Displacement Controlled Fluid Power System with
Flow Sharing Capabilities
Robert Andersson
Mikael Axin
Division of Fluid and Mechanical Engineering Systems
Master Thesis
Department 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 University
by
Robert Andersson
Mikael Axin
LIU-IEI-TEK-A- -09/00577- -SE
Supervisors:
Björn Eriksson
IEI, Linköping University
Daniel Sundkvist
Parker Hannifin AB
Examiner:
Karl-Erik Rydberg
IEI, Linköping University
Linköping, 26 February, 2009
Datum 2009-02-26
Date
Avdelning, Institution
Division, Department
Institutionen för ekonomisk och industriell utveckling
Fluid och mekanisk systemteknik
Department of Management and Engineering
Fluid and Mechanical Engineering Systems
Språk
Language
Rapporttyp
Report category
ISBN
Svenska/Swedish
Licentiatavhandling
ISRN
Engelska/English
Examensarbete
C-uppsats
D-uppsats
Övrig rapport
—
LIU-IEI-TEK-A- -09/00577- -SE
Serietitel och serienummer ISSN
Title of series, numbering
—
URL för elektronisk version
http://www.ep.liu.se
Titel
Title
Displacement Controlled Fluid Power System with Flow Sharing Capabilities
Författare
Robert Andersson
Author
Mikael Axin
Sammanfattning
Abstract
The purpose of this master thesis is to further develop a displacement controlled
fluid power system. It uses similar components as a load sensing system but the
pump is controlled in a different way. Instead of a pressure closed loop control
mode the pump operates in an open control mode where the requested displacement 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 common
pre compensators the flow delivered from the pump needs to be matched by the
valve. A flow map would then be required and problems might occur if incorrect
flow is delivered by the pump. A solution to the problem is to utilize pre compensators with anti saturation. The flow will then be shared proportionally to the
active actuators and no flow map is needed. Since the compensators will make
sure that all flow will reach the actuators the main spool can be manoeuvred to
its end position, which allows additional energy savings.
The displacement controlled system has been designed and simulated using the
simulation software AMESim. All components in the system have been modelled
and validated using a laboratory platform. The system has also been implemented
in 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 controlled system compared to a load sensing system during a short duty cycle.
Nyckelord
Keywords
Energy efficiency, pump pressure margin, compensation, displacement control
Abstract
The purpose of this master thesis is to further develop a displacement controlled
fluid power system. It uses similar components as a load sensing system but the
pump is controlled in a different way. Instead of a pressure closed loop control mode
the pump operates in an open control mode where the requested displacement 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 common
pre compensators the flow delivered from the pump needs to be matched by the
valve. A flow map would then be required and problems might occur if incorrect
flow is delivered by the pump. A solution to the problem is to utilize pre compensators with anti saturation. The flow will then be shared proportionally to the
active actuators and no flow map is needed. Since the compensators will make
sure that all flow will reach the actuators the main spool can be manoeuvred to
its end position, which allows additional energy savings.
The displacement controlled system has been designed and simulated using the
simulation software AMESim. All components in the system have been modelled
and validated using a laboratory platform. The system has also been implemented
in 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 controlled 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. We
would like to thank the whole department for their time and effort. Our supervisor
Daniel Sundkvist can always spare a moment for discussions and he has allowed
us to go our own way during this master thesis. Anders Eliasson has helped us
a lot with the laboratory platform and the test rig. When it comes to technical
issues, Anders Lindström has been of great help.
We would also like to thank Per-Anders Kumlin at Parker Mobile Control
Division. Because of his master thesis and especially the construction of the test
rig, our work became a lot easier.
Our supervisor at the university has been Björn Eriksson. He has helped us a
lot in almost all possible ways, especially with problem regarding the report. He
has also been involved in the simulation and the system design.
Finally, we would like to thank our examiner Karl-Erik Rydberg and our opponent Karl Pettersson.
Linköping, February, 2009
Robert Andersson
Mikael Axin
vii
Contents
1 Introduction
1.1 Background . .
1.2 Purpose . . . .
1.3 Delimitations .
1.4 Method . . . .
1.5 Report Outline
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2 Basic Theory
2.1 Mobile Fluid Power Systems . . . . . . . . . . .
2.1.1 Constant Flow . . . . . . . . . . . . . .
2.1.2 Constant Pressure . . . . . . . . . . . .
2.1.3 Load Sensing . . . . . . . . . . . . . . .
2.2 Pressure Compensation . . . . . . . . . . . . .
2.2.1 Common Pre Compensation . . . . . . .
2.2.2 Pre Compensation with Anti Saturation
2.2.3 Post Compensation . . . . . . . . . . . .
2.3 Displacement Controlled System . . . . . . . .
2.3.1 Flow Mapping . . . . . . . . . . . . . .
2.3.2 Compensation . . . . . . . . . . . . . .
2.3.3 Energy Savings . . . . . . . . . . . . . .
2.4 Flow Forces . . . . . . . . . . . . . . . . . . . .
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3 Design of Simulation Models
3.1 Pump . . . . . . . . . . . . . . . .
3.1.1 Pressure Control . . . . . .
3.1.2 Displacement Control . . .
3.2 Valve . . . . . . . . . . . . . . . . .
3.2.1 Cartridge Valve . . . . . . .
3.2.2 Main Spool . . . . . . . . .
3.2.3 Common Pre Compensator
3.2.4 Pre Compensator with Anti
3.2.5 ∆pp Limiter . . . . . . . . .
3.3 Actuator . . . . . . . . . . . . . . .
3.4 Load Sensing Systems . . . . . . .
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Saturation
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x
Contents
3.5
Displacement Controlled Systems . . . . . . . . . . . . . . . . . . .
4 Validation of Simulation Models
4.1 Laboratory Platform . . . . . .
4.1.1 ∆p/q Test . . . . . . . .
4.1.2 Flow Forces . . . . . . .
4.1.3 Step Response . . . . .
4.2 Test Rig . . . . . . . . . . . . .
4.2.1 Pump . . . . . . . . . .
4.2.2 Load Pressure Feedback
4.2.3 Pressure Losses . . . . .
4.3 Load Sensing Systems . . . . .
4.3.1 Pump Saturation . . . .
4.3.2 Step Response . . . . .
40
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43
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5 Design of Displacement Controlled Systems
5.1 Flow Mapping . . . . . . . . . . . . . . . . . . . . . . . . . . . .
5.1.1 Valve Control . . . . . . . . . . . . . . . . . . . . . . . . .
5.1.2 Pump Control . . . . . . . . . . . . . . . . . . . . . . . .
5.2 System Characteristics . . . . . . . . . . . . . . . . . . . . . . . .
5.2.1 Response Behaviour . . . . . . . . . . . . . . . . . . . . .
5.2.2 Dynamic Stability . . . . . . . . . . . . . . . . . . . . . .
5.2.3 System Pressure . . . . . . . . . . . . . . . . . . . . . . .
5.3 Incorrect Flow Delivery . . . . . . . . . . . . . . . . . . . . . . .
5.3.1 Not Enough Flow is Delivered . . . . . . . . . . . . . . . .
5.3.2 Too Much Flow is Delivered . . . . . . . . . . . . . . . . .
5.4 Displacement Controlled System with Flow Sharing Capabilities
5.4.1 One Actuator . . . . . . . . . . . . . . . . . . . . . . . . .
5.4.2 Two Actuators . . . . . . . . . . . . . . . . . . . . . . . .
5.4.3 Flow Forces . . . . . . . . . . . . . . . . . . . . . . . . . .
5.4.4 Position Feedback . . . . . . . . . . . . . . . . . . . . . .
5.4.5 Different Loads . . . . . . . . . . . . . . . . . . . . . . . .
5.4.6 Redesign of the Flow Sharing Compensator . . . . . . . .
5.4.7 Lowering Motion . . . . . . . . . . . . . . . . . . . . . . .
5.4.8 Cylinder is Unable to Move . . . . . . . . . . . . . . . . .
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70
6 System Improvements - Verifying Measurements
6.1 Pump Pressure Margin Reduction . . . . . . . . .
6.2 Pump Saturation . . . . . . . . . . . . . . . . . . .
6.3 Step Response . . . . . . . . . . . . . . . . . . . .
6.4 Short Duty Cycle . . . . . . . . . . . . . . . . . . .
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7 Summary & Conclusions
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8 Future Work
83
Bibliography
85
A Hydraulic Schematic of a L90LS Valve
87
xii
Contents
Contents
1
List of Figures
2.1
2.2
2.3
2.4
2.5
2.6
2.7
2.8
Constant flow power figure . . . . . . . . .
Constant pressure power figure . . . . . . .
Load sensing power figure . . . . . . . . . .
Common pre compensators [5] . . . . . . .
Pre compensators with anti saturation [5] .
Post compensators [5] . . . . . . . . . . . .
Dispacement controlled system power figure
Flow forces acting on a spool [4] . . . . . .
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11
12
13
14
15
16
18
18
3.1
3.2
3.3
3.4
3.5
3.6
3.7
3.8
3.9
3.10
3.11
3.12
3.13
3.14
3.15
3.16
3.17
3.18
3.19
3.20
3.21
P1 pump controller [5] . . . . . . . . . . . . . . . . . . . . . . . . .
Simulation model of a pressure controlled pump . . . . . . . . . . .
Simulation model of a displacement controlled pump . . . . . . . .
L90LS valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Pressure reducer and cartridge valve . . . . . . . . . . . . . . . . .
Simulation model of the cartridge valve . . . . . . . . . . . . . . .
Main spool . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Restriction area for the main spool . . . . . . . . . . . . . . . . . .
Simulation model of the main spool . . . . . . . . . . . . . . . . .
Pre compensator . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Restriction area for a common pre compensator . . . . . . . . . . .
Simulation model of a pre compensator . . . . . . . . . . . . . . .
Pre compensator with anti saturation . . . . . . . . . . . . . . . .
Restriction area for a pre compensator with anti saturation . . . .
Simulation model of a pre compensator with anti saturation . . . .
Simulation model of a ∆pp limiter . . . . . . . . . . . . . . . . . .
Simulation model of an actuator . . . . . . . . . . . . . . . . . . .
A load sensing system with common pre compensators . . . . . . .
A load sensing system with pre compensators with anti saturation
A displacement controlled system with common pre compensators
A displacement controlled system with flow sharing capabilities . .
22
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40
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4.1
4.2
4.3
4.4
4.5
4.6
4.7
4.8
4.9
4.10
4.11
4.12
4.13
Setup during laboratory tests . . . . . . . . . . . . . . . . . . . . .
Plugged connection between lspb and t2b/t3b . . . . . . . . . . .
∆p/q test using common pre compensators . . . . . . . . . . . . .
∆p/q test using pre compensators with anti saturation . . . . . . .
The influence of flow forces . . . . . . . . . . . . . . . . . . . . . .
The affects of flow forces in laboratory and simulation . . . . . . .
Main spool step response . . . . . . . . . . . . . . . . . . . . . . .
Zettelmeyer 802 Si [5] . . . . . . . . . . . . . . . . . . . . . . . . .
P1 pump validation . . . . . . . . . . . . . . . . . . . . . . . . . .
Pressure build up in the load pressure feedback pipe . . . . . . . .
Pressure losses in the pipe connecting the pump and the valve . . .
A saturated situation with common pre compensators in the test rig
A saturated situation with common pre compensators in simulation
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52
2
Contents
4.14 A saturated situation with pre compensators with anti saturation .
4.15 Step response made in the test rig . . . . . . . . . . . . . . . . . .
4.16 Step response made in the simulation model . . . . . . . . . . . . .
53
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54
5.1
5.2
5.3
5.4
5.5
5.6
5.7
5.8
5.9
5.10
5.11
5.12
5.13
5.14
5.15
5.16
Displacement controlled system [5] . . . . . . . . . . . . . . . . .
Flow field when using common pre compensators . . . . . . . . .
Load sensing and displacement controlled power figure . . . . . .
System characteristics when not enough flow is delivered . . . . .
System characteristics when too much flow is demanded . . . . .
Flow field when using pre compensators with anti saturation . .
Flow field when manoeuvring the main spool to its end position .
∆p controls the flow . . . . . . . . . . . . . . . . . . . . . . . . .
Two actuators with different flow demand . . . . . . . . . . . . .
Two actuators affected by flow forces . . . . . . . . . . . . . . . .
Two actuators controlled by a position feedback . . . . . . . . . .
Two actuators with compensator in its end position . . . . . . . .
Two actuators with different loads and decreased restriction area
Redesign of the quota . . . . . . . . . . . . . . . . . . . . . . . .
Redesign of the quota and the maximal restriction area . . . . .
Two actuators with different loads and a redisigned compensator
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6.1
6.2
6.3
6.4
6.5
6.6
6.7
6.8
6.9
6.10
6.11
6.12
6.13
6.14
6.15
6.16
6.17
6.18
Pump pressure margin in a load sensing system . . . . . . .
Pump pressure margin in a displacement controlled system
Pump saturation in a load sensing system . . . . . . . . . .
Pump saturation in a displacement controlled system . . . .
Step response in a load sensing system . . . . . . . . . . . .
Step response in a displacement controlled system . . . . .
Step response in both systems . . . . . . . . . . . . . . . . .
Short duty cycle [3] . . . . . . . . . . . . . . . . . . . . . . .
Command signals using a load sensing system . . . . . . . .
Actuator positions using a load sensing system . . . . . . .
Pump- and load pressure using a load sensing system . . . .
Pump pressure margin using a load sensing system . . . . .
Command signals in a short duty cycle . . . . . . . . . . . .
Actuator positions in a short duty cycle . . . . . . . . . . .
Pump- and load pressure in a short duty cycle . . . . . . .
Pump pressure margin in a short duty cycle . . . . . . . . .
Power consumption in a short duty cycle . . . . . . . . . . .
Energy consumption in a short duty cycle . . . . . . . . . .
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7.1
Pump pressure margin in load sensing and displacement controlled
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Contents
3
List of Tables
4.1
4.2
Measured parameters in the laboratory tests . . . . . . . . . . . . .
Measured parameters in the test rig . . . . . . . . . . . . . . . . .
44
48
4
Contents
Nomenclature
Ac
Ac,max
Ac1
Ac2
As
As,allowed
As,max
c
Cq
d
Dp
F0
Fs
Fs
k
l
L
np
p
p1
p2
pl
pl,max
po
pp
pr
P
q
qc
qin
qout
qp
Compensator restriction area
Maximal compensator restriction area
Compensator area exposed to control pressure
Compensator area exposed to control pressure
Main spool restriction area
Maximal allowed main spool restriction area
Maximal main spool restriction area
Speed of sound in oil
Flow coefficient
Diameter
Pump displacement
Preload spring force
Spring force
Flow force
Spring stiffness
Length
Length
Pump shaft speed
Pressure
Upstream pressure
Downstream pressure
Load pressure
Maximal load pressure
Pressure reduced by the main orifice
Pump pressure
Pressure reduced by the compensator
Power
Flow
Flow across the compensator
Flow into a volume
Flow out of a volume
Pump flow
5
[m2 ]
[m2 ]
[m2 ]
[m2 ]
[m2 ]
[m2 ]
[m2 ]
[m/s]
[−]
[m]
[m3 /rev]
[N ]
[N ]
[N ]
[N/m]
[m]
[m]
[rev/s]
[P a]
[P a]
[P a]
[P a]
[P a]
[P a]
[P a]
[P a]
[W ]
[m3 /s]
[m3 /s]
[m3 /s]
[m3 /s]
[m3 /s]
6
Contents
qs
t
twave
v
V
w
xc
xs
βe
δ
∆pf
∆pp
∆ps
εp
ηvol,p
λ
ρ
Flow across the main spool
Time
Time for a wave to travel across a volume
Velocity
Volume
Area gradient
Compensator spool position
Main spool position
Effective bulk modulus
Jet angle
Pressure losses
Pump pressure margin
Pressure difference across the main spool
Pump cam position
Volumetric efficiency of the pump
Friction factor
Density
[m3 /s]
[s]
[s]
[m/s]
[m3 ]
[m]
[m]
[m]
[P a]
[ ◦]
[P a]
[P a]
[P a]
[−]
[−]
[−]
[kg/m3 ]
Chapter 1
Introduction
1.1
Background
Mobile fluid power applications of today consume more energy than necessary to
achieve useful work. Since there are future demands for lower fuel consumption,
more energy efficient systems need to be developed. When energy efficiency is a
vital issue load sensing systems are frequently utilized. Therefore it is appropriate
to compare new system proposals with a load sensing system.
Most fluid power systems still utilize mechanical control. But as electric control
becomes more common the system design has got new possibilities. Both pump
and valve could be controlled electronically which allows new control strategies.
One way is to control the displacement of the pump instead of the pressure, which
is done in a load sensing system.
An earlier master thesis at Parker Hannifin [5] has shown that it is possible
to design a displacement controlled system using similar components as in a load
sensing system and improve the energy efficiency. However, this type of system
design raises new control problems that need to be solved in order to guarantee
the system operability under all circumstances.
1.2
Purpose
In the previous master thesis a displacement controlled system was implemented
in a wheel loader application. Although it worked reasonably well there was insufficient knowledge of the system characteristics. In order to further develop the
displacement controlled system such knowledge is a necessity.
The purpose of this master thesis is therefore to gain knowledge about the
system characteristics in order to make further developments. An investigation
regarding further improvement of the energy efficiency should also be made.
7
8
1.3
Introduction
Delimitations
When discussing about energy efficiency in this master thesis, it is the working
hydraulics that is referred to. Neither the steering nor the transmission has been
taken under consideration.
When designing displacement controlled systems, it might not be necessary to
utilize a load sensing pressure compensated directional valve. However, in this
master thesis no other opportunities have been taken under consideration.
The theory about post compensated valves is included in this master thesis
but further investigations have not been made. This is partly because only a pre
compensated valve was available in the test rig.
In the simulation models the leakage has been neglected. The simulation models are only used for comparison and in that point of view the leakage will not
have any influence.
1.4
Method
The chosen approach when developing the displacement controlled system is simulation. In a simulation environment all parameters can be measured and a very
good overview of the system is obtained. AMESim is the chosen simulation software in this master thesis. The software is based on components representing real
physical models and is therefore simple to use and the models easy to overview.
The program comes with standard libraries containing among other things hydraulic components.
When designing simulation models it is necessary to compare the simulation
results with measurements made on real components. In order to validate the
simulation models a laboratory platform as well as a test rig has been used. To
get a good reference, a load sensing system is built using the validated simulation
models. The test rig can then be used to achieve proper system characteristics in
the simulation model of the load sensing system.
Since the displacement controlled system consist of similar components as the
load sensing system, it can be used when developing the displacement controlled
system. Behaviours that hardly can be noticed otherwise can be detected while
simulating and adjustment can be made.
When comparing the displacement controlled system with a load sensing system 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 models
can be made fairly simple and only the dynamics of interest are taken under
consideration. The models are in some cases made in a general way in order to
get an easy comparison. Therefore the simulation models should not be seen as
an exact image of the real components but more as a tool to compare different
system designs.
1.5 Report Outline
1.5
9
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 concept
of flow forces.
In the third chapter the design of all simulation models will be explained, at
first separately and then together in load sensing- and displacement controlled
systems.
In the fourth chapter all simulation models will be validated. Both a laboratory
platform and a test rig are utilized. Some important characteristics in load sensing
systems are also validated. These systems will later be used as a reference when
designing new types of systems.
The fifth chapter will explain how displacement controlled systems could be
designed. Control strategies and the characteristics of the system are discussed.
Simulation models are used to confirm the discussions and finally a new system
proposal is introduced.
In the sixth chapter a load sensing system and the new system proposal are
compared in a wheel loader application. The performance and the energy efficiency
are considered.
The seventh and eighth chapter consists of a summary, some conclusions and
future 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 used
today and their power characteristics. All of the systems have their advantages
and disadvantages. A system that is suitable in one application can be useless in
another. To understand the power comparison between the systems it is important
to know the relation between pressure, flow and power, see equation (2.1).
P =q·p
2.1.1
(2.1)
Constant Flow
Constant flow systems are the most commonly used systems in mobile applications
today. It uses a pump with fixed displacement and an open centre valve. The
system 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 flow
needs to be used. In mobile applications this is often not the case. The unused
flow will be throttled directly to the reservoir through the open centre valve from
the current pressure level. Depending on the application and point of operation
big energy losses might occur, see figure 2.1.
A drawback with the constant flow system is that it is sensitive to actuator
interference. It means that the speed of the lightest load will be affected by the
heaviest load [1].
2.1.2
Constant Pressure
Constant pressure systems utilize a variable pump and a pressure regulator or
a fixed pump and a relief valve to maintain a constant system pressure. If the
actuators operate at the same pressure level as the pump the energy efficiency will
be high. Otherwise big energy losses will occur, see figure 2.2.
Unlike the constant flow system, the constant pressure system is not sensitive
to actuator interference. This implies as long as the pump can supply the system
with enough oil. If that is not the case the pump is saturated and the heaviest
load 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 pressure
feedback. That gives the opportunity to adapt both pressure and flow to what is
currently needed by the actuators which gives high energy efficiency.
The weaknesses that might occur in a load sensing system are oscillations and
slow response. Both are due to the load pressure feedback controlling the pump
pressure [6].
In a load sensing system the pump pressure is continuously adjusted to the
highest 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. ∆pp
is set to overcome all these losses and it is a necessary energy loss to guarantee
the 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 in
load and pump pressure.
2.2.1
Common Pre Compensation
In a common pre compensated load sensing valve the compensator is placed up
stream of the main spool. It acts in the same way as a pressure reducing valve
where the reduced pressure pr acts on one side of the compensator and the load
pressure pl together with a spring on the other, see figure 2.4 [9].
If the pump pressure increases the compensator reduces its orifice area and
vice versa resulting in a constant output pressure. The principle is the same for
the load pressure, when it decreases the compensator will reduce its orifice area
and vice versa.
The spring force depends on the preload, which is constant, and according
to Hook’s law also the position of the compensator. Since the contribution from
the preload is much bigger, the compensator position can be neglected, see equation (2.2).
Fs = F0 + kxc ≈ /F0 kxc / ≈ F0
(2.2)
Equation (2.2) together with the force equilibrium for the compensator, equation (2.3), and the flow equation gives the flow across the main spool.
14
Basic Theory
Figure 2.4: Common pre compensators [5]
F0 + Ac1 pl = Ac1 pr ⇔ F0 = Ac1 (pr − pl )
qs = Cq As
r
2
(pr − pl ) = Cq As
ρ
s 2 F0
ρ Ac1
(2.3)
(2.4)
According to equation (2.4) the flow across the main spool depends on the
main spool restriction area, As .
Equation (2.4) is valid as long as the pump can supply sufficient flow. In a
saturated situation the pump pressure will drop resulting in the compensator with
the heaviest load will open completely. That function will then loose speed or
even stop. Functions operated simultaneously at lower pressure levels will move
normally.
2.2.2
Pre Compensation with Anti Saturation
This type of compensators has the same features as the common pre compensators
but also an anti saturation function. This means that all actuators can be given
the same flow priority independent of variations in load- and pump pressure.
In compensators with anti saturation functions there is no spring keeping the
pressure drop across the main spool constant. The spring force has been replaced
by two pressure signals that constitute ∆pp according to figure 2.5 and equation (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, equation (2.6), and the flow equation gives the flow across the main spool.
Ac1 pp + Ac2 pl = Ac1 pl,max + Ac2 pr ⇔ (pr − pl ) =
qs = Cq As
r
2
(pr − pl ) = Cq As
ρ
Ac1
(pp − pl,max )
Ac2
s 2 Ac1
∆pp
ρ Ac2
(2.6)
(2.7)
According to equation (2.7) the flow across the main spool depends on the
main spool restriction area, As , and the pump pressure margin, ∆pp . When the
pump can supply sufficient flow, ∆pp remains constant and the flow only depends
on the main spool restriction area.
In a saturated situation ∆pp will drop and according to equation (2.7) the flow
across all main spools will decrease proportionally. This is the main difference
comparing to common pre compensated valves where the heaviest function will
decrease in speed or even stop.
16
2.2.3
Basic Theory
Post Compensation
The functionality of post compensation is the same as with pre compensation with
anti saturation. Both will distribute the flow between all functions in proportion
to demand in a saturated situation. The pressure reduced by the main orifice
po acts on one side of the compensator and the maximum load pressure pl,max
together with a spring on the other, see figure 2.6 and equation (2.8). The same
assumption with the spring according to equation (2.2) is done.
Figure 2.6: Post compensators [5]
Ac1 po = Ac1 pl,max + F0 ⇔ F0 = Ac1 (po − pl,max )
(2.8)
Equation (2.5) and (2.8) together with the flow equation confirm that the flow
only depends on the main spool restriction area according to equation (2.9).
s
r
F0
2
2
(pp − po ) = Cq As
(∆pp −
)
(2.9)
qs = Cq As
ρ
ρ
Ac1
In a saturated situation ∆pp will drop and according to equation (2.9) the flow
across all main spools will decrease proportionally.
2.3 Displacement Controlled System
2.3
17
Displacement Controlled System
A displacement controlled system and is a non conventional fluid power system
using similar components as a load sensing system. The difference is that it uses
a displacement controlled pump without feedback instead of a pressure controlled
pump with feedback, which is the case in a load sensing system. This means
that the operator controls both the pump and the valve with the joystick. A
consequence is that the demanded flow must be known [5].
A displacement controlled system does not suffer from the same oscillation
problems as a load sensing system because it utilizes an open control without a
load pressure feedback. The open control also increases the response time of the
pump. By controlling the displacement of the pump, it allows a reduction of ∆pp
which means energy savings [6].
2.3.1
Flow Mapping
To be able to calculate the demanded flow it is necessary to know how much
flow that can pass by the valve. There are several ways to find out but they
all 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 controller
according to equation (2.10). The pump will then deliver the flow demanded from
the operator.
qp = εp Dp np ηvol,p
2.3.2
(2.10)
Compensation
The same compensators used in load sensing system can be used in a displacement
controlled system to prevent load interference. Problems might occur when using
common pre compensated valves. If more flow is delivered from the pump than
can pass by the valve the extra flow will build pressure and the pump pressure will
hit its maximum value which will result in high energy losses. Using a valve with
post compensation or a pre compensated valve with anti saturation will eliminate
these problems. The extra flow will then be shared proportional to the active
actuators.
2.3.3
Energy Savings
Load sensing systems have a fixed ∆pp to overcome the pressure losses in the system. ∆pp is set in order to guarantee the system operability under all possible
situations. However, in some situations parts of ∆pp is throttled in the compensator resulting in energy losses. Using a displacement controlled system this
problem is avoided because the system compensates for pressure losses between
pump and valve itself. The compensator does not need to reduce the pressure and
energy 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 coming
towards the outlet orifice the velocity is increased. Because the absolute pressure
is constant, the static pressure will decrease. The force acting on the spool by the
outlet orifice is therefore less than the force by the inlet orifice. The resulting force
is 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 = |2Cq wxs (p1 − p2 )cos(δ)| + ρlq̇
(2.11)
Flow forces are defined positive in closing direction, explaining the absolute
value on the static part of the equation. δ in equation (2.11) is the angel of the oil
when passing by the outlet orifice, see figure 2.8. This angel is often approximated
to 69 ◦ for small openings of the spool.
Flow forces can be a problem in valves because the position of the spool is
affected. With special geometry of the spool and the housing, the influence of flow
forces 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 about
the system characteristics is necessary. A way of getting such knowledge is to
simulate the whole system. Behaviours that hardly can be noticed otherwise can
be detected while simulating and adjustment can be made.
AMESim is the chosen simulation software in this master thesis. The software
is based on components representing real physical models and is therefore simple to
use and the models easy to overview. The program comes with standard libraries
containing among other things hydraulic components.
In this chapter, all simulation models that is necessary to design load sensing
systems as well as displacement controlled systems will be explained. Geometric
properties are determined from cad drawings and area curves for the spool are
calculated using Parker´s inhouse program Veber. Unknown parameters such
as flow forces and dynamic properties are determined by lab measurements, see
chapter 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 library
can be used when dynamics have less influence and the static behaviour is of
interest.
Because the simulation models are made in a comparative point of view they
should not be seen as an exact image of the real components but more as a tool
to compare different system designs.
3.1
Pump
A variable axial piston pump is modelled in AMESim to simulate Parker´s P1075
pump. The model is a more general pump and do not have the same components
as a real P1075 but parameters are adjusted to strive for similar behaviour. P1075
can either be controlled by pressure or by displacement, therefore two versions of
the 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 controller
is a current which is sent to a solenoid acting on the control valve. The solenoid
acts 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 control
pistons. When the controller request an increase of the pump displacement the
valve is positioned according to figure 3.1. Pump pressure is then directed to the
spring loaded control piston. The other control piston is connected to reservoir,
resulting in an increase of pump displacement. In the other case when a decrease
of displacement is requested, pump pressure is directed to both control pistons.
Because the spring loaded piston has a smaller area the resulting force will decrease
the pump displacement.
The control pistons act on the swash plate which controls the flow. The stroke
of the pump pistons is dependent on the swash plate angel. An increase of the
angel means larger stroke for the pump pistons which means that more flow is
delivered from the pump.
Figure 3.1: P1 pump controller [5]
3.1 Pump
3.1.1
23
Pressure Control
In the simulation model the load pressure is connected to the pump and converted
to 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 then
compared and the result is sent to a pid controller. The gain of the controller
is adjusted to give expected behaviour, see section 4.2.1. The signal output from
the controller is sent to a solenoid where it is transformed to a force acting on the
control valve.
The spring in the control valve is modelled with a mechanical spring. The
spring force at zero displacement is set to zero to avoid static error in the control
loop. When an increase of the pump displacement is requested the controller will
send a signal to the solenoid which will move the spool to the left. Volume 1 is then
connected to pump pressure and volume 2 to the reservoir. Each of the volumes
are connected to the control pistons. If the spool is moved to the right, pump
pressure is connected to volume 1 and 2. This is done by adding an underlap to
piston 1. Volume 1 and 3 are therefore always connected independent of the spool
position.
Volume 1 is connected to the spring loaded control piston and volume 2 to
the other control piston. The velocity and force from the pistons are transformed
to angular velocity and torque using the transformers. The constants connected
to the transformers are the lever arm between the pistons and the swash plate.
The transformers are connected to rotary nodes used to synchronize the motion
between the two control pistons. Relations between the motions is set in order
to make the pistons move the same distance but in opposite direction when the
swash plates angel changes.
A third piston receives the pump pressure and acts on the swash plate by
the transformer. The purpose is to simulate how the pump pistons on the high
pressure 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 to
define maximum angel of the swash plate. An angel sensor receives the angel of
the swash plate and is multiplied with a gain to calculate the displacement of
the pump. The displacement signal is sent to an ideal pump model. Instead of
simulating pump pistons an ideal pump is used as a flow source. Pressure ripples
caused by the pump pistons has not been taken under consideration because it
will not add any dynamics and high frequency ripple will increase the simulation
time.
As a power source a diesel engine with a certain torque and speed capacity
should be used. In the model however the engine is assumed to have enough power
to always keep a given speed. Therefore the engine is modelled as a constant speed
source connected to the pump.
3.1.2
Displacement Control
Instead of using pressures in the control loop, pump displacement and requested
displacement are used, see figure 3.3. Pump displacement is measured from the
24
Design of Simulation Models
displacement sensor and compared with the requested displacement which is an
input to the model and the result is sent to the controller. Since the control error
is smaller the controller gain needs to be higher. The signal output is send to the
solenoid in the same way as the pressure control mode.
Figure 3.2: Simulation model of a pressure controlled pump
3.1 Pump
Figure 3.3: Simulation model of a displacement controlled pump
25
26
3.2
Design of Simulation Models
Valve
The modelled valve is Parker´s L90LS, which is a load sensing and pressure compensated directional valve. It is constructed for many different applications such as
cranes, construction machinery and forest machinery. The valve can be equipped
with up to 12 sections. It is constructed for 320 bar system pressure and a flow of
90 l/min with pressure compensators in each section.
In figure 3.4 a cross section of the valve is shown. The main spool is controlled
by a pressure reducing cartridge valve. To obtain a constant pressure drop across
the main spool, different types of pressure compensators might be used. In this
section common pre compensators and pre compensators with anti saturation will
be explained. L90LS can also be equipped with two port relief valves in each
section 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. When
the machine operator moves the lever a current is sent to a solenoid. The solenoid
transforms the current into a force acting on one side of the cartridge.
The cartridge valve is supplied with pressure from the pump circuit. Because
of varying pressure levels in the pump circuit a pressure reducer is used to provide
the cartridge with a constant pressure, usually 22 or 35 bar. By changing its
restriction 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 of
the main spool is correct. Neither the pressure reducer nor the cartridge itself adds
much dynamics. Most of the dynamics depends on the damping orifice. Because
of 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 of
the spring controlling the main spool is added in the function box. The signal is
then divided depending on the level position being positive or negative. Finally
the pilot pressure is damped in orifice 1. To achieve correct dynamics the diameter
of orifice 1 can be adjusted, see section 4.1.3.
28
Design of Simulation Models
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 spring
package. This spring package is placed on the right side of the spool but acts
in both directions. When no pilot pressure acts on the spool, the spring package
will place the spool in neutral position. To be able to move the spool, the pilot
pressure needs to overcome the preload of the spring. To avoid leakage the spool
has a under lap before it opens.
Only a movement of the main spool in one direction will be explained, in this
case a movement which will result in a lifting motion. If the movement was in the
opposite direction the only difference is that the flow will be directed to the other
motor port resulting in a lowering motion.
When the spool is moved the load pressure holes will be connected with motor
port A. The pressure in that motor port will then represent the load pressure for
its section in the valve. The load pressure is taken through a channel inside the
spool to volume 4, see figure 3.7. From this volume the load pressure is connected
to 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. At
first the flow will only pass through the control notches and when the spool has
been moved some more the whole ring area will open. Figure 3.8 shows how the
restriction area of the spool depends on the spool position from pump port to
motor port. The spool design can be different depending on the application. For
example, the restriction area for port A can be different compared to port B, which
is the case here.
From port A, the flow goes to the actuator, see section 3.3. When it returns
to 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 how
the restriction area of the spool depends on the spool position from motor port to
reservoir.
3.2 Valve
29
Figure 3.7: Main spool
1
1
P−A
B−T
0.9
0.8
0.8
0.7
0.7
Area/max area
Area/max area
0.9
0.6
0.5
0.4
0.6
0.5
0.4
0.3
0.3
0.2
0.2
0.1
0.1
0
0
0.2
0.4
0.6
0.8
Position/max position
1
P−B
A−T
0
0
0.2
0.4
0.6
0.8
Position/max position
Figure 3.8: Restriction area for the main spool
1
30
Design of Simulation Models
A small control notch is milled in the spool, see figure 3.7. When the spool
is 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 the
reservoir. Because of the load pressure feedback, the pump pressure can also drop
and remain at standby pressure.
Simulation Model
The pilot pressure is connected to volume 5 and acts on piston 3. The spring
package is represented by one spring on each side of the main spool. When the
spool is moved towards one of the springs it will be compressed while the other
spring will not be affected. By setting the correct preload and stiffness on the
springs, 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 volume
6 via an orifice in piston 4. The pressure in motor port A is the load pressure and
it will be connected with the compensator via volume 6. If the spool is moved
some more the pump port is connected to motor port A and flow can pass by
the restriction in piston 5. By converting the spool position into restriction area
and hydraulic diameter using a look up table, piston 5 represents both the control
notches and the ring area.
The hydraulic diameter is a commonly used term when handling flow in noncircular tubes and channels. Because of the complex geometry of the spool the
restriction area is assumed be circular. The hydraulic diameter is thus equal to the
diameter of the assumed circular restriction. This will have almost no significance
in 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 flow
passes by the restriction to the reservoir. Piston 6 also uses a look up table to
convert spool position into restriction area and hydraulic diameter.
Volume 6 is also connected to the reservoir via a variable orifice, which represent
the small control notch in figure 3.7. When the main spool is in neutral position
the variable orifice will open and the load pressure will drop. If the main spool is
moved from neutral position the variable orifice will be closed.
The mass in the model represent the mass of the spool and the end stops. The
end stops are set to represent the length of the stroke. The model also consists
of a position sensor. Even if there is no such sensor in the real valve it might be
useful 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 through
a damping orifice before it reaches the compensator. The spring is available in
different designs depending on the stiffness. Differences in the stiffness will result
in different pressure drops across the main spool and thereby different flows, see
equation (2.4).
3.2 Valve
31
Figure 3.9: Simulation model of the main spool
32
Design of Simulation Models
When the main spool is in neutral position the load pressure is the same as
the pressure in the reservoir, see section 3.2.2. The compensator will be in its left
position which means that no flow can pass by the restriction. When the main
spool 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 restriction
will 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 after
the 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 changing
its restriction area resulting in a pressure drop across the compensator. How the
restriction 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 the
compensator strives to maintain the reduced pressure it will open more. If the
pump pressure drops below the load pressure the flow direction will be reversed
and the load will drop. To prevent this there is a check valve function in the
compensator. If the compensator is moved to its right end position the check
valve will close it. The check valve affects the restriction area of the compensator
according to figure 3.11.
Area/max area
1.5
Check valve
Restriction
1
0.5
0
0,5
0,6
0,7
0,8
Position/max position
0,9
1
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 and
acts together with a spring at piston 7. Volume 7 represents the volume between
the orifice and the compensator. The spring parameters can easily be changed in
the model if another spring is more adequate for the current application.
The pump pressure is also represented by a pressure source. When the compensator moves to the right flow can pass by the restriction in piston 8. Here the
compensator position is converted into restriction area and hydraulic diameter by
using a look up table. The pressure after the restriction acts on piston 9 on the
right end of the compensator. Volume 8 represents the volume between the spool
and the sleeve. Because the channel in the compensator is fairly big no orifice is
needed in the model.
After the restriction the flow passes by the check valve restriction, represented
by piston 10, where the position is converted into restriction area and hydraulic
diameter. When the compensator is in control position the check valve restriction
has no influence on the pressure. Volume 9 represents the volume between the
compensator and the main spool. Similar to the main spool, the compensator has
a 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 the
valve section and the maximum load pressure. The load pressure is connected to
the compensator via a damping orifice. The compensator has two different areas
exposed to the control pressures. The load pressure acts on the bigger one on the
left end of the compensator.
The maximum load pressure is connected to the compensator via a damping
orifice. It acts on the right end on the smaller area. On the other end of the
compensator 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 .
34
Design of Simulation Models
When the main spool is in neutral position the load pressure is the same as
the pressure in the reservoir, see section 3.2.2. The compensator will be in its left
position which means that no flow can pass by the restriction. When the main
spool 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 restriction
will act on the bigger area on the right end of the compensator via a channel, see
figure 3.13.
Figure 3.13: Pre compensator with anti saturation
Because of the force balance for the compensator, the reduced pressure after
the restriction will be equal to the load pressure plus a factor of the pump pressure
margin, see equation (2.6). The compensator will find an equilibrium position by
changing its restriction area resulting in a pressure drop across the compensator.
How the restriction area depends on the compensator position can be found in
figure 3.14.
1
Restriction
Area/max area
0.8
0.6
0.4
0.2
0
0
0,1
0,2
0,3
0,4
0,5
0,6
Position/max position
0,7
0,8
0,9
1
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 the
spring used in a common pre compensated valve. Both will determine the pressure
drop 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 a
change in the spring stiffness.
3.2 Valve
35
Simulation Model
The compensator has been modelled as one part, but actually it consists of three
parts. The pressures acting on the left piston are the pump pressure and the load
pressure. Since the pump pressure will be higher than the load pressure in almost
every situation an assumption has been made: The left piston will always be in
its 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. If
the maximum load pressure is higher than the reduced pressure the right piston
will 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).
pr Ac2 = pl,max Ac2
(3.1)
When looking at the force equilibrium for the spool, pump pressure and load
pressure will act to the left on the smaller respectively bigger area. On the right
side the reduced pressure will act on the smaller and bigger area. The force
equilibrium for the spool is shown in equation (3.2).
pl Ac1 + pp Ac2 = pr (Ac1 + Ac2 )
(3.2)
When taking the contribution from the right piston on the spool under consideration, equation (3.1) can be put into equation (3.2) resulting in the force
equilibrium for the compensator, see equation (3.3).
pl Ac1 + pp Ac2 = pr Ac1 + pl,max Ac2
(3.3)
The load pressure is represented by a pressure source. It passes by orifice 3 and
acts at piston 11, which represents the bigger area exposed to the control pressures.
Volume 10 represents the volume between the orifice and the compensator. The
maximum load pressure is also represented by a pressure source. It passes by orifice
4 and acts at piston 12, which represents the smaller area. Volume 11 represents
the volume inside the sleeve.
The pump pressure is represented by a pressure source. It passes by orifice 5
and acts at piston 13, which represents the smaller area. Volume 12 represents the
volume that arises between the spool and the left piston. When the compensator
moves to the right flow can pass by the restriction in piston 14. The compensator
position is converted into restriction area and hydraulic diameter by using a look
up table. The pressure after the restriction passes by orifice 6 and acts on piston
15, which represents the bigger area. Volume 13 represents the volume between the
spool and the sleeve and volume 14 represents the volume between the compensator
and the main spool. The compensator also has a mass with end stops and a
position sensor.
Orifice 6 is modelled as one restriction with 1.5 mm in diameter. This is a
simplification because there are actually four radial holes with 1 mm in diameter
each. Those holes are then connected to a channel inside the spool with 1.5 mm
in diameter. The assumption has been made that the channel is limiting and the
diameter of orifice 6 is set to the diameter of the channel.
36
Design of Simulation Models
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 difference
between pump pressure and maximum load pressure, ∆pp . The cracking pressure
for the valve is the maximum load pressure plus a spring preload. By adjusting
the preload of the spring ∆pp can be limited.
Simulation Model
Pump pressure acts on the poppet. On the opposite side acts the maximum load
pressure together with a spring on piston 16. As long as the pump pressure is less
than 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 be
throttled to the reservoir resulting in a limitation of ∆pp .
Figure 3.16: Simulation model of a ∆pp limiter
3.3 Actuator
3.3
37
Actuator
The actuators in the system are cylinders. The cylinder is connected with two
hoses from the valve and can therefore be controlled in both directions. When oil
is directed to the piston side of the cylinder the stroke increases. Since the rod is
located 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 the
mass and the stroke of the piston. An appropriate value of the viscous friction has
been estimated in order to eliminate oscillations and speed up the simulation. A
signal source is transformed into a force connected to the mass. The signal can
be positive or negative resulting in a force that pulls or pushes the piston. The
cylinder 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 be
utilized to build a load sensing system. To get a better overview of the system, the
super component function in AMESim is used. This means that all components are
hidden in an icon representing the simulation model. To get a better understanding
of 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 transformed
to a pilot pressure acting on the main spool. The main spool will then move and
the 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 the
pump via the load pressure feedback pipe. How to achieve a correct pressure build
up in the pipe can be seen in section 4.2.2
In the pump the controller makes sure that the pump pressure margin is obtained by increasing the displacement. Flow is now delivered to the valve via the
pipe connecting the pump and the valve. To achieve correct pressure losses in the
pipe the diameter and the length are adjusted, see section 4.2.3.
By changing its restriction area, the compensator reduces the pressure before
the oil comes to the main spool. There is also a pressure drop across the main
spool before the oil reaches the actuator. In the simulation model the actuators
are 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 finally
reach the reservoir.
38
Design of Simulation Models
The maximal load pressure will also affect the cracking pressure for the ∆pp
limiter, 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 connected to the pump. If the system is equipped with a pre compensating valve with
anti saturation, the maximal load pressure will be connected to the compensator
as 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
40
3.5
Design of Simulation Models
Displacement Controlled Systems
A model of a displacement controlled system equipped with common pre compensators can be seen in figure 3.20. The only difference compared to the model of
a load sensing system with common pre compensators, see figure 3.18, is that the
load pressure feedback is removed and the displacement of the pump is controlled
directly by the operator via a flow map. If the system is equipped with a pre compensating valve with anti saturation, the maximal load pressure will be connected
to 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
42
Design of Simulation Models
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 and
positions can be measured by sensors during testing. Parameters in the simulation
models 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 to
validate the simulation models. The models will be validated separately at first
and then together in a familiar system, in this case a load sensing system on the
test rig.
4.1
Laboratory Platform
The L90LS valve was set up in a laboratory platform. A variable orifice representing the load is connected to the motor port. By adjusting the area of the orifice
the pressure in the motor port can be set. The orifice is connected to either motor
port A or B depending on which test being made. The oil is directed directly to
the reservoir after the orifice and not through the valve. This is because the motor
port connected to the load should not be affected by the other motor port. To
control 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. An
external pilot pressure representing the maximum load pressure is connected to
the load sensing port lspb and the connection between lspb and t2b/t3b is
plugged, 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
1
2
3
4
5
6
7
8
Quantity
Position of the main spool
Current to the PVC25
Pilot pressure on the spool, A side
Pilot pressure on the spool, B side
Pump pressure
Motor port pressure
Reservoir pressure
Flow
Unit
[mm]
[mA]
[bar]
[bar]
[bar]
[bar]
[bar]
[l/min]
Figure 4.1: Setup during laboratory tests
Figure 4.2: Plugged connection between lspb and t2b/t3b
4.1 Laboratory Platform
45
∆p/q Test
4.1.1
100
100
90
90
80
80
70
70
60
60
Flow [l/min]
Flow [l/min]
The spool position is set to an initial position and the electrical current to the
PVC25 is constant during the test. The motor port pressure is changed from
250 bar to 50 bar and back to 250 bar. The pressure difference between pump
pressure 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 compensators and pre compensators with anti saturation. During the ∆p/q test, the
pressure drop across the valve will change but since the valve is pressure compensated the flow will remain constant. According to the left plot in figure 4.3 and
4.4, both compensators acts as they are supposed to. The flow remains reasonably
constant when ∆p is increased.
The same test can be simulated and the results agrees reasonably well with the
measurements, see the right plot in figure 4.3 and 4.4.
50
40
50
40
30
30
20
20
10
10
0
0
50
100
150
Pump pressure margin [Bar]
200
0
0
50
100
150
Pump pressure margin [Bar]
200
Figure 4.3: ∆p/q test using common pre compensators
The same test was also done with the load connected to motor port B instead
of 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 valve
can be used without them. In order to prevent the flow from going in the wrong
direction a check valve is used instead of a compensator. The pressure difference
between the pump and the motor port will then occur across the main spool. Flow
forces will act only on the meter in orifice since the oil is directed directly to the
reservoir after the load.
The position of the spool is measured during changes in the pressure drop
across the main spool. The influence of flow forces can be seen in the left plot in
Validation of Simulation Models
120
120
100
100
80
80
Flow [l/min]
Flow [l/min]
46
60
60
40
40
20
20
0
0
50
100
150
Pump pressure margin [Bar]
0
0
200
50
100
150
Pump pressure margin [Bar]
200
Figure 4.4: ∆p/q test using pre compensators with anti saturation
6
6
5.5
5.5
5
5
Spool position [mm]
Spool position [mm]
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.
4.5
4
3.5
3
2.5
2
0
4.5
4
3.5
3
2.5
50
100
150
Pump pressure margin [Bar]
200
2
0
50
100
150
Pump pressure margin [Bar]
200
Figure 4.5: The influence of flow forces
On the meter out orifice the influence of flow forces is reduced with special
geometry of the spool. Therefore flow forces on the meter out orifice are set to
zero [8].
4.2 Test Rig
47
6
Spool position [mm]
5.5
5
4.5
4
3.5
3
2.5
2
0
20
40
60
80
100
120
Pump pressure margin [Bar]
140
160
180
200
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 to
the PVC25. The current is then switched off resulting in no pilot pressure and the
main spool in neutral position. When the current is switched on again a step will
be 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 similar
behaviour can be seen in the simulations. Since the orifice is enough to get the
expected dynamics the cartridge valve and the pressure reducer can be neglected,
see section 3.2.1.
6
6
Main spool position in simulation
Main spool position in laboratory
5
5
4
4
Position [mm]
Position [mm]
Main spool position in laboratory
3
3
2
2
1
1
0
0
0.5
1
Time [s]
1.5
2
0
0
0.5
1
Time [s]
1.5
2
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 be
both pressure and displacement controlled allowing a load sensing system and a
48
Validation of Simulation Models
displacement controlled system to be tested on the same machine. The actuators
are cylinders controlling the lift and tilt functions. The valve is equipped with
both common pre compensators and pre compensators with anti saturation for
each function.
Parker´s iqan system controls the hydraulic and consists of an ecu and several
i/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
Pump pressure
Pump pressure
Load pressure
Motor port pressure
Motor port pressure
Motor port pressure
Motor port pressure
Position
Position
Pump rotational speed
Pump cam position
Command signals
Location
Pump
Valve
Valve
Lift A-side
Lift B-side
Tilt A-side
Tilt B-side
Lift cylinder
Tilt cylinder
Engine
IQAN
IQAN
Unit
[bar]
[bar]
[bar]
[bar]
[bar]
[bar]
[bar]
[cm]
[cm]
[rev/min]
[%]
[%]
4.2 Test Rig
4.2.1
49
Pump
The test rig is used to validate the pump. A step is done in the lever and the pump
pressure is measured. Since the measurement is done on a mobile application the
pressure will oscillate with a low frequency because the machine swings when
a step is done. The pump will also cause pressure ripples. To keep the pump
model simple those phenomenon are not considered in the simulation. The most
important thing is the on stroke time, see figure 4.9.
The gain of the pump controller is adjusted for pressure mode and displacement
mode. Since the control error is smaller in displacement mode, a higher gain is
needed compared to load sensing mode. The gain together with the orifice area
for piston 2 in the pump control valve, see figure 3.2, is adjusted to give the right
on stroke time.
120
120
Pump pressure in simulation
Pump pressure in test rig
100
100
80
80
Pressure [Bar]
Pressure [Bar]
Pump pressure in test rig
60
40
20
0
0.8
60
40
20
0.9
1
1.1
Time [s]
1.2
1.3
1.4
0
0.8
0.9
1
1.1
Time [s]
1.2
1.3
1.4
Figure 4.9: P1 pump validation
4.2.2
Load Pressure Feedback
When modelling the load pressure feedback pipe, friction, compressibility and wave
dynamics could be taken into consideration. The friction is however neglected
because there is almost no flow in the pipe. Wave dynamics are only of interest
if the time taken by a wave to travel along the pipe is longer than the sampling
rate, see equation (4.1).
twave =
r
ρ
L
=L
≈ 5ms
c
βe
(4.1)
Since the communication interval with iqan is 10 ms there is no need to take
wave 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 on
the bulk modulus and the volume, see equation (4.2).
dp
βe
= Σqin
dt
V
(4.2)
50
Validation of Simulation Models
The pressure build up can be validated by measurements on the test rig. If a
step is done in the command signal pressure will be built up in the load pressure
feedback. By plotting the load pressure as a function of time the pressure build
up can be seen in figure 4.10.
In the simulation model the compressibility of the fluid and expansion of the
pipe wall are taken into account by using an effective bulk modulus. This is
calculated 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 the
model and the measurement can be seen in figure 4.10.
100
100
Load pressure in simulation
Load pressure in test rig
Load pressure in test rig
80
Pressure [Bar]
Pressure [Bar]
80
60
40
20
0
1
60
40
20
1.1
1.2
Time [s]
1.3
1.4
0
1
1.1
1.2
Time [s]
1.3
1.4
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 must
be kept in mind. Pressure losses arise due to friction between the fluid and the
wall of the pipe and the friction in the fluid. Pressure losses due to friction can
be calculated according to equation (4.3). Also one time losses and losses due to
a disturbance source will arise but they are not considered in this master thesis.
∆pf = λ
l ρv 2
d 2
(4.3)
The test rig can be used to validate pressure losses. The difference between
the pressure at the pump and at the valve are pressure losses, ∆pf . By increasing
the 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 changed
to achieve appropriate pressure losses, see figure 4.11. The friction factor λ is
calculated 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 in
the motor ports is measured next to the valve on the test rig. To compensate for
this simplification the force acting on the cylinder might be made a little bigger
in the model.
4.3 Load Sensing Systems
51
150
10
Pressure [Bar]
Flow [l/min]
Flow in test rig
100
50
0
2
3
4
Time [s]
5
3
4
Time [s]
5
6
10
Pressure [Bar]
Flow in simulation
Flow [l/min]
5
0
2
6
150
100
50
0
2
Pressure losses in test rig
3
4
Time [s]
5
6
Pressure losses in simulation
5
0
2
3
4
Time [s]
5
6
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 but
appropriate values has been estimated. Wave dynamics has not been taken into
account 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 they
are supposed to, two different tests are made in the test rig and compared with
the simulation models. The same tests are also made in the test rig using a
displacement controlled system, see chapter 6.
Since the simulation models are used to evaluate the system characteristics
it cannot be compared exactly to the measurements done on the test rig. The
important thing is that the behaviour is correct. For example, in the test rig
the cylinders have different diameters resulting in different flow demand for the
same velocity. In the simulation models the cylinders are identical to simplify the
comparison between the two functions.
4.3.1
Pump Saturation
When the pump is saturated different system behaviour can be expected depending
on what compensator being used. If the valve is equipped with common pre
compensators actuator interference will occur, see section 2.2.1. When using pre
compensators 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 both
functions until the pump cannot supply sufficient flow. By continuing to increase
the command signal the behaviour in a saturated situation is shown. The heaviest
load is represented by the lift function and the lightest load by the tilt function.
The first test is made with common pre compensators.
52
Validation of Simulation Models
120
0.4
Lift actuator position
Tilt actuator position
Lift and tilt command
100
0.3
Position [m]
Lever [%]
80
60
0.2
40
0.1
20
0
0
1
2
3
Time [s]
4
5
0
0
6
1
2
3
Time [s]
4
5
6
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 until
maximal flow from the pump is delivered. The velocity of the tilt will then continue
to increase while the lift will loose speed. Eventually all pump flow will be delivered
to the tilt and the lift will stop completely.
When the same test is made in the simulation model the same result is achieved
according to figure 4.13. The interesting part is however to analyse how the compensators work to attain this behaviour. When the pump cannot supply sufficient
flow ∆pp will decrease. To prevent the reduced pressure to decrease, the compensator 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 across
the main spool the flow and thus the velocity will decrease. If the reduced pressure drops below the load pressure plus the resulting spring pressure check valve
position will be reached and the function will stop according to section 3.2.3.
100
0.4
Tilt actuator position
Lift actuator position
Position [m]
Lever [%]
Lift and tilt command
50
0
0
1
2
3
4
0.3
0.2
0.1
0
0
5
1
2
Time [s]
4
5
15
Tilt compensator area
Lift compensator area
1
Pressure [Bar]
Area/max area
1.5
0.5
0
0
3
Time [s]
1
2
3
Time [s]
4
5
Pressure drop across lift main spool
10
5
0
0
1
2
3
4
5
Time [s]
Figure 4.13: A saturated situation with common pre compensators in simulation
4.3 Load Sensing Systems
53
This is not a problem for the tilt function because the pressure drop across
the compensator is higher. Its restriction area will also increase but the reduced
pressure can be kept constant according to figure 4.13.
The same test is made using pre compensators with anti saturation. According
to figure 4.14 the same velocity is obtained, also when the pump suffers from
saturation. If the test is made in the simulation model the expected result is
obtained according to figure 4.14.
0.4
Tilt actuator position in test rig
Lift actuator position in test rig
0.4
Position [m]
Position [m]
0.6
0.2
0
0
1
2
3
4
Tilt actuator position in simulation
Lift actuator position in simulation
0.2
0
0
5
1
2
Time [s]
4
5
40
Tilt compensator restriction area
Lift compensator restriction area
1
Pressure [Bar]
Area/max area
1.5
0.5
0
0
3
TIme [s]
1
2
3
Time [s]
4
5
Pump pressure margin
20
0
0
1
2
3
4
5
Time [s]
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 its
restriction area almost constant when the pump is saturated. The other compensator however will increase its restriction area in order to maintain the same
pressure drop across both main spools. Hence, the same velocity for both functions
is obtained.
4.3.2
Step Response
A step is made to validate the simulation model but also to compare the performance between a load sensing system and a displacement controlled system. This
comparison is made in section 6.3.
The test is made with two different functions. The lift function represents the
heaviest load and the tilt function the lightest. A step is made in the command
signal for the tilt and two seconds later a step is made with the lift. Finally the
command 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 of
the pump pressure being increased. When making the same test in the simulation
model the compensator to the tilt can be studied to find out why the velocity
remains constant. As seen in figure 4.16 the behaviour of the pump pressure and
the 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 position
by changing its restriction area. A constant pressure drop across the main spool
54
Validation of Simulation Models
100
150
Lift command
Tilt command
Pump pressure
Pressure [Bar]
Lever [%]
80
60
40
20
100
50
0
0
1
2
3
TIme [s]
4
5
0
0
6
1
2
3
Time [s]
4
5
6
Position [m]
0.8
Lift actuator position
Tilt actuator position
0.6
0.4
0.2
0
0
1
2
3
Time [s]
4
5
6
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 the
pump pressure increases because of the higher load. The compensator will then
find a new equilibrium position by decrease its restriction area. It results in a
higher 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 will
decrease and the compensator will return to its previous position, see figure 4.16.
50
0
0
1
2
3
4
Pump pressure
100
50
0
0
5
Tilt actuator position
Lift actuator position
0.3
0.2
0.1
0
0
150
Time [s]
0.4
Position [m]
Pressure [Bar]
Tilt command
Lift command
Position [mm]
Lever [%]
100
1
2
3
Time [s]
4
5
1
2
3
4
5
Time [s]
6
Tilt compensator position
4
2
0
0
1
2
3
Time [s]
Figure 4.16: Step response made in the simulation model
4
5
Chapter 5
Design of Displacement
Controlled Systems
The main difference when controlling the displacement of the pump instead of the
pressure is that the load pressure feedback is removed. Instead the pump receives
a requested displacement from the operator. This principle might imply higher
energy efficiency, faster response and less oscillations.
Figure 5.1: Displacement controlled system [5]
In this chapter, two different types of displacement controlled systems will be
studied. 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 compensators
will be discussed.
5.1
Flow Mapping
In a displacement controlled system the flow delivered from the pump need to be
matched against the flow received by the valve. In this section it will be explained
what factors that will affect the flow calculation across the valve and from the
pump.
5.1.1
Valve Control
When the operator tilts the lever a current is sent to the cartridge valve. The
cartridge valve will move the main spool with a pilot pressure proportional to the
received current. Since the main spool acts against a spring the pilot pressure is
proportional 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 area
and the pressure drop according to equation (5.1).
r
p
2
∆ps ∝ As ∆ps
qs = Cq As
(5.1)
ρ
The pressure drop across the main spool will be equal to the resulting spring
pressure of the compensator. The spring used in this master thesis will give a
pressure drop between 4.5 and 6.5 bar within the compensators control position.
Control position of a common pre compensator means that the compensator is
able to maintain a pressure drop between the above mentioned values across the
main spool independent of variations in pump- and load pressure. If the pressure
drop goes below 4.5 bar check valve mode is reached and the restriction area will
decrease, see figure 3.11. There are other springs available resulting in different
pressure drops across the main spool but they are not considered in this master
thesis.
A certain lever position will result in a certain restriction area for the main
spool. But since the compensator can hold a pressure drop between 4.5 and 6.5 bar
different flows are possible for the same lever position. In figure 5.2 the flow is
plotted against the lever position. The flow for two different pressure drops is
plotted, one for 4.5 bar and the other for 6.5 bar. This creates a flow field which
the compensator can hold within its control position.
5.1 Flow Mapping
57
180
160
4.5 bar pressure drop
6.5 bar pressure drop
140
Flow [l/min]
120
100
80
60
40
20
0
0
10
20
30
40
50
60
Command signal [%]
70
80
90
100
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 lever
position needs to be transformed to a flow. This could be done by measuring the
command signals and use known relationships between command signals, current,
pilot pressure, spool position and spool area. A pressure drop within the control position of the compensator could then be assumed and the flow calculated
according to equation (5.1).
But there is uncertainty in each step between command signal and flow. The
tolerance of the components and factors like flow forces might affect the calculations. Adding up these uncertainties will make the calculations of the flow too
unsure to use. Another method is required.
A load sensing system with common pre compensators can be used instead to
calculate a flow map. A ramp can be made in the command signal and the flow
measured. The flow can then be plotted as function of the command signal. Since
the load pressure feedback is used to control the system pressure and the pump
can supply sufficient flow, the compensator will be in its control position and hence
the flow will be inside the flow field.
To calculate the required displacement of the pump the flow map is used to
transform a lever position into a flow. The flow demand from each section is
summed up and sent to the pump controller. Here the maximal possible flow is
calculated from the rotational speed of the pump and a displacement is ordered
according to the requested flow. At the same time the lever position is sent to the
valve resulting in a proportional manoeuvring of the main spool.
When measuring the shaft speed of the pump there might be some inaccuracy
in the test rig because of the analogue communication between the ecu and the
pump controller. Since the volumetric efficiency of the pump depends on the
current system pressure it might also be a source of problem. Because of this the
pump will not deliver the exact amount of flow demanded by the operator.
58
5.2
Design of Displacement Controlled Systems
System Characteristics
The characteristics of a displacement controlled system are to some extent different
compared to a load sensing system. In this section differences concerning response
behaviour, 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 command signal and pump respond. At first the joystick generates a pilot pressure
which displaces the main spool. The highest load pressure can then travel through
the load pressure feedback and the pump changes its displacement and generates
flow.
When controlling the displacement of the pump a command signal will generate
a pilot pressure and at the same time a requested displacement will be sent to the
pump. Therefore the pump and the valve should react simultaneously on the
command signal [6] [2].
The pump controller used in the test rig is not optimal for controlling the
displacement. Because of that, there is no focus on the response behaviour in this
master thesis. As shown in section 6.3 the response in a load sensing system is
equal to the response in a displacement controlled system with the current pump
controller.
5.2.2
Dynamic Stability
Pressure controlled pumps operate in a pressure closed loop control mode where
the highest load pressure can change significantly. Factors such as oil temperature, natural frequencies and damping levels might affect this loop. Therefore a
fixed setting of the control parameters must be a compromise across all operating
conditions. Unfortunately some operating conditions might exceed the stability
limit resulting in an increase of the hydraulics tendency to oscillate [6] [2].
When the pump is displacement controlled it operates in an open control since
the requested displacement is set by the operator. Problems related to the load
pressure feedback are therefore terminated and less oscillations can be expected.
However, this is considered outside of this master thesis scope but might be of
interest in future investigations.
5.2.3
System Pressure
In a load sensing system the pump controller will adjust the pump pressure in
order to maintain the pump pressure margin, ∆pp . To control the pressure the
pump changes its displacement resulting in a flow. The flow will hence be changed
automatically in order to maintain ∆pp and the pump pressure will be the highest
load pressure plus ∆pp .
When the pump is displacement controlled it will deliver the flow demanded
from the operator. On its way to the actuators the oil will pass several restrictions
creating 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 the
highest load pressure plus the pressure losses in order to deliver the demanded
flow to the actuators. The similar behaviour can be seen in a constant flow system
when no flow is throttled through the open centre valve.
The pump pressure margin used in load sensing systems is set to a fixed level in
order to transport oil to the actuators across all flow resistances and under the most
unfavourable conditions, such as cold oil or maximal flow rate. However during
other conditions ∆pp is too high and pressure is throttled across the compensators
resulting in a waste of energy.
Because the system pressure is automatically adjusted when controlling the
displacement of the pump, an optimal system pressure is always achieved. Both
flow and pressure is thus adapted to what is needed by the actuators in every possible 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, see
section 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 system
the position of the compensator is unknown. The exact pressure drop across the
main spool and thus the position in the flow field is therefore also unknown. This
means 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 is
outside the flow field. The simulation model of a displacement controlled system
equipped with common pre compensators, see figure 3.20, can be used to simulate
if not enough or too much flow is delivered.
60
Design of Displacement Controlled Systems
5.3.1
Not Enough Flow is Delivered
Two different loads are used in the simulation, the heaviest represent by the lift
function and the lightest by the tilt function. The same constant lever position is
used for both functions. The demanded flow is then decreased from a value inside
the flow field.
1.5
Tilt actuator position
Lift actuator position
0.4
Area/max area
Position [m]
0.5
0.3
0.2
0.1
0
71
70
69
68
67
66
Flow [l/min]
65
64
0.5
1
2
3
4
5
Time [s]
30
Pressure drop across lift main spool
Pressure [Bar]
Pressure [Bar]
1
0
0
63
10
5
0
0
Lift compensator area
1
2
3
Time [s]
4
5
Pump pressure margin
20
10
0
0
1
2
3
4
5
Time [s]
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 the
pump delivers flow according to the flow field. But as the demanded flow decreases
∆pp will decrease. The compensator with the heaviest load will therefore increase
its 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 control
area. The function with the heaviest load will therefore decrease in speed since
the pressure drop across the main spool is decreased. The same behaviour can be
seen in a load sensing system with common pre compensators when the pump is
saturated, see section 4.3.1. The consequence if the pump delivers insufficient flow
is hence actuator interference.
5.3 Incorrect Flow Delivery
5.3.2
61
Too Much Flow is Delivered
It is also interesting to see what happens if the delivered flow is outside the flow
field and too much flow is delivered from the pump. To test this scenario the same
simulation is made as in section 5.3.1, but the demanded flow is now increased
from a value inside the flow field. Only one actuator is needed in the simulation
to show what happens.
0.4
300
Pressure [Bar]
Position [m]
Lift actuator position
0.3
0.2
0.1
0
72
73
74
75
76
Flow [l/min]
77
78
Pump pressure
200
100
0
0
79
3
4
5
15
Lift compensator area
Pressure [Bar]
Area/max area
2
Time [s]
0.8
0.6
0.4
0.2
0
0
1
1
2
3
4
5
Pressure drop across lift main spool
10
5
0
0
Time [s]
1
2
3
4
5
Time [s]
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 independent of the flow increase. Since more flow is delivered from the pump than the
actuator receives pressure will be built in the pipe connecting the pump and the
valve according to equation (5.2).
dp
βe
= Σ(qin − qout )
dt
V
(5.2)
When the pressure is built up the compensator will reduce its restriction area
in order to maintain a constant pressure drop across the main spool. The pressure
in the pipe connecting the pump and the valve will increase until it exceeds the
cracking pressure for the ∆pp limiter. The flow will then be throttled to the
reservoir in order to maintain a maximum system pressure. In figure 5.5 the
cracking pressure for the ∆pp limiter is set very high in order to see when the
pressure increases. If a lower cracking pressure is set, the ∆pp limiter will throttle
the flow to the reservoir resulting in energy losses.
Demanding to much flow will result in energy losses. Since the pump pressure
increases and the load pressure remains constant unnecessary pressure losses will
occur across the compensator. Also the ∆pp limiter will contribute to the energy
losses. The extra flow delivered from the pump, which not passes through the
valve, will be throttled to the reservoir resulting in high energy losses.
62
Design of Displacement Controlled Systems
5.4
Displacement Controlled System with Flow
Sharing Capabilities
As explained in section 5.3 problems occur when controlling the displacement of
the pump and using a common pre compensating valve. Too many factors are
unknown and the consequence if an incorrect flow is delivered from the pump is
either energy losses or actuator interference. This is not acceptable and a better
solution is a necessity.
The flow across the main spool is dependent of the restriction area and the
pressure drop according to equation (5.1). If the area is kept constant by the
operator and the flow varies because of for example inaccuracy in the pump control
the pressure drop also needs to vary. With common pre compensators the pressure
drop 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 a
bigger flow field.
A solution to the problem is to utilize pre compensators with anti saturation
instead of common pre compensators. The flow across the main spool is then
dependent on the pump pressure margin, ∆pp , instead of the resulting spring
pressure, see equation (2.7). The compensator will make sure that the pressure
drop across the main spool becomes exactly what is needed for the flow to pass
by. Hence the pressure drop across the main spool can theoretically be whatever
and the flow field will be infinitely large, see figure 5.6.
180
160
140
4.5 bar pressure drop
6.5 bar pressure drop
Small pressure drop
Large pressure drop
Flow [l/min]
120
100
80
60
40
20
0
0
10
20
30
40
50
60
Command signal [%]
70
80
90
100
Figure 5.6: Flow field when using pre compensators with anti saturation
Another solution is to utilize post compensators. That solution is however not
considered in this master thesis, mainly because a post compensated valve is not
available on the test rig.
The reason why this is working is because the pump pressure is automatically
adjusted by the system. If a higher ∆pp is needed to deliver all flow to the actuators
the pump pressure will increase and vice versa. Observe that this implies only if
the pump is displacement controlled. If the pump is pressure controlled the flow
is automatically adjusted in order to maintain the system pressure. Observe also
that 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 reaches
the 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 with
the pump saturation problem. When sufficient flow cannot be delivered the compensator will share the flow in proportion to demand. Used in a displacement
controlled system, the compensator will share the flow in every situation by adjusting the pressure drop across the main spool. Therefore the compensator will
from now on be called flow sharing compensator.
In a load sensing system the flow is controlled by the restriction area of the
main spool. When the pump is displacement controlled and the valve is equipped
with a flow sharing compensator this is not the case. The flow is then controlled
by the pump and the valve will cause unnecessary pressure losses. It is therefore
smart to increase the restriction area and reduce the pressure drop but still get the
same flow, see equation (5.1). The main spool can hence always be manoeuvred
to its end position and the pressure drop is adjusted to match the flow delivered
from the pump, see figure 5.7.
180
160
140
Flow [l/min]
120
100
80
60
40
20
0
0
10
20
30
40
50
60
Command signal [%]
70
80
90
100
Figure 5.7: Flow field when manoeuvring the main spool to its end position
In the following sections the simulation model of a displacement controlled
system equipped with flow sharing compensators, see figure 3.21, will be utilized
to validate the control principle. From now on the system will be referred to a
displacement controlled system with flow sharing capabilities.
5.4.1
One Actuator
In the simulation model the flow delivered from the pump is increased according
to figure 5.8. The main spool is manoeuvred to its end position during the whole
simulation and the load is kept at a constant level. The compensator will make
sure that all flow is delivered to the actuator by increasing the pressure drop across
the main spool. The pump pressure is therefore also increased when more flow
is delivered from the pump, partly because the bigger pressure drop and partly
because increased friction losses in the pipe connecting the pump and the valve.
64
Design of Displacement Controlled Systems
200
140
Pump pressure
Pressure [Bar]
Flow [l/min]
Delivered flow
150
100
50
0
1
2
3
Time [s]
4
120
100
80
1
5
40
2
5
Pressure drop across main spool
Pressure [Bar]
Pressure [Bar]
4
10
Pump pressure margin
30
20
10
0
1
3
Time [s]
2
3
Time [s]
4
5
8
6
4
2
0
1
2
3
Time [s]
4
5
Figure 5.8: ∆p controls the flow
5.4.2
Two Actuators
If two actuators are used simultaneously the compensators will make sure that the
pressure drop across the main spools is the same. To get a correct flow distribution
to 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 position
as before and the other section should be manoeuvred in proportion to the flow
demand. Because the pressure drop is the same across both main spools, the
restriction 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 command
signal was sent to the valves. Here the command signals are recalculated in order
to manoeuvre the section with the highest flow demand completely and the other
in proportion to the flow demand. The original command signal is sent to the
pump in order to deliver the flow demanded from the operator. A nice feature
is that operator can decide the characteristics on the lift and tilt map. This is
because the compensator will make sure that the pressure drop across the main
spool is exactly what is needed for the flow to pass by.
When testing this in the simulation model one actuator is activated and all
flow delivered from the pump reaches that actuator. Then another actuator with
a bigger flow need is activated. As seen in figure 5.9 the right amount of flow is
delivered to both actuators because the main spools are manoeuvred in proportion
to the flow demand.
5.4 Displacement Controlled System with Flow Sharing Capabilities 65
100
1.5
Demanded flow tilt
Demanded flow lift
Delivered flow tilt
Delivered flow lift
Position / max position
Flow [l/min]
75
Tilt main spool position
Lift main spool position
50
25
0
0
1
2
3
4
1
0.5
0
0
5
1
2
Time [s]
3
4
5
Time [s]
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 a
real application this is not the case. When including flow forces in the simulation
the flow delivered to the actuators might not be exactly as demanded. What
happens is that one actuator gets more flow and the other less flow according to
figure 5.10. This is however not a big problem because it will hardly be noticed in
a real application.
100
1.5
Demanded flow tilt
Demanded flow lift
Delivered flow tilt
Delivered flow lift
Position/ max position
Flow [l/min]
75
Tilt main spool position
Lift main spool position
50
25
0
0
1
2
3
Time [s]
4
5
1
0.5
0
0
1
2
3
4
5
Time [s]
Figure 5.10: Two actuators affected by flow forces
This problem is not unique for displacement controlled systems. Also in load
sensing systems, flow forces will have influence on the main spool position and
thus the flow. The current to the cartridge valve and the pilot pressure together
66
Design of Displacement Controlled Systems
with the tolerances of the components will also affect the main spool restriction
area.
5.4.4
Position Feedback
A way of solving this is to include a position feedback in the control loop. If the
actual position of the main spool is known the current sent to the valve can be
adjusted in order to achieve the reference position. If the same test as previous is
made with a position feedback a correct flow distribution is achieved according to
figure 5.11. Since a position sensor on the main spool is not available on the test
rig this solution is left outside the scope of this master thesis.
125
1.5
Demanded flow tilt
Demanded flow lift
Delivered flow tilt
Delivered flow lift
Position/ max position
Flow [l/min]
100
Tilt main spool position
Lift main spool position
75
50
1
0.5
25
0
0
1
2
3
Time [s]
4
5
0
0
1
2
3
4
5
Time [s]
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 equal
pressure drop across the main spools is obtained. But according to figure 5.12 this
is not the case. It is because the compensator at the heaviest load reaches its end
position and it is therefore unable to maintain the same pressure drop as the other
compensator.
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 be
throttled across the compensator at the heaviest load. The flow across the compensator can hence be calculated according to equation (5.3). Those equations
are only valid if the compensator is in its control position. When a flow sharing
compensator is utilized control position means that the compensator not reaches
its end position.
Lever [%]
100
Tilt command
Lift command
50
0
0
1
2
3
4
5
Position/ max position
5.4 Displacement Controlled System with Flow Sharing Capabilities 67
2
Lift main spool position
Tilt main spool position
1.5
1
0.5
0
0
1
2
Time [s]
4
5
0.5
Pressure drop across tilt main spool
Pressure drop across lift main spool
Position [m]
Pressure [Bar]
10
5
0
0
3
Time [s]
1
2
3
4
5
Tilt actuator position
Lift actuator position
0.4
0.3
0.2
0.1
0
0
1
Time [s]
2
3
4
5
Time [s]
Figure 5.12: Two actuators with compensator in its end position
s Ac1
2
1−
∆pp
qc = Cq Ac
ρ
Ac2
(5.3)
Since the flow across the main spool and the compensator is the same, the
equations can be put together.
s s Ac1
2 Ac1
2
1−
∆pp = Cq Ac
∆pp
(5.4)
Cq As
ρ Ac2
ρ
Ac2
When equation (5.4) is simplified a relationship between the restriction areas
and the areas exposed to the control pressures is obtained according to equation (5.5).
r
Ac2
−1
(5.5)
As = Ac
Ac1
If the restriction area of the compensator is maximized and the compensator
is designed like it is today the maximal allowed restriction area of the main spool
can be calculated.
r
Ac2
As,allowed = Ac,max
− 1 < As,max
(5.6)
Ac1
According to equation (5.6) the compensator cannot be in its control position
while the main spool is fully open. One opportunity is to limit the manoeuvring
of the main spool so that a bigger restriction area is not achieved. However, a
decrease of the restriction area means an increase of the pressure drop and thus a
decrease of the energy efficiency. When simulating this solution the compensator
will not reach its end position and the same pressure drop across the main spools
is maintained according to figure 5.13.
Design of Displacement Controlled Systems
100
Lever [%]
Tilt command
Lift command
50
0
0
1
2
3
4
5
Position/ max position
68
2
Tilt main spool position
Lift main spool position
1.5
1
0.5
0
0
1
2
Time [s]
4
5
0.5
Pressure drop across tilt main spool
Pressure drop across lift main spool
20
Position [m]
Pressure [Bar]
30
10
0
0
3
Time [s]
1
2
3
4
Tilt actuator position
Lift actuator position
0.4
0.3
0.2
0.1
0
0
5
1
Time [s]
2
3
4
5
Time [s]
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 areas
exposed to the control pressures can be changed. The restriction area of the main
spool 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 the
maximal restriction area of the main spool. If the quota is decreased, the main
spool can be manoeuvred to its end position. The interesting part is however how
this will affect the pressure drop across the valve. The flow is kept at a constant
level and ∆p is plotted against the main spool restriction area.
150
100
pressure drop across main spool
allowed restriction area for main spool
maximal restriction area for main spool
0.8
Pressure [−]
Main spool restriction area [%]
1
equation (5.5)
allowed restriction area for main spool
maximal restriction area for main spool
0.6
0.4
50
0.2
0
0.2
0.3
0.4
0.5
0.6
Ac1/Ac2 [−]
0.7
0.8
0
25
50
75
100
Main spool restriction area [%]
Figure 5.14: Redesign of the quota
The right plot in figure 5.14 shows that a decrease of the quota and thus an
increase of the restriction area implies a lower ∆p for the same flow.
5.4 Displacement Controlled System with Flow Sharing Capabilities 69
A second step in the redesign could be to increase the maximal restriction
area of the compensator. A new optimal point can be found in the left plot in
figure 5.15 and the pressure drop would decrease even more according to the right
plot in figure 5.15.
150
pressure drop with present area
maximal restriction area for main spool
pressure drop with increased area
0.8
Pressure [−]
Main spool restriction area [%]
1
equation (5.5)
maximal restriction area for main spool
equation (5.5) with increased area
100
0.6
0.4
50
0.2
0
0.2
0.3
0.4
0.5
0.6
Ac1/Ac2 [−]
0.7
0
0.8
25
50
75
100
Main spool restriction area [%]
Figure 5.15: Redesign of the quota and the maximal restriction area
Lever [%]
100
Tilt command
Lift command
50
0
0
1
2
3
4
5
Position/ max position
A redesign of the compensator would give additional energy savings but since
the compensator available on the test rig is of original design this is not further
investigated in this master thesis. When comparing figure 5.16 and 5.13, it shows
the same performance but with a decreased pressure drop across the main spool.
2
Tilt main spool position
Lift main spool position
1.5
1
0.5
0
0
1
2
Time [s]
4
5
0.5
Pressure drop across tilt main spool
Pressure drop across lift main spool
Position [m]
Pressure [Bar]
10
5
0
0
3
Time [s]
1
2
3
4
Time [s]
5
Tilt actuator position
Lift actuator position
0.4
0.3
0.2
0.1
0
0
1
2
3
4
5
Time [s]
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 the
meter out restriction. This is due to the gravitational force pushing the actuator.
70
Design of Displacement Controlled Systems
Because of that, the main spool is controlled in proportional to the command
signal when making a lowering motion. To avoid cavitations the pump will deliver
enough flow to the piston rod side.
By controlling the main spool in proportional to the command signal no additional energy savings across the valve is obtained. But the energy savings concerning ∆pp is still there because the system will adjust the pump pressure to a
minimum level.
A way of getting additional energy savings while making a lowering motion
is to utilize an unloaded lowering. Oil to the piston rod side will then partly be
delivered from the reservoir instead of the pump. This solution is however not
considered 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 other
reason is unable to continue its movement. In a load sensing system the flow is
changed automatically to maintain the system pressure. If a cylinder is unable
to continue its movement, no flow is needed to that section in order to maintain
the system pressure. Hence, the other actuators will not be affected and the right
amount of flow will be delivered from the pump.
In a displacement controlled system the pump will continue to deliver the flow
demanded from the operator, even though a cylinder is unable to continue its
movement. If the flow has nowhere else to go, it will go to another actuator
resulting in an increase of the speed for that actuator. This is not desirable and a
solution is required.
One solution is to utilize port relief valves in the motor ports. When the
system pressure reaches the cracking pressure for the port relief valve the flow will
be 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 to
the reservoir from a high pressure level. It is also not desirable to set the cracking
pressure for the port relief valves at the same level as the maximal allowed system
pressure.
A better solution is to utilize a simple position sensor on the cylinder. When
the cylinder reaches its end position the flow demanded from the operator to that
section can be cancelled. It means that the pump will decrease its displacement
and only deliver the flow demanded to the other actuators. The flow demand is
cancelled until the operator demands a movement in the opposite direction. This
solution will save a lot of energy but is not further investigated in this master
thesis.
Chapter 6
System Improvements Verifying Measurements
To verify the simulation model of a displacement controlled system with flow
sharing capabilities the system has been implemented in the test rig. Since the
same components as in a load sensing system is utilized no hardware changes is
necessary.
To get a proper comparison between a load sensing system and a displacement
controlled system with flow sharing capabilities measurements has been made on
the test rig. During the tests the main spool has been manoeuvred to its end position despite that it will affect the system operability. This is because the system
potential concerning energy efficiency should be highlighted. If a displacement
controlled system with flow sharing capabilities is implemented in a commercial
application the compensator needs to be redesigned, see section 5.4.6. If the compensator is redesigned the main spool could be manoeuvred to its end position
with 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 the
pump and the valve depends on the flow. In a load sensing system ∆pp is set to a
constant value resulting in unnecessary pressure losses across the compensator. In
a displacement controlled system with flow sharing capabilities ∆pp is continuously
adjusted to a minimum level, see section 5.4. In figure 6.1, ∆pp as function of
the flow is shown for a load sensing system and in figure 6.2 for a displacement
controlled system with flow sharing capabilities.
71
72
System Improvements - Verifying Measurements
100
60
Position [cm]
Lever [%]
Lift command
50
0
0
1
2
3
4
Time [s]
5
6
20
0
0
7
1
2
3
4
Time [s]
5
6
7
40
Pressure [Bar]
150
Flow [l/min]
Lift actuator position
40
100
50
0
0
1
2
3
4
Time [s]
5
6
Pump pressure margin
30
20
10
0
0
7
1
2
3
4
Time [s]
5
6
7
Figure 6.1: Pump pressure margin in a load sensing system
100
60
Position [cm]
Lever [%]
Lift command
50
0
0
1
2
3
4
Time [s]
5
6
20
0
0
7
1
2
3
4
Time [s]
5
6
7
40
Pressure [Bar]
150
Flow [l/min]
Lift actuator position
40
100
50
0
0
1
2
3
4
Time [s]
5
6
7
Pump pressure margin
30
20
10
0
0
1
2
3
4
Time [s]
5
6
Figure 6.2: Pump pressure margin in a displacement controlled system
7
6.2 Pump Saturation
6.2
73
Pump Saturation
To show how a displacement controlled system with flow sharing capabilities behaves 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 should
be shared proportionally between all active functions. As seen in figure 6.3 and
6.4 the system acts in the same way as a load sensing system with anti saturation
but with a lower ∆pp when the pump is not saturated.
100
60
Position [cm]
Lever [%]
Lift and tilt command
50
0
0
1
2
3
4
Time [s]
5
6
1
2
3
4
Time [s]
5
6
7
40
Pressure [Bar]
Pressure [Bar]
Tilt actuator position
Lift actuator position
20
0
0
7
100
Pump pressure
50
0
0
40
1
2
3
4
Time [s]
5
6
Pump pressure margin
20
0
0
7
1
2
3
4
Time [s]
5
6
7
5
6
7
Figure 6.3: Pump saturation in a load sensing system
100
60
Position [cm]
Lever [%]
Lift and tilt command
50
0
0
1
2
3
4
Time [s]
5
6
1
2
3
4
Time [s]
40
Pressure [Bar]
Pressure [Bar]
Tilt actuator position
Lift actuator position
20
0
0
7
100
Pump pressure
50
0
0
40
1
2
3
4
Time [s]
5
6
7
Pump pressure margin
20
0
0
1
2
3
4
Time [s]
5
Figure 6.4: Pump saturation in a displacement controlled system
6
7
74
System Improvements - Verifying Measurements
6.3
Step Response
Even though the response behaviour is outside the scope of this master thesis
it 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 in
section 4.3.2 is made with a displacement controlled system with flow sharing
capabilities. According to figure 6.7 the response in the systems is comparable.
100
60
Lift command
Tilt command
50
Position [cm]
Lever [%]
80
60
40
20
0
0
Lift actuator position
Tilt actuator position
40
30
20
10
1
2
3
Time [s]
4
5
0
0
6
1
2
3
Time [s]
4
5
6
5
6
Figure 6.5: Step response in a load sensing system
100
60
Lift command
Tilt command
50
Position [cm]
Lever [%]
80
60
40
20
0
0
Lift actuator position
Tilt actuator position
40
30
20
10
1
2
3
Time [s]
4
5
6
0
0
1
2
3
Time [s]
4
Figure 6.6: Step response in a displacement controlled system
6.4 Short Duty Cycle
75
100
50
Lift command both system
Tilt command both system
40
Position [cm]
Lever [%]
80
60
40
20
0
0
Lift actuator position, LS
Lift actuator position, New system
Tilt actuator position, LS
Tilt actuator position, New system
30
20
10
1
2
3
Time [s]
4
5
6
0
0
1
2
3
Time [s]
4
5
6
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 controlled system with flow sharing capabilities is a short duty cycle for a wheel
loader application. This test is suitable because it is important to compare the
two systems under realistic circumstances. Another type of standardized test could
also be made but since the system is implemented in a wheel loader application
the short duty cycle is the chosen one.
Only the working hydraulics is considered in the measurements. Neither the
steering nor the transmission has been taken under consideration.
Figure 6.8: Short duty cycle [3]
76
System Improvements - Verifying Measurements
When making a short duty cycle the wheel loader drives into the gravel pile
to fill the bucket. The lift and tilt functions are then increased. The wheel loader
will then drive backwards towards the reversing point and then towards the load
receiver. The wheel loader will empty its bucket on the load receiver by decreasing
the tilt function. Finally the wheel loader drives backwards towards the reversing
point 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 controlled
system with flow sharing capabilities two identical tests need to be made. This is
almost impossible and if the tests are not identical a proper comparison cannot be
done.
A solution to the problem is to make one test with the displacement controlled
system. Because the performance is similar the exact same test could have been
made with a load sensing system. The only thing that differs is the pump pressure
margin, ∆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 controlled
system in order to maintain a constant ∆pp of 25 bar, a similar test but with a
load sensing system is obtained.
A short duty cycle test with a load sensing system has also been made to verify
that ∆pp in fact is 25 bar. As seen in figure 6.12 a constant ∆pp of 25 bar is a
good approximation.
100
Lift command
Tilt command
Lever [%]
50
0
−50
−100
0
5
10
15
20
25
30
35
Time [s]
Figure 6.9: Command signals using a load sensing system
0.5
Lift actuator position
Tilt actuator position
Position [m]
0.4
0.3
0.2
0.1
0
0
5
10
15
20
25
30
35
Time [s]
Figure 6.10: Actuator positions using a load sensing system
6.4 Short Duty Cycle
77
250
Pump pressure
Load pressure
Pressure [Bar]
200
150
100
50
0
0
5
10
15
20
25
30
35
Time [s]
Figure 6.11: Pump- and load pressure using a load sensing system
60
Pump pressure margin
Average pump pressure margin
Pressure [Bar]
50
40
30
20
10
0
0
5
10
15
20
25
30
35
Time [s]
Figure 6.12: Pump pressure margin using a load sensing system
In the following figures a short duty cycle using a displacement controlled
system with flow sharing capabilities is shown. The corresponding test using a
load 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 of
a load sensing system in this application.
78
System Improvements - Verifying Measurements
100
Lift demand
Tilt demand
Lever [%]
50
0
−50
−100
0
5
10
15
20
Time [s]
25
30
35
40
Figure 6.13: Command signals in a short duty cycle
0.7
Lift actuator position
Tilt actuator position
0.6
Position [m]
0.5
0.4
0.3
0.2
0.1
0
0
5
10
15
20
Time [s]
25
30
35
40
Figure 6.14: Actuator positions in a short duty cycle
250
Pump pressure, displacement controlled system
Pump pressure, load sensing system
Load pressure
Pressure [Bar]
200
150
100
50
0
0
5
10
15
20
Time [s]
25
30
35
40
Figure 6.15: Pump- and load pressure in a short duty cycle
50
Pump pressure margin, displacement controlled system
Pump pressure margin, load sensing system
Pressure [Bar]
40
30
20
10
0
0
5
10
15
20
Time [s]
25
30
35
40
Figure 6.16: Pump pressure margin in a short duty cycle
6.4 Short Duty Cycle
79
30
Power, displacment controlled system
Power, load sensing system
Power [kW]
25
20
15
10
5
0
0
5
10
15
20
Time [s]
25
30
35
40
Figure 6.17: Power consumption in a short duty cycle
200
Energy, displacment controlled system
Energy, load sensing system
Energy [kJ]
150
100
50
0
0
5
10
15
20
Time [s]
25
30
35
40
Figure 6.18: Energy consumption in a short duty cycle
80
System Improvements - Verifying Measurements
Chapter 7
Summary & Conclusions
A displacement controlled system is an alternative to a load sensing system when
energy efficiency is a vital issue. Except for the pump, the exact same components
can be utilized and no additional sensors are necessary. The increased energy
efficiency is due to a lower pump pressure margin since the system will compensate
for pressure losses between pump and valve itself. Hence, no unnecessary energy
losses across the compensator will occur.
Originally the idea of a displacement controlled system was to match the manoeuvring of the main spool with the flow delivered from the pump using common
pre compensators. Such system design would require a very good flow map and
an exact knowledge of the current shaft speed and volumetric efficiency of the
pump. The flow map would also be different depending on the application and
the tolerance of the components. This is not practically achievable in all kinds of
applications.
By using flow sharing compensators instead of common pre compensators all
these problems will be eliminated and additional energy savings across the main
spool is enabled. The compensator will make sure that the pressure drop across
the main spool becomes exactly what is needed for the flow to pass by independent
of the main spool restriction area. Hence the flow delivered from the pump does
not need to be matched against the manoeuvring of the main spool. Neither the
flow map nor the shaft speed nor the volumetric efficiency is then a problem.
Additional energy savings across the main spool can be achieved by manoeuvre
the 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 then
obtained.
In a displacement controlled system with flow sharing capabilities there is hence
no unnecessary energy losses. The compensator at the heaviest load will be completely open and also the main spool with the highest flow demand. If further
energy savings should be obtained, additional pumps or transformers need to be
utilized.
In a load sensing system there is no unnecessary pressure drop across the compensator 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 unnecessary
pressure losses across the main spool either. Hence when maximal flow is delivered and the main spool is completely open, the energy efficiency in a load sensing
system will be equal to a displacement controlled system with flow sharing capabilities. Otherwise, the energy efficiency will be higher in a displacement controlled
system.
40
Pump pressure margin, displacement controlled system
Pump pressure margin, load sensing system
35
Pressure [Bar]
30
25
20
15
10
5
0
0
10
20
30
40
50
60
Flow [l/min]
70
80
90
100
110
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 both
systems if maximal flow is delivered from the pump. Otherwise it can be decreased
a lot. In an application that not uses all available pump capacity during its duty
cycle, the energy efficiency can be increased if a displacement controlled system
is utilized instead of a load sensing system. The wheel loader application used in
this 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 efficiency
cannot be increased with a displacement controlled system. Minimal energy losses
across the compensator and the main spool are already achieved with a load sensing
system. But there are other advantages with a displacement controlled system.
Even though the energy efficiency cannot be increased there is still a potential of
better response and less oscillations.
To summarize, the proposed displacement controlled system with flow sharing
capabilities has better or equal energy efficiency compared to a traditional load
sensing system. It has also a potential of better response and less oscillations. To
switch from a load sensing system would require a displacement controlled pump
and an electronically controlled valve. Neither a flow map nor the accuracy of the
pump nor additional sensors needs to be considered.
Chapter 8
Future Work
• The potential of a better response when controlling the displacement of the
pump needs to be studied, see section 5.2.1. For example, an external pilot
pressure 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 drop
across all main spools because it will reach its end position. An investigation
regarding how to redesign the compensator needs to be made in order to get
a correct flow distribution to several actuators, see section 5.4.6.
• In this master thesis only common pre compensators and flow sharing compensators have been studied. However, it would be interesting to study post
compensators as well.
• To deal with flow forces acting on the main spool a position feedback can
be utilized, see section 5.4.3 and 5.4.4. This will also improve the flow
distribution if several actuators are activated.
• During a lowering motion there are opportunities to get additional energy
savings when controlling the displacement of the pump. For example, an
unloaded lowering might be utilized, see section 5.4.7.
• When the cylinder reaches its end position a position sensor can cancel the
flow 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 hydraulics, problems might occur when controlling the displacement. A study
regarding this problem needs to be done.
• Improve the timing between the pump and the valve by sending signals in
the right moment. Thus preventing the valve to rush ahead and open before
the 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 Verlag
Aachen, 2007.
[3] Reno Filla. Operator and machine models for dynamic simulation of construction machinery. LiTH, 2005.
[4] Institutionen för konstruktions & produktionsteknik. Formelsamling i Hydraulik och pneumatik. Linköpings Tekniska Högskola, 1995.
[5] Per-Anders Kumlin. Implementation of a flow controlled hydraulic work system. 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, New
York, 1967. ISBN 0-471-59617-5.
[8] Bo Nilstam. Oral, December 2008.
[9] K-E. Rydberg O. Olsson. Kompendium i hydraulik.
konstruktions- & produktionsteknik, 1993.
85
Institutionen för
86
Bibliography
Appendix A
Hydraulic Schematic of a
L90LS Valve
87
88
Hydraulic Schematic of a L90LS Valve