensator position can be found in figure 3.11.

If the pump suffers from saturation the pump pressure will drop. Because thecompensator strives to maintain the reduced pressure it will open more. If thepump pressure drops below the load pressure the flow direction will be reversedand the load will drop. To prevent this there is a check valve function in thecompensator. If the compensator is moved to its right end position the checkvalve will close it. The check valve affects the restriction area of the compensatoraccording to figure 3.11.

10,90,80,70,60,50

0.5

1

1.5


Are

a/m

ax a

rea

Check valveRestriction

Figure 3.11: Restriction area for a common pre compensator

3.2 Valve 33

Simulation Model

The load pressure is represented by a pressure source. It passes by orifice 2 andacts together with a spring at piston 7. Volume 7 represents the volume betweenthe orifice and the compensator. The spring parameters can easily be changed inthe model if another spring is more adequate for the current application.

The pump pressure is also represented by a pressure source. When the com-pensator moves to the right flow can pass by the restriction in piston 8. Here thecompensator position is converted into restriction area and hydraulic diameter byusing a look up table. The pressure after the restriction acts on piston 9 on theright end of the compensator. Volume 8 represents the volume between the spooland the sleeve. Because the channel in the compensator is fairly big no orifice isneeded in the model.

After the restriction the flow passes by the check valve restriction, representedby piston 10, where the position is converted into restriction area and hydraulicdiameter. When the compensator is in control position the check valve restrictionhas no influence on the pressure. Volume 9 represents the volume between thecompensator and the main spool. Similar to the main spool, the compensator hasa mass with end stops and a position sensor.

Figure 3.12: Simulation model of a pre compensator

3.2.4 Pre Compensator with Anti Saturation

The external pressures acting on the compensator are the load pressure in thevalve section and the maximum load pressure. The load pressure is connected tothe compensator via a damping orifice. The compensator has two different areasexposed to the control pressures. The load pressure acts on the bigger one on theleft end of the compensator.

The maximum load pressure is connected to the compensator via a dampingorifice. It acts on the right end on the smaller area. On the other end of thecompensator acts the pump pressure via a damping orifice on the smaller area.The resulting pressure acting on the smaller area is the pump pressure margin,∆pp.


When the main spool is in neutral position the load pressure is the same asthe pressure in the reservoir, see section 3.2.2. The compensator will be in its leftposition which means that no flow can pass by the restriction. When the mainspool is moved the compensator will sense the load pressure and move to the right.Flow is then allowed to pass by the restriction and the pressure after the restrictionwill act on the bigger area on the right end of the compensator via a channel, seefigure 3.13.

Figure 3.13: Pre compensator with anti saturation

Because of the force balance for the compensator, the reduced pressure afterthe restriction will be equal to the load pressure plus a factor of the pump pressuremargin, see equation (2.6). The compensator will find an equilibrium position bychanging its restriction area resulting in a pressure drop across the compensator.How the restriction area depends on the compensator position can be found infigure 3.14.

10,90,80,70,60,50,40,30,20,100

0.2

0.4

0.6

0.8

1


Are

a/m

ax a

rea

Restriction

Figure 3.14: Restriction area for a pre compensator with anti saturation

The factor of the pump pressure margin can be compared to the stiffness of thespring used in a common pre compensated valve. Both will determine the pressuredrop and thus the flow across the main spool according to equation (2.4) and (2.7).If the pump pressure margin for some reason changes, it can be compared to achange in the spring stiffness.

3.2 Valve 35

Simulation Model

The compensator has been modelled as one part, but actually it consists of threeparts. The pressures acting on the left piston are the pump pressure and the loadpressure. Since the pump pressure will be higher than the load pressure in almostevery situation an assumption has been made: The left piston will always be inits left end position and thus not affect the force balance of the spool.

The right piston on the other hand will affect the force balance of the spool. Ifthe maximum load pressure is higher than the reduced pressure the right pistonwill push the spool to the left and in the opposite case pull the spool to the right.The force equilibrium for the right piston is shown in equation (3.1).

prAc2 = pl,maxAc2 (3.1)

When looking at the force equilibrium for the spool, pump pressure and loadpressure will act to the left on the smaller respectively bigger area. On the rightside the reduced pressure will act on the smaller and bigger area. The forceequilibrium for the spool is shown in equation (3.2).

plAc1 + ppAc2 = pr(Ac1 +Ac2) (3.2)

When taking the contribution from the right piston on the spool under con-sideration, equation (3.1) can be put into equation (3.2) resulting in the forceequilibrium for the compensator, see equation (3.3).

plAc1 + ppAc2 = prAc1 + pl,maxAc2 (3.3)

The load pressure is represented by a pressure source. It passes by orifice 3 andacts at piston 11, which represents the bigger area exposed to the control pressures.Volume 10 represents the volume between the orifice and the compensator. Themaximum load pressure is also represented by a pressure source. It passes by orifice4 and acts at piston 12, which represents the smaller area. Volume 11 representsthe volume inside the sleeve.

The pump pressure is represented by a pressure source. It passes by orifice 5and acts at piston 13, which represents the smaller area. Volume 12 represents thevolume that arises between the spool and the left piston. When the compensatormoves to the right flow can pass by the restriction in piston 14. The compensatorposition is converted into restriction area and hydraulic diameter by using a lookup table. The pressure after the restriction passes by orifice 6 and acts on piston15, which represents the bigger area. Volume 13 represents the volume between thespool and the sleeve and volume 14 represents the volume between the compensatorand the main spool. The compensator also has a mass with end stops and aposition sensor.

Orifice 6 is modelled as one restriction with 1.5 mm in diameter. This is asimplification because there are actually four radial holes with 1 mm in diametereach. Those holes are then connected to a channel inside the spool with 1.5 mmin diameter. The assumption has been made that the channel is limiting and thediameter of orifice 6 is set to the diameter of the channel.


Figure 3.15: Simulation model of a pre compensator with anti saturation

3.2.5 ∆pp Limiter

The ∆pp limiter is a pilot controlled pressure relief valve used to limit the differencebetween pump pressure and maximum load pressure, ∆pp. The cracking pressurefor the valve is the maximum load pressure plus a spring preload. By adjustingthe preload of the spring ∆pp can be limited.

Simulation Model

Pump pressure acts on the poppet. On the opposite side acts the maximum loadpressure together with a spring on piston 16. As long as the pump pressure is lessthan the maximum load pressure plus the spring preload the valve will be closed.If the pump pressure is higher, the valve will open and pump pressure will bethrottled to the reservoir resulting in a limitation of ∆pp.

Figure 3.16: Simulation model of a ∆pp limiter

3.3 Actuator 37

3.3 Actuator

The actuators in the system are cylinders. The cylinder is connected with twohoses from the valve and can therefore be controlled in both directions. When oilis directed to the piston side of the cylinder the stroke increases. Since the rod islocated on the opposite side less flow will be sent back to the valve.

The cylinder in the model is not designed to represent a specific cylinder.Instead it can be seen as a general cylinder. A mass model is used to define themass and the stroke of the piston. An appropriate value of the viscous friction hasbeen estimated in order to eliminate oscillations and speed up the simulation. Asignal source is transformed into a force connected to the mass. The signal canbe positive or negative resulting in a force that pulls or pushes the piston. Thecylinder is simulated with a simple model from the standard hydraulic library.

Figure 3.17: Simulation model of an actuator

3.4 Load Sensing Systems

In this section, the simulation models explained earlier in this chapter will beutilized to build a load sensing system. To get a better overview of the system, thesuper component function in AMESim is used. This means that all components arehidden in an icon representing the simulation model. To get a better understandingof the system, it will be explained what happens when the operator tilts its lever.

When the signal source in the lever is activated an electric signal is transformedto a pilot pressure acting on the main spool. The main spool will then move andthe load pressure will be in contact with the compensator and the shuttle valve.If no other pressure acts on the shuttle valve it will send the load pressure to thepump via the load pressure feedback pipe. How to achieve a correct pressure buildup in the pipe can be seen in section 4.2.2

In the pump the controller makes sure that the pump pressure margin is ob-tained by increasing the displacement. Flow is now delivered to the valve via thepipe connecting the pump and the valve. To achieve correct pressure losses in thepipe the diameter and the length are adjusted, see section 4.2.3.

By changing its restriction area, the compensator reduces the pressure beforethe oil comes to the main spool. There is also a pressure drop across the mainspool before the oil reaches the actuator. In the simulation model the actuatorsare of the same size in order to simplify the model. When the actuator is moving,oil from the piston rod side will be throttled across the main spool and finallyreach the reservoir.


The maximal load pressure will also affect the cracking pressure for the ∆pplimiter, referred to as the pls valve. The system also consists of a PLS limiter,which limits the load pressure and therefore also the load pressure feedback con-nected to the pump. If the system is equipped with a pre compensating valve withanti saturation, the maximal load pressure will be connected to the compensatoras well.

Figure 3.18: A load sensing system with common pre compensators

3.4 Load Sensing Systems 39

Figure 3.19: A load sensing system with pre compensators with anti saturation


3.5 Displacement Controlled Systems

A model of a displacement controlled system equipped with common pre compen-sators can be seen in figure 3.20. The only difference compared to the model ofa load sensing system with common pre compensators, see figure 3.18, is that theload pressure feedback is removed and the displacement of the pump is controlleddirectly by the operator via a flow map. If the system is equipped with a pre com-pensating valve with anti saturation, the maximal load pressure will be connectedto the compensator as well, see figure 3.21

Figure 3.20: A displacement controlled system with common pre compensators

3.5 Displacement Controlled Systems 41

Figure 3.21: A displacement controlled system with flow sharing capabilities

Chapter 4

Validation of Simulation

Models

To be able to validate simulation models and determine unknown parameters,laboratory measurements are necessary. In the laboratory, flows, pressures andpositions can be measured by sensors during testing. Parameters in the simulationmodels can then be adjusted to strive for the same result as in the laboratory test.

In this chapter, a laboratory platform as well as a test rig will be utilized tovalidate the simulation models. The models will be validated separately at firstand then together in a familiar system, in this case a load sensing system on thetest rig.

4.1 Laboratory Platform

The L90LS valve was set up in a laboratory platform. A variable orifice represent-ing the load is connected to the motor port. By adjusting the area of the orificethe pressure in the motor port can be set. The orifice is connected to either motorport A or B depending on which test being made. The oil is directed directly tothe reservoir after the orifice and not through the valve. This is because the motorport connected to the load should not be affected by the other motor port. Tocontrol the main spool a PVC25 is connected to a current source, see figure 4.1.

During the tests the pump pressure is set to a constant value of 250 bar. Anexternal pilot pressure representing the maximum load pressure is connected tothe load sensing port lspb and the connection between lspb and t2b/t3b isplugged, see figure 4.2. The pilot pressure is set to a constant value of 230 bar.The difference between the pump pressure and the maximum load pressure, ∆pp,is therefore 20 bar during the tests.

The parameters measured during the laboratory test are shown in table 4.1.

43

44 Validation of Simulation Models

Table 4.1: Measured parameters in the laboratory tests

Number Quantity Unit

1 Position of the main spool [mm]2 Current to the PVC25 [mA]3 Pilot pressure on the spool, A side [bar]4 Pilot pressure on the spool, B side [bar]5 Pump pressure [bar]6 Motor port pressure [bar]7 Reservoir pressure [bar]8 Flow [l/min]

Figure 4.1: Setup during laboratory tests

Figure 4.2: Plugged connection between lspb and t2b/t3b

4.1 Laboratory Platform 45

4.1.1 ∆p/q Test

The spool position is set to an initial position and the electrical current to thePVC25 is constant during the test. The motor port pressure is changed from250 bar to 50 bar and back to 250 bar. The pressure difference between pumppressure and load pressure, ∆p, is hence changed from 0 to 200 bar.

This is made for different initial positions of the spool with common pre com-pensators and pre compensators with anti saturation. During the ∆p/q test, thepressure drop across the valve will change but since the valve is pressure compen-sated the flow will remain constant. According to the left plot in figure 4.3 and4.4, both compensators acts as they are supposed to. The flow remains reasonablyconstant when ∆p is increased.

The same test can be simulated and the results agrees reasonably well with themeasurements, see the right plot in figure 4.3 and 4.4.

0 50 100 150 2000

10

20

30

40

50

60

70

80

90

100

Pump pressure margin [Bar]

Flo

w [l

/min

]

0 50 100 150 2000

10

20

30

40

50

60

70

80

90

100


Flo

w [l

/min

]

Figure 4.3: ∆p/q test using common pre compensators

The same test was also done with the load connected to motor port B insteadof motor port A. Since the spool is almost symmetric the result is the same.

4.1.2 Flow Forces

A ∆p/q test is also made without pressure compensators because the L90LS valvecan be used without them. In order to prevent the flow from going in the wrongdirection a check valve is used instead of a compensator. The pressure differencebetween the pump and the motor port will then occur across the main spool. Flowforces will act only on the meter in orifice since the oil is directed directly to thereservoir after the load.

The position of the spool is measured during changes in the pressure dropacross the main spool. The influence of flow forces can be seen in the left plot in


0 50 100 150 2000

20

40

60

80

100

120


Flo

w [l

/min

]

0 50 100 150 2000

20

40

60

80

100

120


Flo

w [l

/min

]

Figure 4.4: ∆p/q test using pre compensators with anti saturation

figure 4.5.In the simulation model the Kjet factor in piston 5 and 6, see figure 3.9, can be

adjusted to give the right behaviour, see the right plot in figure 4.5 and figure 4.6.

0 50 100 150 2002

2.5

3

3.5

4

4.5

5

5.5

6


Spo

ol p

ositi

on [m

m]

0 50 100 150 2002

2.5

3

3.5

4

4.5

5

5.5

6


Spo

ol p

ositi

on [m

m]

Figure 4.5: The influence of flow forces

On the meter out orifice the influence of flow forces is reduced with specialgeometry of the spool. Therefore flow forces on the meter out orifice are set tozero [8].

4.2 Test Rig 47

0 20 40 60 80 100 120 140 160 180 2002

2.5

3

3.5

4

4.5

5

5.5

6


Spo

ol p

ositi

on [m

m]

Figure 4.6: The affects of flow forces in laboratory and simulation

4.1.3 Step Response

The main spool position is set to an initial position by changing the current tothe PVC25. The current is then switched off resulting in no pilot pressure and themain spool in neutral position. When the current is switched on again a step willbe made in the current and thus in the main spool position, see figure 4.7.

By adjusting the orifice in the simulation model of the PVC25 the similarbehaviour can be seen in the simulations. Since the orifice is enough to get theexpected dynamics the cartridge valve and the pressure reducer can be neglected,see section 3.2.1.

0 0.5 1 1.5 20

1

2

3

4

5

6

Time [s]

Pos

ition

[mm

]

0 0.5 1 1.5 20

1

2

3

4

5

6

Time [s]

Pos

ition

[mm

]

Main spool position in laboratory Main spool position in simulationMain spool position in laboratory

Figure 4.7: Main spool step response

4.2 Test Rig

A compact wheel loader, Zettelmeyer 802 Si, is used as a test rig, see figure 4.8.The machine is equipped with a P1 pump and a L90LS valve. The pump can beboth pressure and displacement controlled allowing a load sensing system and a


displacement controlled system to be tested on the same machine. The actuatorsare cylinders controlling the lift and tilt functions. The valve is equipped withboth common pre compensators and pre compensators with anti saturation foreach function.

Parker´s iqan system controls the hydraulic and consists of an ecu and severali/o units. Sensors are connected to the system and iqan is used to collect data.

Figure 4.8: Zettelmeyer 802 Si [5]

The parameters measured during testing are shown in table 4.2.

Table 4.2: Measured parameters in the test rig

Quantity Location Unit

Pump pressure Pump [bar]Pump pressure Valve [bar]Load pressure Valve [bar]Motor port pressure Lift A-side [bar]Motor port pressure Lift B-side [bar]Motor port pressure Tilt A-side [bar]Motor port pressure Tilt B-side [bar]Position Lift cylinder [cm]Position Tilt cylinder [cm]Pump rotational speed Engine [rev/min]Pump cam position IQAN [%]Command signals IQAN [%]

4.2 Test Rig 49

4.2.1 Pump

The test rig is used to validate the pump. A step is done in the lever and the pumppressure is measured. Since the measurement is done on a mobile application thepressure will oscillate with a low frequency because the machine swings whena step is done. The pump will also cause pressure ripples. To keep the pumpmodel simple those phenomenon are not considered in the simulation. The mostimportant thing is the on stroke time, see figure 4.9.

The gain of the pump controller is adjusted for pressure mode and displacementmode. Since the control error is smaller in displacement mode, a higher gain isneeded compared to load sensing mode. The gain together with the orifice areafor piston 2 in the pump control valve, see figure 3.2, is adjusted to give the righton stroke time.

0.8 0.9 1 1.1 1.2 1.3 1.40

20

40

60

80

100

120

Time [s]

Pre

ssur

e [B

ar]

0.8 0.9 1 1.1 1.2 1.3 1.40

20

40

60

80

100

120

Time [s]

Pre

ssur

e [B

ar]

Pump pressure in test rig Pump pressure in simulationPump pressure in test rig

Figure 4.9: P1 pump validation

4.2.2 Load Pressure Feedback

When modelling the load pressure feedback pipe, friction, compressibility and wavedynamics could be taken into consideration. The friction is however neglectedbecause there is almost no flow in the pipe. Wave dynamics are only of interestif the time taken by a wave to travel along the pipe is longer than the samplingrate, see equation (4.1).

twave =L

c= L

√

ρ

βe≈ 5ms (4.1)

Since the communication interval with iqan is 10 ms there is no need to takewave dynamics into account. The pipe can hence be considered as a closed volume.The time it takes for the pressure to be built up in a closed volume depends onthe bulk modulus and the volume, see equation (4.2).

dp

dt= Σqin

βeV

(4.2)


The pressure build up can be validated by measurements on the test rig. If astep is done in the command signal pressure will be built up in the load pressurefeedback. By plotting the load pressure as a function of time the pressure buildup can be seen in figure 4.10.

In the simulation model the compressibility of the fluid and expansion of thepipe wall are taken into account by using an effective bulk modulus. This iscalculated based on the wall thickness and Young’s modulus for the wall material.The length and diameter gives the volume of the pipe. A comparison between themodel and the measurement can be seen in figure 4.10.

1 1.1 1.2 1.3 1.40

20

40

60

80

100

Time [s]

Pre

ssur

e [B

ar]

1 1.1 1.2 1.3 1.40

20

40

60

80

100

Time [s]

Pre

ssur

e [B

ar]

Load pressure in test rig Load pressure in simulationLoad pressure in test rig

Figure 4.10: Pressure build up in the load pressure feedback pipe

4.2.3 Pressure Losses

When modelling the pipe between the pump and the valve pressure losses mustbe kept in mind. Pressure losses arise due to friction between the fluid and thewall of the pipe and the friction in the fluid. Pressure losses due to friction canbe calculated according to equation (4.3). Also one time losses and losses due toa disturbance source will arise but they are not considered in this master thesis.

∆pf = λl

d

ρv2

2(4.3)

The test rig can be used to validate pressure losses. The difference betweenthe pressure at the pump and at the valve are pressure losses, ∆pf . By increasingthe flow, ∆pf as a function of the flow is obtained, see figure 4.11.

In the simulation model, the length and diameter of the pipe can be changedto achieve appropriate pressure losses, see figure 4.11. The friction factor λ iscalculated by AMESim and the density is considered constant.

Pressure losses also occur in the pipes connecting the valve and the actuators.This is however not considered in the simulation model because the pressure inthe motor ports is measured next to the valve on the test rig. To compensate forthis simplification the force acting on the cylinder might be made a little biggerin the model.


2 3 4 5 60

50

100

150

Time [s]

Flo

w [l

/min

]

2 3 4 5 60

5

10

Time [s]

Pre

ssur

e [B

ar]

2 3 4 5 60

50

100

150

Time [s]

Flo

w [l

/min

]

2 3 4 5 60

5

10

Time [s]

Pre

ssur

e [B

ar]

Flow in test rig

Flow in simulation

Pressure losses in test rig

Pressure losses in simulation

Figure 4.11: Pressure losses in the pipe connecting the pump and the valve

The compressibility in the pump- and actuator pipe is difficult to validate butappropriate values has been estimated. Wave dynamics has not been taken intoaccount for the same reason as in section 4.2.2.

4.3 Load Sensing Systems

To make sure that the simulation models of the load sensing systems act as theyare supposed to, two different tests are made in the test rig and compared withthe simulation models. The same tests are also made in the test rig using adisplacement controlled system, see chapter 6.

Since the simulation models are used to evaluate the system characteristicsit cannot be compared exactly to the measurements done on the test rig. Theimportant thing is that the behaviour is correct. For example, in the test rigthe cylinders have different diameters resulting in different flow demand for thesame velocity. In the simulation models the cylinders are identical to simplify thecomparison between the two functions.

4.3.1 Pump Saturation

When the pump is saturated different system behaviour can be expected dependingon what compensator being used. If the valve is equipped with common precompensators actuator interference will occur, see section 2.2.1. When using precompensators with anti saturation that problem can be avoided, see section 2.2.2.

A way to test this on the test rig is to increase the lever position to bothfunctions until the pump cannot supply sufficient flow. By continuing to increasethe command signal the behaviour in a saturated situation is shown. The heaviestload is represented by the lift function and the lightest load by the tilt function.The first test is made with common pre compensators.


0 1 2 3 4 5 60

20

40

60

80

100

120

Time [s]

Leve

r [%

]

0 1 2 3 4 5 60

0.1

0.2

0.3

0.4

Time [s]

Pos

ition

[m]

Lift and tilt command Lift actuator positionTilt actuator position

Figure 4.12: A saturated situation with common pre compensators in the test rig

As seen in figure 4.12, both functions will move with the same velocity untilmaximal flow from the pump is delivered. The velocity of the tilt will then continueto increase while the lift will loose speed. Eventually all pump flow will be deliveredto the tilt and the lift will stop completely.

When the same test is made in the simulation model the same result is achievedaccording to figure 4.13. The interesting part is however to analyse how the com-pensators work to attain this behaviour. When the pump cannot supply sufficientflow ∆pp will decrease. To prevent the reduced pressure to decrease, the compen-sator at the lift function will increase its restriction area according to figure 4.13.When the compensator no longer can maintain a constant pressure drop acrossthe main spool the flow and thus the velocity will decrease. If the reduced pres-sure drops below the load pressure plus the resulting spring pressure check valveposition will be reached and the function will stop according to section 3.2.3.

0 1 2 3 4 50

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 50

0.1

0.2

0.3

0.4

Time [s]

Pos

ition

[m]

0 1 2 3 4 50

0.5

1

1.5

Time [s]

Are

a/m

ax a

rea

0 1 2 3 4 50

5

10

15

Time [s]

Pre

ssur

e [B

ar]

Lift and tilt command

Pressure drop across lift main spool

Tilt actuator positionLift actuator position

Tilt compensator areaLift compensator area

Figure 4.13: A saturated situation with common pre compensators in simulation


This is not a problem for the tilt function because the pressure drop acrossthe compensator is higher. Its restriction area will also increase but the reducedpressure can be kept constant according to figure 4.13.

The same test is made using pre compensators with anti saturation. Accordingto figure 4.14 the same velocity is obtained, also when the pump suffers fromsaturation. If the test is made in the simulation model the expected result isobtained according to figure 4.14.

0 1 2 3 4 50

0.2

0.4

0.6

Time [s]

Pos

ition

[m]

0 1 2 3 4 50

0.2

0.4

TIme [s]

Pos

ition

[m]

0 1 2 3 4 50

0.5

1

1.5

Time [s]

Are

a/m

ax a

rea

0 1 2 3 4 50

20

40

Time [s]

Pre

ssur

e [B

ar]

Tilt actuator position in test rigLift actuator position in test rig

Tilt actuator position in simulationLift actuator position in simulation

Tilt compensator restriction areaLift compensator restriction area

Pump pressure margin

Figure 4.14: A saturated situation with pre compensators with anti saturation

As seen in figure 4.14 the compensator at the lightest section will hold itsrestriction area almost constant when the pump is saturated. The other com-pensator however will increase its restriction area in order to maintain the samepressure drop across both main spools. Hence, the same velocity for both functionsis obtained.

4.3.2 Step Response

A step is made to validate the simulation model but also to compare the perfor-mance between a load sensing system and a displacement controlled system. Thiscomparison is made in section 6.3.

The test is made with two different functions. The lift function represents theheaviest load and the tilt function the lightest. A step is made in the commandsignal for the tilt and two seconds later a step is made with the lift. Finally thecommand signal to the lift is shut off, see figure 4.15.

According to figure 4.15 the velocity of the tilt is barely affected despite ofthe pump pressure being increased. When making the same test in the simulationmodel the compensator to the tilt can be studied to find out why the velocityremains constant. As seen in figure 4.16 the behaviour of the pump pressure andthe position of the actuators correspond to the measurements on the test rig.

When the step in the tilt is made the compensator finds its equilibrium positionby changing its restriction area. A constant pressure drop across the main spool


0 1 2 3 4 5 6

0

20

40

60

80

100

TIme [s]

Leve

r [%

]

0 1 2 3 4 5 60

50

100

150

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 5 60

0.2

0.4

0.6

0.8

Time [s]

Pos

ition

[m]

Lift commandTilt command

Pump pressure

Lift actuator positionTilt actuator position

Figure 4.15: Step response made in the test rig

and thus a constant flow is then obtained. When the step in the lift is made thepump pressure increases because of the higher load. The compensator will thenfind a new equilibrium position by decrease its restriction area. It results in ahigher pressure drop across the compensator but still a constant reduced pressure.Hence the flow remains constant. When the lift is shut off the pump pressure willdecrease and the compensator will return to its previous position, see figure 4.16.

0 1 2 3 4 50

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 50

50

100

150

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 50

0.1

0.2

0.3

0.4

Time [s]

Pos

ition

[m]

0 1 2 3 4 50

2

4

6

Time [s]

Pos

ition

[mm

]

Tilt commandLift command


Pump pressure

Tilt compensator position

Figure 4.16: Step response made in the simulation model

Chapter 5

Design of Displacement

Controlled Systems

The main difference when controlling the displacement of the pump instead of thepressure is that the load pressure feedback is removed. Instead the pump receivesa requested displacement from the operator. This principle might imply higherenergy efficiency, faster response and less oscillations.

Figure 5.1: Displacement controlled system [5]

In this chapter, two different types of displacement controlled systems will bestudied. The differences between these systems are what compensator being used.

55

56 Design of Displacement Controlled Systems

At first a displacement controlled system equipped with common pre compensatorswill be discussed.

5.1 Flow Mapping

In a displacement controlled system the flow delivered from the pump need to bematched against the flow received by the valve. In this section it will be explainedwhat factors that will affect the flow calculation across the valve and from thepump.

5.1.1 Valve Control

When the operator tilts the lever a current is sent to the cartridge valve. Thecartridge valve will move the main spool with a pilot pressure proportional to thereceived current. Since the main spool acts against a spring the pilot pressure isproportional to the main spool position when the preload of the spring is overcome.A position corresponds to a certain restriction area of the main spool, see figure 3.8.

The flow across the main spool will depend on the main spool restriction areaand the pressure drop according to equation (5.1).

qs = CqAs

√

2

ρ∆ps ∝ As

√

∆ps (5.1)

The pressure drop across the main spool will be equal to the resulting springpressure of the compensator. The spring used in this master thesis will give apressure drop between 4.5 and 6.5 bar within the compensators control position.Control position of a common pre compensator means that the compensator isable to maintain a pressure drop between the above mentioned values across themain spool independent of variations in pump- and load pressure. If the pressuredrop goes below 4.5 bar check valve mode is reached and the restriction area willdecrease, see figure 3.11. There are other springs available resulting in differentpressure drops across the main spool but they are not considered in this masterthesis.

A certain lever position will result in a certain restriction area for the mainspool. But since the compensator can hold a pressure drop between 4.5 and 6.5 bardifferent flows are possible for the same lever position. In figure 5.2 the flow isplotted against the lever position. The flow for two different pressure drops isplotted, one for 4.5 bar and the other for 6.5 bar. This creates a flow field whichthe compensator can hold within its control position.

5.1 Flow Mapping 57

0 10 20 30 40 50 60 70 80 90 1000

20

40

60

80

100

120

140

160

180

Command signal [%]

Flo

w [l

/min

]

4.5 bar pressure drop6.5 bar pressure drop

Figure 5.2: Flow field when using common pre compensators

5.1.2 Pump Control

In order to deliver a flow from the pump corresponding to the flow field the leverposition needs to be transformed to a flow. This could be done by measuring thecommand signals and use known relationships between command signals, current,pilot pressure, spool position and spool area. A pressure drop within the con-trol position of the compensator could then be assumed and the flow calculatedaccording to equation (5.1).

But there is uncertainty in each step between command signal and flow. Thetolerance of the components and factors like flow forces might affect the calcula-tions. Adding up these uncertainties will make the calculations of the flow toounsure to use. Another method is required.

A load sensing system with common pre compensators can be used instead tocalculate a flow map. A ramp can be made in the command signal and the flowmeasured. The flow can then be plotted as function of the command signal. Sincethe load pressure feedback is used to control the system pressure and the pumpcan supply sufficient flow, the compensator will be in its control position and hencethe flow will be inside the flow field.

To calculate the required displacement of the pump the flow map is used totransform a lever position into a flow. The flow demand from each section issummed up and sent to the pump controller. Here the maximal possible flow iscalculated from the rotational speed of the pump and a displacement is orderedaccording to the requested flow. At the same time the lever position is sent to thevalve resulting in a proportional manoeuvring of the main spool.

When measuring the shaft speed of the pump there might be some inaccuracyin the test rig because of the analogue communication between the ecu and thepump controller. Since the volumetric efficiency of the pump depends on thecurrent system pressure it might also be a source of problem. Because of this thepump will not deliver the exact amount of flow demanded by the operator.


5.2 System Characteristics

The characteristics of a displacement controlled system are to some extent differentcompared to a load sensing system. In this section differences concerning responsebehaviour, dynamic stability and system pressure will be discussed.

5.2.1 Response Behaviour

In a load sensing system a sequence of operations must take place between com-mand signal and pump respond. At first the joystick generates a pilot pressurewhich displaces the main spool. The highest load pressure can then travel throughthe load pressure feedback and the pump changes its displacement and generatesflow.

When controlling the displacement of the pump a command signal will generatea pilot pressure and at the same time a requested displacement will be sent to thepump. Therefore the pump and the valve should react simultaneously on thecommand signal [6] [2].

The pump controller used in the test rig is not optimal for controlling thedisplacement. Because of that, there is no focus on the response behaviour in thismaster thesis. As shown in section 6.3 the response in a load sensing system isequal to the response in a displacement controlled system with the current pumpcontroller.

5.2.2 Dynamic Stability

Pressure controlled pumps operate in a pressure closed loop control mode wherethe highest load pressure can change significantly. Factors such as oil tempera-ture, natural frequencies and damping levels might affect this loop. Therefore afixed setting of the control parameters must be a compromise across all operatingconditions. Unfortunately some operating conditions might exceed the stabilitylimit resulting in an increase of the hydraulics tendency to oscillate [6] [2].

When the pump is displacement controlled it operates in an open control sincethe requested displacement is set by the operator. Problems related to the loadpressure feedback are therefore terminated and less oscillations can be expected.However, this is considered outside of this master thesis scope but might be ofinterest in future investigations.

5.2.3 System Pressure

In a load sensing system the pump controller will adjust the pump pressure inorder to maintain the pump pressure margin, ∆pp. To control the pressure thepump changes its displacement resulting in a flow. The flow will hence be changedautomatically in order to maintain ∆pp and the pump pressure will be the highestload pressure plus ∆pp.

When the pump is displacement controlled it will deliver the flow demandedfrom the operator. On its way to the actuators the oil will pass several restrictionscreating pressure losses. Also friction in the pipes will result in pressure losses,

5.3 Incorrect Flow Delivery 59

see section 4.2.3. Hence the pump pressure will be automatically adjusted to thehighest load pressure plus the pressure losses in order to deliver the demandedflow to the actuators. The similar behaviour can be seen in a constant flow systemwhen no flow is throttled through the open centre valve.

The pump pressure margin used in load sensing systems is set to a fixed level inorder to transport oil to the actuators across all flow resistances and under the mostunfavourable conditions, such as cold oil or maximal flow rate. However duringother conditions ∆pp is too high and pressure is throttled across the compensatorsresulting in a waste of energy.

Because the system pressure is automatically adjusted when controlling thedisplacement of the pump, an optimal system pressure is always achieved. Bothflow and pressure is thus adapted to what is needed by the actuators in every pos-sible situation and the energy efficiency will therefore be very good, see figure 5.3.Observe that this implies as long as the pump do not deliver too much flow, seesection 5.3.2.

Figure 5.3: Load sensing and displacement controlled power figure

5.3 Incorrect Flow Delivery

When calculating the flow map with measurements from the load sensing systemthe position of the compensator is unknown. The exact pressure drop across themain spool and thus the position in the flow field is therefore also unknown. Thismeans that the flow map could be close to the boundary of the flow field.

It is interesting to see what happens if the flow delivered from the pump isoutside the flow field. The simulation model of a displacement controlled systemequipped with common pre compensators, see figure 3.20, can be used to simulateif not enough or too much flow is delivered.


5.3.1 Not Enough Flow is Delivered

Two different loads are used in the simulation, the heaviest represent by the liftfunction and the lightest by the tilt function. The same constant lever position isused for both functions. The demanded flow is then decreased from a value insidethe flow field.

6364656667686970710

0.1

0.2

0.3

0.4

0.5

Flow [l/min]

Pos

ition

[m]

0 1 2 3 4 50

0.5

1

1.5

Time [s]

Are

a/m

ax a

rea

0 1 2 3 4 50

5

10

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 50

10

20

30

Time [s]

Pre

ssur

e [B

ar]


Lift compensator area

Pressure drop across lift main spool Pump pressure margin

Figure 5.4: System characteristics when not enough flow is delivered

As seen in figure 5.4, the velocity will be the same for the functions when thepump delivers flow according to the flow field. But as the demanded flow decreases∆pp will decrease. The compensator with the heaviest load will therefore increaseits restriction area to maintain a constant pressure drop across the main spool. As∆pp decreases the compensator is unable to hold a pressure drop within its controlarea. The function with the heaviest load will therefore decrease in speed sincethe pressure drop across the main spool is decreased. The same behaviour can beseen in a load sensing system with common pre compensators when the pump issaturated, see section 4.3.1. The consequence if the pump delivers insufficient flowis hence actuator interference.

5.3 Incorrect Flow Delivery 61

5.3.2 Too Much Flow is Delivered

It is also interesting to see what happens if the delivered flow is outside the flowfield and too much flow is delivered from the pump. To test this scenario the samesimulation is made as in section 5.3.1, but the demanded flow is now increasedfrom a value inside the flow field. Only one actuator is needed in the simulationto show what happens.

72 73 74 75 76 77 78 790

0.1

0.2

0.3

0.4

Flow [l/min]

Pos

ition

[m]

0 1 2 3 4 50

0.2

0.4

0.6

0.8

Time [s]

Are

a/m

ax a

rea

0 1 2 3 4 50

5

10

15

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 50

100

200

300

Time [s]P

ress

ure

[Bar

]

Lift actuator position Pump pressure

Lift compensator area Pressure drop across lift main spool

Figure 5.5: System characteristics when too much flow is demanded

As seen in figure 5.5, the velocity of the actuator will remain constant inde-pendent of the flow increase. Since more flow is delivered from the pump than theactuator receives pressure will be built in the pipe connecting the pump and thevalve according to equation (5.2).

dp

dt= Σ(qin − qout)

βeV

(5.2)

When the pressure is built up the compensator will reduce its restriction areain order to maintain a constant pressure drop across the main spool. The pressurein the pipe connecting the pump and the valve will increase until it exceeds thecracking pressure for the ∆pp limiter. The flow will then be throttled to thereservoir in order to maintain a maximum system pressure. In figure 5.5 thecracking pressure for the ∆pp limiter is set very high in order to see when thepressure increases. If a lower cracking pressure is set, the ∆pp limiter will throttlethe flow to the reservoir resulting in energy losses.

Demanding to much flow will result in energy losses. Since the pump pressureincreases and the load pressure remains constant unnecessary pressure losses willoccur across the compensator. Also the ∆pp limiter will contribute to the energylosses. The extra flow delivered from the pump, which not passes through thevalve, will be throttled to the reservoir resulting in high energy losses.


5.4 Displacement Controlled System with Flow

Sharing Capabilities

As explained in section 5.3 problems occur when controlling the displacement ofthe pump and using a common pre compensating valve. Too many factors areunknown and the consequence if an incorrect flow is delivered from the pump iseither energy losses or actuator interference. This is not acceptable and a bettersolution is a necessity.

The flow across the main spool is dependent of the restriction area and thepressure drop according to equation (5.1). If the area is kept constant by theoperator and the flow varies because of for example inaccuracy in the pump controlthe pressure drop also needs to vary. With common pre compensators the pressuredrop is only allowed to vary between 4.5 and 6.5 bar because of the spring stiffness,resulting in a small flow field, see figure 5.2. It is therefore desirable to have abigger flow field.

A solution to the problem is to utilize pre compensators with anti saturationinstead of common pre compensators. The flow across the main spool is thendependent on the pump pressure margin, ∆pp, instead of the resulting springpressure, see equation (2.7). The compensator will make sure that the pressuredrop across the main spool becomes exactly what is needed for the flow to passby. Hence the pressure drop across the main spool can theoretically be whateverand the flow field will be infinitely large, see figure 5.6.

0 10 20 30 40 50 60 70 80 90 1000

20

40

60

80

100

120

140

160

180

Command signal [%]

Flo

w [l

/min

]

4.5 bar pressure drop6.5 bar pressure dropSmall pressure dropLarge pressure drop

Figure 5.6: Flow field when using pre compensators with anti saturation

Another solution is to utilize post compensators. That solution is however notconsidered in this master thesis, mainly because a post compensated valve is notavailable on the test rig.

The reason why this is working is because the pump pressure is automaticallyadjusted by the system. If a higher ∆pp is needed to deliver all flow to the actuatorsthe pump pressure will increase and vice versa. Observe that this implies only ifthe pump is displacement controlled. If the pump is pressure controlled the flowis automatically adjusted in order to maintain the system pressure. Observe alsothat the flow map or inaccuracy in the pump controller no longer is a problem.The compensators will make sure that all flow delivered from the pump reachesthe actuators independent of the main spool restriction area.

5.4 Displacement Controlled System with Flow Sharing Capabilities 63

The pre compensator with anti saturation was originally designed to deal withthe pump saturation problem. When sufficient flow cannot be delivered the com-pensator will share the flow in proportion to demand. Used in a displacementcontrolled system, the compensator will share the flow in every situation by ad-justing the pressure drop across the main spool. Therefore the compensator willfrom now on be called flow sharing compensator.

In a load sensing system the flow is controlled by the restriction area of themain spool. When the pump is displacement controlled and the valve is equippedwith a flow sharing compensator this is not the case. The flow is then controlledby the pump and the valve will cause unnecessary pressure losses. It is thereforesmart to increase the restriction area and reduce the pressure drop but still get thesame flow, see equation (5.1). The main spool can hence always be manoeuvredto its end position and the pressure drop is adjusted to match the flow deliveredfrom the pump, see figure 5.7.

0 10 20 30 40 50 60 70 80 90 1000

20

40

60

80

100

120

140

160

180

Command signal [%]

Flo

w [l

/min

]

Figure 5.7: Flow field when manoeuvring the main spool to its end position

In the following sections the simulation model of a displacement controlledsystem equipped with flow sharing compensators, see figure 3.21, will be utilizedto validate the control principle. From now on the system will be referred to adisplacement controlled system with flow sharing capabilities.

5.4.1 One Actuator

In the simulation model the flow delivered from the pump is increased accordingto figure 5.8. The main spool is manoeuvred to its end position during the wholesimulation and the load is kept at a constant level. The compensator will makesure that all flow is delivered to the actuator by increasing the pressure drop acrossthe main spool. The pump pressure is therefore also increased when more flowis delivered from the pump, partly because the bigger pressure drop and partlybecause increased friction losses in the pipe connecting the pump and the valve.


1 2 3 4 50

50

100

150

200

Time [s]

Flo

w [l

/min

]

1 2 3 4 50

2

4

6

8

10

Time [s]P

ress

ure

[Bar

]

1 2 3 4 50

10

20

30

40

Time [s]

Pre

ssur

e [B

ar]

1 2 3 4 580

100

120

140

Time [s]

Pre

ssur

e [B

ar]

Delivered flow Pump pressure

Pump pressure margin Pressure drop across main spool

Figure 5.8: ∆p controls the flow

5.4.2 Two Actuators

If two actuators are used simultaneously the compensators will make sure that thepressure drop across the main spools is the same. To get a correct flow distributionto the actuators the restriction area of the main spool can be used as a flow divider.The section with the highest flow demand can be manoeuvred to its end positionas before and the other section should be manoeuvred in proportion to the flowdemand. Because the pressure drop is the same across both main spools, therestriction area will determine the flow according to equation (5.1).

This can be achieved by manipulating the command signals to the valves.When using common pre compensators a current corresponding to the commandsignal was sent to the valves. Here the command signals are recalculated in orderto manoeuvre the section with the highest flow demand completely and the otherin proportion to the flow demand. The original command signal is sent to thepump in order to deliver the flow demanded from the operator. A nice featureis that operator can decide the characteristics on the lift and tilt map. This isbecause the compensator will make sure that the pressure drop across the mainspool is exactly what is needed for the flow to pass by.

When testing this in the simulation model one actuator is activated and allflow delivered from the pump reaches that actuator. Then another actuator witha bigger flow need is activated. As seen in figure 5.9 the right amount of flow isdelivered to both actuators because the main spools are manoeuvred in proportionto the flow demand.


0 1 2 3 4 50

25

50

75

100

Time [s]

Flo

w [l

/min

]

0 1 2 3 4 50

0.5

1

1.5

Time [s]

Pos

ition

/ m

ax p

ositi

on

Tilt main spool positionLift main spool position

Demanded flow tiltDemanded flow liftDelivered flow tiltDelivered flow lift

Figure 5.9: Two actuators with different flow demand

5.4.3 Flow Forces

In the previous simulations flow forces was not taken under consideration. In areal application this is not the case. When including flow forces in the simulationthe flow delivered to the actuators might not be exactly as demanded. Whathappens is that one actuator gets more flow and the other less flow according tofigure 5.10. This is however not a big problem because it will hardly be noticed ina real application.

0 1 2 3 4 50

25

50

75

100

Time [s]

Flo

w [l

/min

]

0 1 2 3 4 50

0.5

1

1.5

Time [s]

Pos

ition

/ max

pos

ition



Figure 5.10: Two actuators affected by flow forces

This problem is not unique for displacement controlled systems. Also in loadsensing systems, flow forces will have influence on the main spool position andthus the flow. The current to the cartridge valve and the pilot pressure together


with the tolerances of the components will also affect the main spool restrictionarea.

5.4.4 Position Feedback

A way of solving this is to include a position feedback in the control loop. If theactual position of the main spool is known the current sent to the valve can beadjusted in order to achieve the reference position. If the same test as previous ismade with a position feedback a correct flow distribution is achieved according tofigure 5.11. Since a position sensor on the main spool is not available on the testrig this solution is left outside the scope of this master thesis.

0 1 2 3 4 50

25

50

75

100

125

Time [s]

Flo

w [l

/min

]

0 1 2 3 4 50

0.5

1

1.5

Time [s]

Pos

ition

/ max

pos

ition



Figure 5.11: Two actuators controlled by a position feedback

5.4.5 Different Loads

In the previous simulations the force acting on the actuators has been equal.When simulation different loads the compensators should make sure that an equalpressure drop across the main spools is obtained. But according to figure 5.12 thisis not the case. It is because the compensator at the heaviest load reaches its endposition and it is therefore unable to maintain the same pressure drop as the othercompensator.

The flow across the main spool can be calculated according to equation (2.7).Since a part of ∆pp is throttled across the main spool, the other part will bethrottled across the compensator at the heaviest load. The flow across the com-pensator can hence be calculated according to equation (5.3). Those equationsare only valid if the compensator is in its control position. When a flow sharingcompensator is utilized control position means that the compensator not reachesits end position.


0 1 2 3 4 50

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 50

0.5

1

1.5

2

Time [s]

Pos

ition

/ max

pos

ition

0 1 2 3 4 50

0.1

0.2

0.3

0.4

0.5

Time [s]

Pos

ition

[m]

0 1 2 3 4 50

5

10

Time [s]

Pre

ssur

e [B

ar]


Lift main spool positionTilt main spool position

Pressure drop across tilt main spoolPressure drop across lift main spool


Figure 5.12: Two actuators with compensator in its end position

qc = CqAc

√

2

ρ

(

1−Ac1Ac2

)

∆pp (5.3)

Since the flow across the main spool and the compensator is the same, theequations can be put together.

CqAs

√

2

ρ

(

Ac1Ac2

)

∆pp = CqAc

√

2

ρ

(

1−Ac1Ac2

)

∆pp (5.4)

When equation (5.4) is simplified a relationship between the restriction areasand the areas exposed to the control pressures is obtained according to equa-tion (5.5).

As = Ac

√

Ac2Ac1− 1 (5.5)

If the restriction area of the compensator is maximized and the compensatoris designed like it is today the maximal allowed restriction area of the main spoolcan be calculated.

As,allowed = Ac,max

√

Ac2Ac1− 1 < As,max (5.6)

According to equation (5.6) the compensator cannot be in its control positionwhile the main spool is fully open. One opportunity is to limit the manoeuvringof the main spool so that a bigger restriction area is not achieved. However, adecrease of the restriction area means an increase of the pressure drop and thus adecrease of the energy efficiency. When simulating this solution the compensatorwill not reach its end position and the same pressure drop across the main spoolsis maintained according to figure 5.13.


0 1 2 3 4 50

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 50

0.5

1

1.5

2

Time [s]

Pos

ition

/ max

pos

ition

0 1 2 3 4 50

0.1

0.2

0.3

0.4

0.5

Time [s]

Pos

ition

[m]

0 1 2 3 4 50

10

20

30

Time [s]

Pre

ssur

e [B

ar]





Figure 5.13: Two actuators with different loads and decreased restriction area

5.4.6 Redesign of the Flow Sharing Compensator

Another possibility is to redesign the compensator. As a first step the areasexposed to the control pressures can be changed. The restriction area of the mainspool is plotted against the quota of the areas exposed to the control pressures.

In the left plot in figure 5.14, it can be seen how a redesign will affect themaximal restriction area of the main spool. If the quota is decreased, the mainspool can be manoeuvred to its end position. The interesting part is however howthis will affect the pressure drop across the valve. The flow is kept at a constantlevel and ∆p is plotted against the main spool restriction area.

0.2 0.3 0.4 0.5 0.6 0.7 0.80

50

100

150

Ac1/Ac2 [−]

Mai

n sp

ool r

estr

ictio

n ar

ea [%

]

25 50 75 1000

0.2

0.4

0.6

0.8

1

Main spool restriction area [%]

Pre

ssur

e [−

]

equation (5.5)allowed restriction area for main spoolmaximal restriction area for main spool

pressure drop across main spoolallowed restriction area for main spoolmaximal restriction area for main spool

Figure 5.14: Redesign of the quota

The right plot in figure 5.14 shows that a decrease of the quota and thus anincrease of the restriction area implies a lower ∆p for the same flow.


A second step in the redesign could be to increase the maximal restrictionarea of the compensator. A new optimal point can be found in the left plot infigure 5.15 and the pressure drop would decrease even more according to the rightplot in figure 5.15.

0.2 0.3 0.4 0.5 0.6 0.7 0.80

50

100

150

Ac1/Ac2 [−]

Mai

n sp

ool r

estr

ictio

n ar

ea [%

]

25 50 75 1000

0.2

0.4

0.6

0.8

1

Main spool restriction area [%]

Pre

ssur

e [−

]

equation (5.5)maximal restriction area for main spoolequation (5.5) with increased area

pressure drop with present areamaximal restriction area for main spoolpressure drop with increased area

Figure 5.15: Redesign of the quota and the maximal restriction area

A redesign of the compensator would give additional energy savings but sincethe compensator available on the test rig is of original design this is not furtherinvestigated in this master thesis. When comparing figure 5.16 and 5.13, it showsthe same performance but with a decreased pressure drop across the main spool.

0 1 2 3 4 50

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 50

0.5

1

1.5

2

Time [s]

Pos

ition

/ max

pos

ition

0 1 2 3 4 50

5

10

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 50

0.1

0.2

0.3

0.4

0.5

Time [s]

Pos

ition

[m]





Figure 5.16: Two actuators with different loads and a redisigned compensator

5.4.7 Lowering Motion

When making a lowering motion the velocity of the actuator is determined by themeter out restriction. This is due to the gravitational force pushing the actuator.


Because of that, the main spool is controlled in proportional to the commandsignal when making a lowering motion. To avoid cavitations the pump will deliverenough flow to the piston rod side.

By controlling the main spool in proportional to the command signal no ad-ditional energy savings across the valve is obtained. But the energy savings con-cerning ∆pp is still there because the system will adjust the pump pressure to aminimum level.

A way of getting additional energy savings while making a lowering motionis to utilize an unloaded lowering. Oil to the piston rod side will then partly bedelivered from the reservoir instead of the pump. This solution is however notconsidered in this master thesis.

5.4.8 Cylinder is Unable to Move

Problems might occur when a cylinder reaches its end position or by some otherreason is unable to continue its movement. In a load sensing system the flow ischanged automatically to maintain the system pressure. If a cylinder is unableto continue its movement, no flow is needed to that section in order to maintainthe system pressure. Hence, the other actuators will not be affected and the rightamount of flow will be delivered from the pump.

In a displacement controlled system the pump will continue to deliver the flowdemanded from the operator, even though a cylinder is unable to continue itsmovement. If the flow has nowhere else to go, it will go to another actuatorresulting in an increase of the speed for that actuator. This is not desirable and asolution is required.

One solution is to utilize port relief valves in the motor ports. When thesystem pressure reaches the cracking pressure for the port relief valve the flow willbe throttled to the reservoir. No extra flow will then speed up the other actuators.This solution will however result in high energy losses since the flow is throttled tothe reservoir from a high pressure level. It is also not desirable to set the crackingpressure for the port relief valves at the same level as the maximal allowed systempressure.

A better solution is to utilize a simple position sensor on the cylinder. Whenthe cylinder reaches its end position the flow demanded from the operator to thatsection can be cancelled. It means that the pump will decrease its displacementand only deliver the flow demanded to the other actuators. The flow demand iscancelled until the operator demands a movement in the opposite direction. Thissolution will save a lot of energy but is not further investigated in this masterthesis.

Chapter 6

System Improvements -

Verifying Measurements

To verify the simulation model of a displacement controlled system with flowsharing capabilities the system has been implemented in the test rig. Since thesame components as in a load sensing system is utilized no hardware changes isnecessary.

To get a proper comparison between a load sensing system and a displacementcontrolled system with flow sharing capabilities measurements has been made onthe test rig. During the tests the main spool has been manoeuvred to its end po-sition despite that it will affect the system operability. This is because the systempotential concerning energy efficiency should be highlighted. If a displacementcontrolled system with flow sharing capabilities is implemented in a commercialapplication the compensator needs to be redesigned, see section 5.4.6. If the com-pensator is redesigned the main spool could be manoeuvred to its end positionwith maintained system operability.

6.1 Pump Pressure Margin Reduction

In section 4.2.3 it was shown how the pressure losses in the pipe connecting thepump and the valve depends on the flow. In a load sensing system ∆pp is set to aconstant value resulting in unnecessary pressure losses across the compensator. Ina displacement controlled system with flow sharing capabilities ∆pp is continuouslyadjusted to a minimum level, see section 5.4. In figure 6.1, ∆pp as function ofthe flow is shown for a load sensing system and in figure 6.2 for a displacementcontrolled system with flow sharing capabilities.

71

72 System Improvements - Verifying Measurements

0 1 2 3 4 5 6 70

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 6 70

20

40

60

Time [s]

Pos

ition

[cm

]

0 1 2 3 4 5 6 70

50

100

150

Time [s]

Flo

w [l

/min

]

0 1 2 3 4 5 6 70

10

20

30

40

Time [s]P

ress

ure

[Bar

]

Lift command Lift actuator position


Figure 6.1: Pump pressure margin in a load sensing system

0 1 2 3 4 5 6 70

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 6 70

20

40

60

Time [s]

Pos

ition

[cm

]

0 1 2 3 4 5 6 70

50

100

150

Time [s]

Flo

w [l

/min

]

0 1 2 3 4 5 6 70

10

20

30

40

Time [s]

Pre

ssur

e [B

ar]

Lift command Lift actuator position


Figure 6.2: Pump pressure margin in a displacement controlled system

6.2 Pump Saturation 73

6.2 Pump Saturation

To show how a displacement controlled system with flow sharing capabilities be-haves in a saturated situation the same test as in section 4.3.1 has been made.Because the system is equipped with flow sharing compensators the flow shouldbe shared proportionally between all active functions. As seen in figure 6.3 and6.4 the system acts in the same way as a load sensing system with anti saturationbut with a lower ∆pp when the pump is not saturated.

0 1 2 3 4 5 6 70

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 6 70

20

40

60

Time [s]

Pos

ition

[cm

]

0 1 2 3 4 5 6 70

50

100

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 5 6 70

20

40

Time [s]

Pre

ssur

e [B

ar]

Lift and tilt command Tilt actuator positionLift actuator position

Pump pressure Pump pressure margin

Figure 6.3: Pump saturation in a load sensing system

0 1 2 3 4 5 6 70

50

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 6 70

20

40

60

Time [s]

Pos

ition

[cm

]

0 1 2 3 4 5 6 70

50

100

Time [s]

Pre

ssur

e [B

ar]

0 1 2 3 4 5 6 70

20

40

Time [s]

Pre

ssur

e [B

ar]

Lift and tilt command Tilt actuator positionLift actuator position

Pump pressure Pump pressure margin

Figure 6.4: Pump saturation in a displacement controlled system


6.3 Step Response

Even though the response behaviour is outside the scope of this master thesisit is important to show a similar response compared to a load sensing system.Otherwise the comparison in section 6.4 cannot be made. The same test as insection 4.3.2 is made with a displacement controlled system with flow sharingcapabilities. According to figure 6.7 the response in the systems is comparable.

0 1 2 3 4 5 60

20

40

60

80

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 60

10

20

30

40

50

60

Time [s]

Pos

ition

[cm

]



Figure 6.5: Step response in a load sensing system

0 1 2 3 4 5 60

20

40

60

80

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 60

10

20

30

40

50

60

Time [s]

Pos

ition

[cm

]



Figure 6.6: Step response in a displacement controlled system

6.4 Short Duty Cycle 75

0 1 2 3 4 5 60

20

40

60

80

100

Time [s]

Leve

r [%

]

0 1 2 3 4 5 60

10

20

30

40

50

Time [s]

Pos

ition

[cm

]

Lift command both systemTilt command both system

Lift actuator position, LSLift actuator position, New systemTilt actuator position, LSTilt actuator position, New system

Figure 6.7: Step response in both systems

6.4 Short Duty Cycle

The final test when comparing a load sensing system with a displacement con-trolled system with flow sharing capabilities is a short duty cycle for a wheelloader application. This test is suitable because it is important to compare thetwo systems under realistic circumstances. Another type of standardized test couldalso be made but since the system is implemented in a wheel loader applicationthe short duty cycle is the chosen one.

Only the working hydraulics is considered in the measurements. Neither thesteering nor the transmission has been taken under consideration.

Figure 6.8: Short duty cycle [3]


When making a short duty cycle the wheel loader drives into the gravel pileto fill the bucket. The lift and tilt functions are then increased. The wheel loaderwill then drive backwards towards the reversing point and then towards the loadreceiver. The wheel loader will empty its bucket on the load receiver by decreasingthe tilt function. Finally the wheel loader drives backwards towards the reversingpoint while the bucket is lowered in order to begin a new cycle, see figure 6.8 [3].

In order to compare a load sensing system with a displacement controlledsystem with flow sharing capabilities two identical tests need to be made. This isalmost impossible and if the tests are not identical a proper comparison cannot bedone.

A solution to the problem is to make one test with the displacement controlledsystem. Because the performance is similar the exact same test could have beenmade with a load sensing system. The only thing that differs is the pump pressuremargin, ∆pp. In a load sensing system, ∆pp is set to a constant value of 25 bar.Hence if the pump pressure is adjusted in test with the displacement controlledsystem in order to maintain a constant ∆pp of 25 bar, a similar test but with aload sensing system is obtained.

A short duty cycle test with a load sensing system has also been made to verifythat ∆pp in fact is 25 bar. As seen in figure 6.12 a constant ∆pp of 25 bar is agood approximation.

0 5 10 15 20 25 30 35−100

−50

0

50

100

Time [s]

Leve

r [%

]


Figure 6.9: Command signals using a load sensing system

0 5 10 15 20 25 30 350

0.1

0.2

0.3

0.4

0.5

Time [s]

Pos

ition

[m]


Figure 6.10: Actuator positions using a load sensing system


0 5 10 15 20 25 30 350

50

100

150

200

250

Time [s]

Pre

ssur

e [B

ar]

Pump pressureLoad pressure

Figure 6.11: Pump- and load pressure using a load sensing system

0 5 10 15 20 25 30 350

10

20

30

40

50

60

Time [s]

Pre

ssur

e [B

ar]

Pump pressure marginAverage pump pressure margin

Figure 6.12: Pump pressure margin using a load sensing system

In the following figures a short duty cycle using a displacement controlledsystem with flow sharing capabilities is shown. The corresponding test using aload sensing system is also shown in the same figures.

As seen in figure 6.18, the energy consumption can be decreased with 14 %during a short duty cycle when using a displacement controlled system instead ofa load sensing system in this application.


0 5 10 15 20 25 30 35 40−100

−50

0

50

100

Time [s]

Leve

r [%

]

Lift demandTilt demand

Figure 6.13: Command signals in a short duty cycle

0 5 10 15 20 25 30 35 400

0.1

0.2

0.3

0.4

0.5

0.6

0.7

Time [s]

Pos

ition

[m]


Figure 6.14: Actuator positions in a short duty cycle

0 5 10 15 20 25 30 35 400

50

100

150

200

250

Time [s]

Pre

ssur

e [B

ar]

Pump pressure, displacement controlled systemPump pressure, load sensing systemLoad pressure

Figure 6.15: Pump- and load pressure in a short duty cycle

0 5 10 15 20 25 30 35 400

10

20

30

40

50

Time [s]

Pre

ssur

e [B

ar]

Pump pressure margin, displacement controlled systemPump pressure margin, load sensing system

Figure 6.16: Pump pressure margin in a short duty cycle


0 5 10 15 20 25 30 35 400

5

10

15

20

25

30

Time [s]

Pow

er [k

W]

Power, displacment controlled systemPower, load sensing system

Figure 6.17: Power consumption in a short duty cycle

0 5 10 15 20 25 30 35 400

50

100

150

200

Time [s]

Ene

rgy

[kJ]

Energy, displacment controlled systemEnergy, load sensing system

Figure 6.18: Energy consumption in a short duty cycle

Chapter 7

Summary & Conclusions

A displacement controlled system is an alternative to a load sensing system whenenergy efficiency is a vital issue. Except for the pump, the exact same componentscan be utilized and no additional sensors are necessary. The increased energyefficiency is due to a lower pump pressure margin since the system will compensatefor pressure losses between pump and valve itself. Hence, no unnecessary energylosses across the compensator will occur.

Originally the idea of a displacement controlled system was to match the ma-noeuvring of the main spool with the flow delivered from the pump using commonpre compensators. Such system design would require a very good flow map andan exact knowledge of the current shaft speed and volumetric efficiency of thepump. The flow map would also be different depending on the application andthe tolerance of the components. This is not practically achievable in all kinds ofapplications.

By using flow sharing compensators instead of common pre compensators allthese problems will be eliminated and additional energy savings across the mainspool is enabled. The compensator will make sure that the pressure drop acrossthe main spool becomes exactly what is needed for the flow to pass by independentof the main spool restriction area. Hence the flow delivered from the pump doesnot need to be matched against the manoeuvring of the main spool. Neither theflow map nor the shaft speed nor the volumetric efficiency is then a problem.

Additional energy savings across the main spool can be achieved by manoeuvrethe main spool to its end position independent of the flow delivered from the pump.A minimal pressure drop across the main spool for the corresponding flow is thenobtained.

In a displacement controlled system with flow sharing capabilities there is henceno unnecessary energy losses. The compensator at the heaviest load will be com-pletely open and also the main spool with the highest flow demand. If furtherenergy savings should be obtained, additional pumps or transformers need to beutilized.

In a load sensing system there is no unnecessary pressure drop across the com-pensator when maximal flow is delivered from the pump. This is because the high

81

82 Summary & Conclusions

pressure losses in the pipe connecting the pump and the valve, see section 4.2.3.If the operator manoeuvres the spool to its end position there is no unnecessarypressure losses across the main spool either. Hence when maximal flow is deliv-ered and the main spool is completely open, the energy efficiency in a load sensingsystem will be equal to a displacement controlled system with flow sharing capabil-ities. Otherwise, the energy efficiency will be higher in a displacement controlledsystem.

0 10 20 30 40 50 60 70 80 90 100 1100

5

10

15

20

25

30

35

40

Flow [l/min]

Pre

ssur

e [B

ar]

Pump pressure margin, displacement controlled systemPump pressure margin, load sensing system

Figure 7.1: Pump pressure margin in load sensing and displacement controlled

As seen in figure 7.1, the pump pressure margin must be the same in bothsystems if maximal flow is delivered from the pump. Otherwise it can be decreaseda lot. In an application that not uses all available pump capacity during its dutycycle, the energy efficiency can be increased if a displacement controlled systemis utilized instead of a load sensing system. The wheel loader application used inthis master thesis could for example decrease its energy consumption with 14 %during a short duty cycle.

In extreme applications using all available pump capacity the energy efficiencycannot be increased with a displacement controlled system. Minimal energy lossesacross the compensator and the main spool are already achieved with a load sensingsystem. But there are other advantages with a displacement controlled system.Even though the energy efficiency cannot be increased there is still a potential ofbetter response and less oscillations.

To summarize, the proposed displacement controlled system with flow sharingcapabilities has better or equal energy efficiency compared to a traditional loadsensing system. It has also a potential of better response and less oscillations. Toswitch from a load sensing system would require a displacement controlled pumpand an electronically controlled valve. Neither a flow map nor the accuracy of thepump nor additional sensors needs to be considered.

Chapter 8

Future Work

• The potential of a better response when controlling the displacement of thepump needs to be studied, see section 5.2.1. For example, an external pilotpressure could be attached to the control valve of the pump.

• The potential of less oscillations also needs to be studied, see section 5.2.2.

• The flow sharing compensators cannot maintain a constant pressure dropacross all main spools because it will reach its end position. An investigationregarding how to redesign the compensator needs to be made in order to geta correct flow distribution to several actuators, see section 5.4.6.

• In this master thesis only common pre compensators and flow sharing com-pensators have been studied. However, it would be interesting to study postcompensators as well.

• To deal with flow forces acting on the main spool a position feedback canbe utilized, see section 5.4.3 and 5.4.4. This will also improve the flowdistribution if several actuators are activated.

• During a lowering motion there are opportunities to get additional energysavings when controlling the displacement of the pump. For example, anunloaded lowering might be utilized, see section 5.4.7.

• When the cylinder reaches its end position a position sensor can cancel theflow delivered from the pump to that actuator, which means energy savings,see section 5.4.8.

• If unknown functions are connected to the same pump as the working hy-draulics, problems might occur when controlling the displacement. A studyregarding this problem needs to be done.

• Improve the timing between the pump and the valve by sending signals inthe right moment. Thus preventing the valve to rush ahead and open beforethe pump can provide a flow.

• Let professional drivers test the system in different applications.

83

84 Future Work

Bibliography

[1] Mobile Hydraulic Technology. Parker Hannifin, 1999.

[2] Milan Djurovic. Energiesparende Antriebssysteme für die Arbeitshydraulik

mobiler Arbeitsmaschinen Elektrohydraulishes Flow Matching. Shaker VerlagAachen, 2007.

[3] Reno Filla. Operator and machine models for dynamic simulation of construc-

tion machinery. LiTH, 2005.

[4] Institutionen för konstruktions & produktionsteknik. Formelsamling i Hy-

draulik och pneumatik. Linköpings Tekniska Högskola, 1995.

[5] Per-Anders Kumlin. Implementation of a flow controlled hydraulic work sys-

tem. Parker Hannifin, 2008.

[6] Christoph Latour. Electrohydraulic Flow Matching: The next generation of

load-sensing controls. Bosch Rexroth, 2007.

[7] Herbert E. Merritt. Hydraulic Control Systems. John Wiley & Sons, NewYork, 1967. ISBN 0-471-59617-5.

[8] Bo Nilstam. Oral, December 2008.

[9] K-E. Rydberg O. Olsson. Kompendium i hydraulik. Institutionen förkonstruktions- & produktionsteknik, 1993.

85

86 Bibliography

Appendix A

Hydraulic Schematic of a

L90LS Valve

87

88 Hydraulic Schematic of a L90LS Valve

Flow Sharing Capability

Documents

Transcript of Flow Sharing Capability