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Linkping Studies in Science and Technology.Thesis No. 1595
Fluid Power Systems for Mobile
Applications
with a Focus on Energy Eciency and DynamicCharacteristics
Mikael Axin
LIU-TEK-LIC-2013:29Division of Fluid and Mechatronic Systems
Department of Management and EngineeringLinkping University,
SE581 83 Linkping, Sweden
Linkping 2013
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ISBN 978-91-7519-600-8ISSN 0280-7971LIU-TEK-LIC-2013:29
Copyright c 2013 by Mikael AxinDepartment of Management and
EngineeringLinkping UniversitySE-581 83 Linkping, Sweden
Printed in Sweden by LiU-Tryck, Linkping, 2013.
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To Jennie
Mycket f mnniskor lever i dag de
esta gr frberedelser fr att leva i
morgon.
Jonathan Swift (1667-1745)
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Att inse att man r okunnig r ett bra
steg mot kunskap.
Benjamin Disraeli (1804-1881)
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Abstract
This thesis studies an innovative working hydraulic system
design formobile applications. The purpose is to improve the energy
eciency andthe dynamic characteristics compared to load sensing
systems withoutincreasing the complexity or adding additional
components.
The system analysed in this thesis is referred to as ow control.
Thefundamental dierence compared to load sensing systems is that
thepump is controlled based on the operators command signals rather
thanfeedback signals from the loads. This control approach enables
higherenergy eciency since the pressure dierence between pump and
load isgiven by the system resistance rather than a prescribed pump
pressuremargin. High power savings are possible especially at
medium ow rates.
Furthermore, load sensing systems suer from poor dynamic
charac-teristics since the pump is operated in a closed loop
control mode. Thismight result in an oscillatory behaviour. Flow
control systems have nostability issues attached to the load
pressure feedback since there is none.
Pressure compensators are key components in ow control
systems.This thesis addresses the ow matching problem which occurs
whenusing conventional compensators in combination with a ow
controlledpump. Flow sharing pressure compensators solve this
problem since thepump ow will be distributed between all active
functions. A novel con-trol approach where the directional valve is
controlled without aectingthe cylinder velocity with the objective
of optimizing the damping isproposed.
In this research, both theoretical studies and practical
implementa-tions demonstrate the capability of ow control systems.
Experimentsshow a reduced pump pressure margin and energy saving
possibilities ina short loading cycle for a wheel loader
application.
i
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ii
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Acknowledgements
The work presented in this thesis has been carried out at the
Divisionof Fluid and Mechatronic Systems (Flumes) at Linkping
University.There are several people who have made this thesis
possible and towhom I would like to express my gratitude.
First of all I would like to thank my supervisor, Prof. Petter
Krus, forhis support, supervision and valuable inputs to my work. I
am also verygrateful to Prof. Jan-Ove Palmberg, former head of
division. Thankyou for giving me the opportunity to be a part of
this division. I wouldlike to give special thanks to Dr. Bjrn
Eriksson for his great support inmy work. I would also like to
thank all my other colleagues for makingthe university a fun place
to work at.
Thanks go to Parker Hannin AB for their nancial involvement
andtheir help with hardware and other resources.
Most of all, I would like to thank my family and friends for
alwaysbeing there for me. My greatest gratitude goes to you Jennie,
my won-derful love, for sharing life with me.
Linkping, April, 2013
Mikael Axin
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iv
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Papers
The following three appended papers are arranged in
chronological orderof publication and will be referred to by their
Roman numerals. Allpapers are printed in their originally published
state with the exceptionof minor errata and changes in text and
gure layout in order to maintainconsistency throughout the
thesis.
In papers [I], [II] and [III], the rst author is the main
author, respon-sible for the work presented, with additional
support from the co-writers.A short summary of each paper can be
found in chapter 8.
[I] M. Axin, B. Eriksson, and J.-O. Palmberg. Energy Ecient
LoadAdapting System Without Load Sensing - Design and Evalu-ation.
In: The 11th Scandinavian International Conference onFluid Power
(SICFP09). Linkping, Sweden, June 2009.
[II] M. Axin, B. Eriksson, J.-O. Palmberg, and P. Krus.
DynamicAnalysis of Single Pump, Flow Controlled Mobile Systems.
In:The Twelfth Scandinavian International Conference on FluidPower
(SICFP11). Vol. 2. Tampere, Finland, May 2011, pp. 223238.
[III] M. Axin, J.-O. Palmberg, and P. Krus. Optimized Damping
inCylinder Drives Using the Meter-out Orice - Design and
Exper-imental Verication. In: 8th International Fluid Power
Confer-ence (IFK). Vol. 1. Dresden, Germany, Mar. 2012, pp.
579591.
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Papers not included
Paper [IV] is not included in the thesis but constitutes an
importantpart of the background. The two rst authors are the main
authors,responsible for the work presented, with additional support
from theco-writers.
[IV] M. Axin, R. Braun, A. DellAmico, B. Eriksson, P. Nordin,
K.Pettersson, I. Staack, and P. Krus. Next Generation
SimulationSoftware using Transmission Line Elements. In: Fluid
Power andMotion Control (FPMC). Bath, UK, Sept. 2010, pp.
265276.
vi
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Contents
1 Introduction 1
1.1 Motivation and Needs . . . . . . . . . . . . . . . . . . . .
1
1.2 Aims . . . . . . . . . . . . . . . . . . . . . . . . . . . .
. . 2
1.3 Delimitations . . . . . . . . . . . . . . . . . . . . . . .
. . 2
1.4 Contribution . . . . . . . . . . . . . . . . . . . . . . . .
. 2
2 Mobile Working Hydraulic Systems 3
2.1 Valve Controlled Systems . . . . . . . . . . . . . . . . . .
4
2.1.1 Open-centre . . . . . . . . . . . . . . . . . . . . . .
4
2.1.2 Constant Pressure . . . . . . . . . . . . . . . . . .
6
2.1.3 Load Sensing . . . . . . . . . . . . . . . . . . . . .
7
2.1.4 Individual Metering . . . . . . . . . . . . . . . . .
9
2.2 Valveless Systems . . . . . . . . . . . . . . . . . . . . .
. . 13
2.2.1 Secondary Control using Transformers . . . . . . . 13
2.2.2 Pump Controlled Actuators . . . . . . . . . . . . . 14
2.2.3 Electro Hydraulic Actuators . . . . . . . . . . . . .
15
2.3 System Summary . . . . . . . . . . . . . . . . . . . . . . .
17
3 The Flow Control Concept 19
3.1 Pressure Compensators . . . . . . . . . . . . . . . . . . .
21
3.1.1 Traditional Compensators . . . . . . . . . . . . . .
21
3.1.2 Flow Sharing Compensators . . . . . . . . . . . . . 22
3.2 Pump and Valve Control Approaches . . . . . . . . . . . .
24
3.2.1 Flow Control using Traditional Compensators . . . 24
3.2.2 Flow Control using Flow Sharing Compensators . 26
3.3 Energy Eciency . . . . . . . . . . . . . . . . . . . . . . .
28
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4 Dynamic Analysis 314.1 Mathematical Model . . . . . . . . . .
. . . . . . . . . . . 324.2 Pump Stability . . . . . . . . . . . .
. . . . . . . . . . . . 34
4.2.1 Load Sensing Systems . . . . . . . . . . . . . . . .
344.2.2 Flow Control Systems . . . . . . . . . . . . . . . . 35
4.3 Damping . . . . . . . . . . . . . . . . . . . . . . . . . .
. 374.3.1 Active Control of the Inlet Orice . . . . . . . . .
374.3.2 Design and Control of the Outlet Orice . . . . . . 40
5 Experimental Results 435.1 Energy Eciency Improvements . . . .
. . . . . . . . . . 43
5.1.1 Hardware Requirements . . . . . . . . . . . . . . .
435.1.2 A Demonstrator System . . . . . . . . . . . . . . . 44
5.2 Improved Damping . . . . . . . . . . . . . . . . . . . . . .
47
6 Summary and Conclusions 49
7 Outlook 51
8 Review of Papers 53
Appended papers
I Energy Ecient System Without Load Sensing 61
II Dynamic Analysis of Flow Controlled Systems 85
III Optimized Damping Using the Meter-out Orice 109
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Nomenclature
The quantities used in this thesis are listed in the table.
Capital lettersare used for linearized and Laplace transformed
variables.
Quantity Description Unity
Ac Cylinder area m2
Ac1 Compensator area exposed to control pressure m2
Ac2 Compensator area exposed to control pressure m2
As Directional valve opening area m2
Bp Viscous friction coecient Ns/mCq Flow coecient -Fs
Compensator spring stiness NKca Flow-pressure coecient for the
inlet orice m
3/Pa sKcaopt Kca which gives the highest damping m
3/Pa s
Kcb Flow-pressure coecient for the outlet orice m3/Pa s
Kcbopt Kcb which gives the highest damping m3/Pa s
Lp Pump inductance Pa s2/m3
mL
Load mass kgPa Pressure on the piston side of the cylinder
PaPamax Maximum pressure on the piston side PaPb Pressure on the
piston rod side of the cylinder Pap
LLoad pressure Pa
pLmax
Maximum load pressure PaPp Pump pressure Papr Reduced pressure
Paps Supply pressure PaQa Flow into the cylinder m3/sQb Flow out of
the cylinder m3/sq
LLoad ow m3/s
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Qp Pump ow m3/sQpref Pump ow demand m
3/ss Laplace variable 1/sU Mechanical gear ratio -Va Volume at
the piston side of the cylinder m3
Vb Volume at the piston rod side of the cylinder m3
Vp Pump hose volume m3
Xp Piston position me Bulk modulus Pai Parameter for the inlet
orice -o Parameter for the outlet orice -hmax Maximum damping -pp
Pump pressure margin PaPp Pump pressure margin PaPpref Pump
pressure margin demand Pa Cylinder area ratio - Density kg/m3
Go Open loop transfer functionGpFC Pump transfer functionGpLS
Pump transfer functionGva Inlet valve transfer functionGvea Inlet
valve transfer functionGvb Outlet valve transfer functionHs Pump
hose transfer functionZ
LLoad transfer function
x
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1Introduction
Fluid power systems are used in a wide range of applications,
mobile aswell as industrial. In mobile machinery, such as
construction, forestryand agricultural machines, uid power is used
for both propulsion sys-tems and working hydraulics. An example of
working hydraulics is thesystem controlling the bucket motion of an
excavator. This thesis coversthe area of working hydraulics in
mobile machinery. An innovative sys-tem design is presented and
discussed in relation to both existing andnot yet commercially
available mobile hydraulic systems.
1.1 Motivation and Needs
There are several dierent reasons for preferring uid power
systemsto other technologies. Fluid power components have a
superior powerdensity compared to other technologies, for example
electrical compo-nents [1]. Furthermore, uid power systems have the
ability to handleforce impacts, which makes it more robust than for
example mechanicaltransmissions. Fluid power components are
generally available at lowercost compared to other technologies,
especially for high power applica-tions [1]. Another property of
uid power systems is their good heattransfer capability.
Fluid power systems also present some challenges. The most
impor-tant one concerns their energy eciency [2] [3]. Much progress
has beenmade in making the individual components more ecient [4]
[5]. How-ever, each component has its own optimum working
condition, whichoften leads to poor overall system eciency [5].
When improving energy eciency in uid power systems, the
trend
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Fluid Power Systems for Mobile Applications
is to use additional components and more sophisticated control
algo-rithms [2] [6]. Meanwhile, less attention has been paid to the
dynamicproperties. A hydraulic system with poor dynamic properties
has a ten-dency to oscillate, which has a negative impact on both
the productivityof the application and the comfort of the
operator.
1.2 Aims
The purpose of this thesis is to investigate and analyse how the
energyeciency and the dynamic characteristics of the working
hydraulics inmobile machinery can be improved. Dierent valve
concepts are studied.The hypothesis is that there exist valve
controlled systems which im-prove the energy eciency and the
dynamic properties compared to loadsensing systems, without
increasing the complexity or adding additionalcomponents. The
solutions presented in this thesis are demonstratedthrough both
simulation and experiments.
1.3 Delimitations
This thesis concerns the energy eciency and dynamic
characteristicsof mobile uid power systems. Other aspects, such as
manufacturingand marketing are not taken up. Industrial hydraulics
and propulsionsystems are not included in this work. The thesis is
also limited to thehydraulic system; the combustion engine powering
the hydraulic pumpis therefore not included. The eld of digital
hydraulics is not includedin this thesis.
1.4 Contribution
The most important contribution of this thesis is a deeper
understandingof the dynamic characteristics of ow control systems.
Novel ways ofdesigning and controlling the directional valves in
order to optimize thedamping are proposed and demonstrated. Energy
measurements, whereow control and load sensing systems are compared
in a wheel loaderapplication, are performed analytically and veried
by experiments.
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2Mobile Working
Hydraulic Systems
Mobile hydraulic applications distinguish themselves from other
hy-draulic applications, such as industrial hydraulics, because the
pressureand ow demand varies greatly over time and between dierent
func-tions. Unlike other hydraulic applications, several functions
are oftensupplied by one single pump. This means that the total
installed poweron the actuator side is generally considerably
higher than the installedpump power. This is possible because the
actuators almost never requiretheir maximum power at the same
time.
Fluid power systems have successfully been used in mobile
machinesfor several decades. Because of the machines versatility,
dierent hy-draulic systems have been developed for dierent
applications. Impor-tant properties for hydraulic systems are
energy eciency, dynamic char-acteristics and complexity. However,
the order of these properties variesfor dierent applications. The
following sections give an overview ofthe most commonly used
working hydraulic systems of today. It alsopresents some
interesting system designs that have not yet been com-mercialized
but are attracting considerable attention both in industry aswell
as academia. Energy eciency, dynamic characteristics and
systemcomplexity are discussed and compared.
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Fluid Power Systems for Mobile Applications
2.1 Valve Controlled Systems
The most common hydraulic systems in mobile machines are
systemsbased on valve control. Common to these systems is that they
can besupplied by one single pump, which gives a cost eective and
compactsystem solution. Four dierent hydraulic system designs are
here cate-gorized by open-centre, constant pressure, load sensing
and individualmetering.
2.1.1 Open-centre
Today, most hydraulic systems in mobile machines are of the
open-centretype. In such systems, the directional valves are
designed so that theentire pump ow is directed to tank when no
valve is activated. Thisis commonly achieved by providing the
directional valve with a chan-nel in the centre position connecting
the pump port and the tank, seegure 2.1a. By means of this
open-centre channel, the system pressureis kept at a low level
while the system is idle and the valve is closed.These systems are
designed for use with xed displacement pumps andare therefore often
called constant ow systems.
(a) Simplied schematic of anopen-centre system.
ow
useful wasted powerpre
ssure
power
load demand
system operation point
(b) Pressure and ow characteristicsin an open-centre system.
Figure 2.1 Schematic and eciency of an open-centre system.
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Mobile Working Hydraulic Systems
When a valve is shifted from its centre position, the
open-centre chan-nel begins to close and the pump pressure
increases. Simultaneously,the pump port is connected to either of
the load ports, depending onthe direction of spool movement, while
the other load port is connectedto tank. When the pump ow is
restricted so that the pump pressureis higher than the load
pressure, the check valve opens and there willbe a ow to the load.
The rate of this ow is thus not only dependenton spool
displacement, but also on load pressure. This is called
loaddependency.
If several valves are activated simultaneously, the ow to each
ac-tuator will not only be dependent on its own load, but also on
otheractivated loads. This means that the pressure level at one
load canheavily inuence the speed of another actuator, a phenomenon
calledload interaction.
Another disadvantage of open-centre systems is that the ow is
loaddependent. For heavy loads, the major part of the lever stroke
is used torestrict the pump ow in order to obtain a high pump
pressure. Only aminor part of the stroke is then left for
controlling the speed. This mightbe a serious problem if a heavy
load is be positioned with accuracy, asis often the case for
instance with mobile cranes.
The fact that the ow is load dependent is from a dynamic point
ofview actually an advantage. It gives the system a naturally high
damp-ing, which means that the system is less prone to
oscillations. To obtaindamping from a valve, the ow has to increase
when the pressure dropacross the valve increases and vice versa.
Damping is a preferred prop-erty when handling large inertia loads,
for example the swing functionof a mobile crane.
The most important disadvantage of open-centre systems is that
itmay have poor energy eciency. High energy losses accur when
liftingheavy loads slowly; the pump pressure needs to be high but
only a minorpart of the ow is directed to the load, see gure 2.1b.
Most of the owis then directed through the open-centre channel to
the tank with a highpressure drop, resulting in high energy
losses.
To summarize, open-centre systems have the following advantages
anddisadvantages:
Advantages The system is simple and robust. It has high
damping,which makes it suitable for heavy mobile applications.
Disadvantages Poor overall eciency and interaction between
simul-
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Fluid Power Systems for Mobile Applications
taneously operated loads. The actuator velocity does not
corre-spond to a specic lever displacement but is also a function
of theload pressure.
2.1.2 Constant Pressure
A constant pressure system can be realized using a pressure
controlledvariable displacement pump or a xed pump working against
a pressurerelief valve. In this section, the pressure controlled
pump solution willbe discussed because of its higher eciency, see
gure 2.2a. When thesystem is idle, each directional valve has a
closed pump port and thevariable pump is de-stroked to a small
displacement, compensating forits own losses and thus keeping the
pressure constant. The directionalvalves are of closed centre
type.
(a) Simplied schematic of aconstant pressure system.
useful power
wasted power
ow
pre
ssure
load demand
system operation point
(b) Pressure and ow characteristicsin a constant pressure
system.
Figure 2.2 Schematic and eciency of a constant pressure
system.
There is a ow to the actuator when its directional valve is
shiftedfrom neutral position. Simultaneously, the pump controller
increases itsdisplacement in order to maintain a constant system
pressure. The owrate is dependent on both spool displacement and
load pressure. Conse-quently, constant pressure systems suer from
load dependency. How-ever, the controllability of these systems is
better than in open-centre
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Mobile Working Hydraulic Systems
systems as far as interaction between actuators is concerned.
This is be-cause there is no dependency between the load pressure
and the pumppressure. From a dynamic point of view, constant
pressure systems havesimilar characteristics to open-centre systems
due to their load depen-dency. The damping is therefore high.
Regarding energy eciency, constant pressure systems are a
goodchoice if the present loads tend to be constant. The pump
pressureis then matched against the mentioned constant load.
However, if theload situation alters, high losses might occur. This
is especially truewhen raising a light load with a high velocity,
see gure 2.2b. The mainpart of the entire pressure drop then occurs
across the directional valveand only a minor part is used to lift
the load. The major fraction of thetotal power is therefore spent
in heating the oil. Consequently, these sys-tems not only have
large energy losses but also often need extra energyto cool the
oil.
Advantages No interaction between simultaneously operated loads
anda high damping.
Disadvantages Poor eciency for light loads and the actuator
velocitydoes not correspond to a specic lever displacement but is
also afunction of the load pressure.
2.1.3 Load Sensing
Load sensing systems use a variable displacement pump and closed
cen-tre valves, similar to constant pressure systems. However, the
pump con-troller is designed in a dierent way. Instead of
maintaining a constantpressure, the pump pressure is continuously
adapted according to thehighest load, see gure 2.3. Another load
sensing system design wouldbe to use a xed displacement pump and a
pressure relief valve, adapt-ing its cracking pressure according to
the highest load. That solution,however, is not discussed in this
thesis because of its lower eciency.An early review of load sensing
systems was made by Andersson in [7].
When all directional valves are closed, the pump is de-stroked,
main-taining a low system pressure. When a valve is shifted from
neutralposition, the pump controller senses the load and increases
its pressure,thereby allowing a ow to the actuator. Since the pump
pressure con-tinuously adapts to the load, a specic lever
displacement results in acertain ow, independent of the load
pressure.
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Fluid Power Systems for Mobile Applications
Figure 2.3 Simplied schematic of a load sensing system.
Load sensing systems have no load dependency as long as only
oneload is controlled. However, when several loads are operated
simultane-ously, only the heaviest load will be load independent.
All lighter loadswill suer from both load dependency and load
interaction. In applica-tions where controllability is an important
feature, the valves are oftenequipped with pressure compensators.
The pressure drop across eachdirectional valve is then kept at a
constant level and all functions arethereby load independent and
there will be no load interaction. Pressurecompensators are studied
in detail in section 3.1.
One weakness of load sensing systems using pressure
compensatedvalves is the hydraulic damping. The primary design
endeavours toachieve low inuence on the ow from the load pressure.
This decreasesthe damping capability of the valve. The dynamics of
load sensing sys-tems are studied in more detail in chapter 4.
Load sensing systems have high energy eciency since the pump
con-tinuously adapts its pressure just above the highest load. A
pressuredierence, usually around 20-30 bar, between pump and load
is necess-ary to overcome losses in hoses and valves. This pressure
margin is oftenset substantially higher than necessary to ensure it
is high enough at alloperational points. More details regarding the
pressure margin can befound in section 3.3. When several functions
are operated simultane-
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Mobile Working Hydraulic Systems
ously, high losses might occur at lighter loads. An example is
when alight load is operated with a high velocity and a heavier
load is activatedat the same time, see gure 2.4.
pre
ssure
ow
useful powerwasted power
useful power
pump pressure margin
load 1
load 2
system operationpoint
Figure 2.4 Pressure and ow characteristics in a load sensing
system.
To summarize, pressure compensated load sensing systems have
thefollowing advantages and disadvantages:
Advantages Energy eciency is high although pressure and ow
de-mands vary greatly over time and between dierent functions.
Thesystem has excellent controllability since there is no load
interac-tion and no load dependency.
Disadvantages Low damping, meaning that the system can show
anoscillatory behaviour in certain points of operation. High losses
atlighter loads when several functions are operated
simultaneously.A needless pressure loss in most points of operation
due to anexcessive pressure margin.
2.1.4 Individual Metering
A step forward from load sensing systems using conventional
spool valvesis to decouple the inlet and outlet orices in the
valve, see gure 2.5.Numerous congurations for individual metering
systems have been de-veloped, both in academia as well as in
industry [8]. These conceptsprovide a higher degree of freedom as
all four orices are separated
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Fluid Power Systems for Mobile Applications
and can be controlled individually. The main benet of this
increasedfreedom is that the ow paths can be changed during
operation. Fourdierent operational cases can be identied [9], see
gure 2.7.
Figure 2.5 Simplied schematic of an individual metering
system.
Normal operation The load is operated as in a conventional
system;oil is withdrawn from the pump and the return oil is fed to
tank. Inconventional systems, the outlet orice opening area is
determinedby the spool position. For an independent metering
system, theoutlet orice can be separately controlled with the
objective of, forexample, reducing metering losses.
Regenerative operation When operating a light and a heavy load
si-multaneously, it is possible to use the cylinder as a discrete
trans-former. This is done by connecting both load ports to the
pumpline, thereby increasing the pressure level and decreasing the
owlevel. As can be seen in gure 2.6, this might result in
substantiallylower power losses.
Energy neutral operation This mode is benecial when, for
exam-ple, lowering a load while the system pressure level is high.
Insteadof taking high pressure oil from the pump and throttling it
acrossthe control valve, the oil could instead be withdrawn from
tank.
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Mobile Working Hydraulic Systems
No power is then needed from the system pump. The return oil
isfed to tank.
Recuperative operation This mode is similar to energy neutral
op-eration. However, instead of feeding the return oil to tank, it
isdirected into the pump line. The cylinder thus works as a pumpand
can be used to drive other loads or operate the system pumpas a
motor.
pre
ssure
ow
useful powerload 1
load 2
wasted power
useful power
system operation point
Figure 2.6 It is possible to reduce the losses when using the
regenera-tive operation mode in individual metering systems.
Advantages Possibility to optimize the eciency by means of
changingthe ow paths. Recuperation of energy is possible. Active
dampingmeasures are easy to implement because of increased
exibility.
Disadvantages Complex controller and often sensor dependent for
owpath selection.
11
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Fluid Power Systems for Mobile Applications
(a) Normal operation, the load isoperated as in a conventional
sys-tem.
(b) Regenerative operation, bothload ports are connected to
thepump with the objective to even outpressure dierences between
loads.
(c) Energy neutral operation, a loadcan be lowered without any
powerfrom the pump.
(d) Recuperative operation, thecylinder works as a pump
enablingenergy recuperation.
Figure 2.7 Flow paths for the dierent operational cases in
individualmetering systems.
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Mobile Working Hydraulic Systems
2.2 Valveless Systems
One hot research topic in the area of mobile hydraulics is
systems inwhich the control valves are eliminated along with the
metering losses,see gure 2.8. Three interesting concepts are here
categorised by hy-draulic transformers, pump controlled actuators
and electro hydraulicactuators. Such systems are not yet common
commercially in mobileapplications but can be found in, for
example, the aerospace indus-try [10].
pre
ssure
ow
useful power
useful power
load 1
load 2
Figure 2.8 Pressure and ow characteristics in a valveless
system. Allmetering losses are ideally eliminated.
2.2.1 Secondary Control using Transformers
In a secondary control system, all actuators are connected to a
commonpressure rail. This can be realized, for example, by a
pressure controlledpump. An accumulator is often connected to the
common pressure rail,allowing the pump to be downsized according to
the mean ow over aduty cycle. It also allows the possibility of
storing recuperated powergenerated by the load [11]. Controlling
hydraulic motors is the mostcommon use of secondary control [12].
This technology can, however,not be used directly on linear
cylinder drives since the piston area isxed. In that case, a
hydraulic transformer is required, see gure 2.9.
A hydraulic transformer converts an input ow at a certain
pressurelevel to a dierent output ow at the expense of a change in
pressure
13
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Fluid Power Systems for Mobile Applications
level, ideally maintaining the hydraulic power. One way of
realizing atransformer is to combine two hydraulic machines, where
at least one hasa variable displacement. However, the eciency is
limited, mainly be-cause at least one of the machines will operate
under partial loading [13].In recent years, an innovative
transformer concept has been developedby the Dutch company Innas BV
[14]. The conventional transformerwith 2 hydraulic machines has
been replaced by one axial piston unit,thereby avoiding partial
loading conditions. A mean eciency of 93%in a broad region of
operation has been reported [5].
Figure 2.9 Simplied schematic of a constant pressure
systemequipped with transformers and an accumulator, enabling
energy recu-
peration.
Advantages High eciency, even when several loads are operated
si-multaneously. Possibility to recuperate energy from the
loads.
Disadvantages The number of hydraulic machines is increased,
whichmeans more required space and a higher cost. Low damping
sincethe ow is load independent.
2.2.2 Pump Controlled Actuators
Instead of using one pump to supply all actuators, each actuator
has adedicated pump in pump controlled actuator systems. To control
the
14
-
Mobile Working Hydraulic Systems
speed, the pump displacement setting is used as the nal control
element.All losses are thereby ideally eliminated. In reality, the
losses are heavilydependent on the eciency of the system pumps.
Pump controlled ac-tuator systems can principally be dierentiated
into two dierent circuitlayouts, either with the pump arranged in a
closed circuit [15] [16] or inan open circuit [17], see gure
2.10.
Since all actuators have their own dedicated pump, each has to
besized to handle maximum speed. A typical example of a
dimensioningmotion is the lowering bucket motion in a wheel loader.
The loweringow can be several times higher than the maximum pump ow
in asimilar valve controlled system. The dierence is that all ow
has to behandled by the pump in pump controlled actuator systems.
In singlepump systems, the pump can also be downsized since not
every loadis actuated at full speed simultaneously very often. For
these reasons,the total installed displacement tends to be high in
pump controlledactuator systems.
Figure 2.10 Simplied schematic of a pump controlled actuator
sys-tem. It can be realized in both closed- and open circuit.
Advantages High eciency, even when several loads are operated
si-multaneously.
Disadvantages One hydraulic machine for each actuator, which
meansmore required space and a higher cost. Low damping since the
owis load independent.
2.2.3 Electro Hydraulic Actuators
The main component in electro hydraulic actuator systems, often
re-ferred to as eha, is a xed displacement bidirectional hydraulic
pump.
15
-
Fluid Power Systems for Mobile Applications
An electric motor is usually used to power the pump, enabling
activecontrol of the rotational speed and thereby the ow to the
actuator,see gure 2.11. A conventional eha requires a symmetrical
actuator inorder to ensure ow balance. In that case, no additional
oil reservoirs orcontrol valves are needed [10]. However, the pump
in mobile applicationsis usually powered by and mechanically
coupled to an internal combus-tion engine. Moreover, asymmetrical
cylinders are predominantly usedin mobile applications. Some
modications of the conventional eha lay-out are therefore needed.
Solutions for handling asymmetrical cylindershave been proposed
[18].
Figure 2.11 Simplied schematic of an electro hydraulic actuator
sys-tem. An electric motor is controlling the rotational speed of a
xed bidi-
rectional pump.
In eha systems, there are no metering losses and the pump only
oper-ates when control action is needed. One disadvantage of using
a speedcontrolled xed displacement pump is that the volumetric
eciency is of-ten compromised at low pump speeds. However, the
overall eciency ineha systems is still higher than in, for example,
pump controlled actua-tor systems due to the pumps low eciency at
small displacements [19].The eha technology has become
well-established in the aerospace indus-try due to its eciency and
compactness.
Advantages High energy eciency and does not require a
variablepump.
Disadvantages Requires one hydraulic machine and an electric
motorfor each actuator. Low damping since the ow is load
independent.
16
-
Mobile Working Hydraulic Systems
2.3 System Summary
The systems described in this chapter are all used in dierent
appli-cations. Open-centre and constant pressure systems can be
consideredrather simple and often inecient system layouts. In such
systems, thecomponent costs themselves are often important and
eciency might beof less importance. Load sensing systems are often
a good compromisebetween eciency and complexity, but suer from poor
dynamic charac-teristics due to their closed loop pressure control.
More details regardingthe dynamics of load sensing systems can be
found in chapter 4.
When more than one load is actuated, often only the heaviest
loadcan be operated eciently in single pump systems. This issue is
solvedin valveless systems. When all loads have their own dedicated
pump,the pressure can always be matched against the present load.
It canbe realized with transformers, pump controlled actuators or
eha, whereeha is the most ecient solution. However, one has to bear
in mind thatvalveless systems might require several valves to
handle, for example,asymmetrical cylinder actuation and safety
requirements [17] [19].
This thesis proposes a system design that can be placed
somewherebetween load sensing and pump controlled actuators, see
gure 2.12. Itis called ow control system. It is similar to pump
controlled actuatorsbecause the pump displacement setting is used
to control the speed ofthe actuators. However, only one system pump
is needed and it usessimilar valves to load sensing systems. A
complete description of owcontrol systems is given in the following
chapters.
17
-
Fluid Power Systems for Mobile Applications
Complexity
Ener
gye
cien
cy
Open-centre
Constant pressure
Load sensing
Traditional
Pump controlled actuators
Independent metering
Flow control
Hydraulic transformers
eha
Intelligent
Present
and
per
form
ance
hydraulics
hydraulics
Figure 2.12 The system design analysed in this theses, ow
controlsystem, can be placed somewhere between load sensing and
pump con-
trolled actuators in terms of energy eciency and
performance.
18
-
3The Flow Control
Concept
In mobile hydraulic systems, the actuation of dierent loads is
controlledby joystick signals. These signals pose either a ow or
pressure demandfrom the operator. In applications where velocity
control is important,the signals from the operator often correspond
to ow demands. Anexample is load sensing systems equipped with
pressure compensators.The compensators maintain a constant pressure
drop across the direc-tional valves, which make the signals from
the operator correspond toow demands. Nevertheless, the pump in
these kinds of systems is stilloften pressure controlled.
In systems where the operators signals correspond to ow demands,
itseems more natural to also control the pump by ow. This approach
hassome benets regarding energy eciency, dynamic characteristics
andincreased exibility compared to load sensing systems. It also
presentssome challenges, for example the compensator design.
The idea of ow control is to use the joystick signals to control
thepump ow and the valve opening simultaneously. This means that
thepump software needs information about the ow demands for
dierentfunctions. The pump displacement setting is controlled
according to thesum of all requested load ows.
When no function is activated, the pump is de-stroked,
delivering noow to the system, and all directional valves are
closed. Activating ajoystick will simultaneously open a valve and
increase the displacementof the pump. Pressure is built up in the
pump hose and when the pumppressure becomes higher than the load
pressure there will be a ow to
19
-
Fluid Power Systems for Mobile Applications
the actuator. When stationary, the ow delivered by the pump will
go tothe load. The pump pressure will therefore adapt itself to a
level neededby the system. This results in eciency improvements
compared to loadsensing systems, which are described in more detail
in section 3.3.
Figure 3.1 Simplied schematic of a ow control system. The
pumpdisplacement setting and the valve openings are controlled
simultaneously
by the operators joystick signals.
Flow control systems will suer from load dependency if more
thanone load is activated simultaneously. This can be solved by
introducingsensors into the system. Zhe [20] used the velocities of
the actuatorsas the main feedback signals for pump and valve
control. Jongebloed etal. [21] used pressure sensors at all load
ports for the valve control. Tooptimize energy eciency, the valve
at the highest load can be opened toits maximum while lighter loads
are controlled by their valve openings.
Load dependency can also be solved by using pressure
compensators.Since the pump is ow controlled, there will be dierent
demands on thecompensator functionality compared to load sensing
systems. However,it also opens up new possibilities regarding the
control of the directionalvalves. Details regarding the compensator
requirements and the controlapproaches are described in sections
3.1 and 3.2.
Flow control systems have many similarities with load sensing
systems.Except for the pump controller, the two systems are almost
equivalent.
20
-
The Flow Control Concept
The pump controller used in ow control systems could also be
usedin pump controlled actuators. All components needed in ow
controlsystems are therefore available on the market [22].
3.1 Pressure Compensators
In some mobile uid power applications, load dependency and load
in-teraction are undesired system characteristics. An example is
forestrymachines, where the operator wants to position the load
with accuracy.Pressure compensators are commonly used in these
kinds of applicationsto ensure good handling capabilities. Two
dierent types of compen-sators can be realized, which are explained
in section 3.1.1 and 3.1.2.In applications with less demand for
accuracy, it is also possible to takeadvantage of ow forces for the
pressure compensation functionality.
3.1.1 Traditional Compensators
The most common design is to place the compensator upstream of
thedirectional valve. The reduced pressure is then working against
the loadpressure and a preloaded spring, see gure 3.2a. The force
equilibriumfor the compensator, equation (3.1), together with the
ow equationgives the ow across the directional valve. According to
equation (3.2),the compensator spring force sets the pressure drop
across the directionalvalve, making the ow load independent.
Fs + Ac1pL = Ac1pr pr pL =FsAc1
(3.1)
qL
= CqAs
2
(pr pL) = CqAs
2
(FsAc1
)(3.2)
It is also possible to achieve the same functionality by placing
thecompensator downstream of the directional valve. In that case,
thesupply pressure is working against the reduced pressure and a
springaccording to gure 3.2b. The force equilibrium, equation
(3.3), togetherwith the ow equation gives the same result, equation
(3.2) comparedwith (3.4).
Fs + Ac1pr = Ac1ps ps pr =FsAc1
(3.3)
21
-
Fluid Power Systems for Mobile Applications
ps pr pL
Ac1
Ac1Fs
As qL
(a) The compensator is placed up-stream of the directional
valve.
ps
pr
pL
Ac1
Ac1Fs
As qL
(b) The compensator is placeddownstream of the
directionalvalve.
Figure 3.2 Two dierent ways of realizing a traditional pressure
com-pensator. The pressure drop across the directional valve is set
by the
compensator spring force.
qL
= CqAs
2
(ps pr) = CqAs
2
(FsAc1
)(3.4)
These types of compensators are designed for use with a pressure
con-trolled pump. In case of the pump being saturated, the supply
pressurewill drop, resulting in the compensator spool at the
heaviest load open-ing completely. That function will lose speed
and possibly even stop.Functions operated simultaneously at lower
pressure levels will, however,move normally.
3.1.2 Flow Sharing Compensators
Another design is to implicate the highest load pressure into
the compen-sator. When the pressure is actively controlled, this
design is equivalentto the traditional compensator design. However,
its characteristics aredierent when the pump is saturated. All
functions will then be giventhe same priority, which means that all
functions will decrease in speed.This ow sharing functionality can
be achieved by placing a compensatoreither downstream or upstream
of the directional valve.
In case of the compensator being located downstream of the
direc-tional valve, the reduced pressure is working against the
highest loadpressure and a spring, see equation (3.5) and gure
3.3a. The pumppressure margin is dened according to equation (3.6)
and the ow canbe calculated according to equation (3.7).
Ac1pr = Ac1pLmax + Fs pr = pLmax +FsAc1
(3.5)
22
-
The Flow Control Concept
pp = ps pLmax (3.6)
qL
= CqAs
2
(ps pr) = CqAs
2
(pp Fs
Ac1
)(3.7)
ps pr pL
Ac1
Ac1Fs
As qL
pLmax
(a) The compensator is placeddownstream of the
directionalvalve.
ps
pr pL
Ac1
Ac1
As qL
pLmax
Ac2
Ac2Fs
(b) The compensator is placed up-stream of the directional
valve.
Figure 3.3 Two dierent ways of realizing a ow sharing pressure
com-pensator. The pressure drop across the directional valve is set
by the pump
pressure margin.
The ow sharing pressure compensator placed upstream of the
direc-tional valve is similar to its traditional equivalent.
Instead of a spring,two pressure signals that constitute the pump
pressure margin are act-ing on the compensator, see gure 3.3b.
Equation (3.6) together withthe force equilibrium for the
compensator, equation (3.8), gives the owaccording to equation
(3.9). The spring in this type of compensator isnot required for
the functionality. It can rather be used as a designparameter for,
for example, prioritization [23].
Ac2ps + Ac1pL = Ac2pLmax + Ac1pr + Fs
(pr pL) =Ac2Ac1
(ps pLmax ) FsAc1
(3.8)
qL
= CqAs
2
(pr pL) = CqAs
2
(Ac2Ac1
pp FsAc1
)(3.9)
Flow sharing pressure compensators will distribute the entire
pumpow relative to the individual valve openings also when the pump
issaturated. A pressure controlled pump which has been saturated
cannot
23
-
Fluid Power Systems for Mobile Applications
control the pressure and can therefore be seen as a ow
controlled pump.These compensators are therefore appropriate to use
together with a owcontrolled pump.
3.2 Pump and Valve Control Approaches
In ow control systems, the operators joystick signals control
the pumpow and the valve opening simultaneously. For this to work
properly,the system software needs knowledge about every ow
consumer in thesystem. However, solutions for attaching auxiliary
functions withoutknowledge about their ow demand have been
presented in [24] and [25].Dierent control approaches are possible
depending on whether tradi-tional compensators or ow sharing
compensators are used.
3.2.1 Flow Control using Traditional Compensators
When all directional valves are closed, the pump is ideally
de-strokedto zero, delivering no ow to the system. When the
operator movesthe joystick, signals are sent to the pump and the
valve simultaneously.The valve is shifted from neutral position and
the pump starts to deliverow. Since the valve is traditionally
pressure compensated, the springforce sets the pressure drop across
the directional valve, and thereby theabsolute ow level that the
valve is expecting, see gure 3.4. When thepump is delivering ow,
pressure is built up in the hose connecting thepump and the valve.
There will be a ow to the load when the pumppressure is higher than
the load pressure. This works ne as long as theow sent by the pump
equals the ow expected by the valve. If this isnot the case, two
situations may occur.
The pump ow is too low This is the same case as when the pump
issaturated in a load sensing system. The consequences will be
thatthe compensator spool at the highest load will open
completely,resulting in a decrease in speed for that load. It will
possibly evenstop.
The pump ow is too high Both compensator spools will close
moreand the pump pressure will increase until the system relief
valveopens. The throttle losses will be huge and the system will
emergeas a constant pressure system.
24
-
The Flow Control Concept
Figure 3.4 Simplied schematic of a ow control system using
tra-ditional pressure compensators. The system can also be realized
with
traditional compensators placed downstream of the directional
valves.
The reason for this is that traditional pressure compensators
controlthe absolute ow across the directional valve by reducing the
pumppressure relative to the load pressure of its own load. This
works neas long as the pump pressure is actively controlled, with
for instance aload pressure feedback. Otherwise, the ow situation
in the system isover-determined.
A lot of research solving this ow matching problem has been
pre-sented. Djurovic and Helduser [26] introduce a position sensor
placedon the directional valve. It allows precise knowledge of the
ow expectedby the valve. It is also possible to equip the
compensator with a positionsensor [27]. If no compensator is close
to fully opened, the pump ow istoo high. In case of the pump ow
being too low, the compensator atthe highest load would be
completely opened. A bleed-o valve to tankis proposed by several
authors [24] [26] [27]. A small overow is then ac-ceptable, which
could be used in closed loop control if a position sensoris added.
Fedde and Harms [28] discuss the pros and cons with overow
25
-
Fluid Power Systems for Mobile Applications
and underow when using a bleed-o valve. Grsbrink et al. [29]
[30]propose a system design where the pump is pressure controlled
for lowpump ows and ow controlled for high ow rates. It is also
possible toshift from ow control to pressure control in case of an
undesirable press-ure build up [31]. A review of solutions to the
ow matching problem inow control systems using traditional
compensators has been made byDjurovic in [32].
3.2.2 Flow Control using Flow Sharing Compensators
There are alternatives to address this ow matching problem
withoutadding additional components or sensors to the system. The
key is toimplicate the highest load pressure into the compensator
and thus getthe ow sharing behaviour described in section 3.1.2.
The compensatorsthan act as relief valves instead of reducing
valves and all valve sectionswill work against the highest load
pressure, see gur 3.5. This has beenstudied in, for example, [22]
and [33].
Figure 3.5 Simplied schematic of a ow control system using
owsharing pressure compensators. The system can also be realized
with ow
sharing compensators placed downstream of the directional
valves.
26
-
The Flow Control Concept
Using a ow controlled pump in combination with ow sharing
press-ure compensators opens up new possibilities in terms of
controlling thedirectional valves independently of the cylinder
velocity. This can beexplained with a small example. Imagine that
two loads are active, therst with 50% of the maximum velocity and
the second with 25% ofthe maximum velocity. The directional valves
have an opening area of50% and 25% and the ow delivered by the pump
is constant. Bothdirectional valve openings are now increased, the
rst to 100% and thesecond to 50%. The pump ow is still the same.
Since the ow sharingpressure compensators will distribute the
entire pump ow relative tothe individual valve openings, the
velocities will be unchanged. Whathappens is that the absolute
pressure drop across both directional valveshas been reduced, see
gure 3.6.
0 0.2 0.4 0.6 0.8 10
0.2
0.4
0.6
0.8
1
Pump flowVelocity, load 1Velocity, load 2
Flow
and
velo
city
[-]
Time [-]
(a) The pump ow and both actu-ator velocities are constant.
0 0.2 0.4 0.6 0.8 10
0.2
0.4
0.6
0.8
1
Pressure dropOpening area, load 1Opening area, load 2
Time [-]
Pre
ssure
dro
pan
dop
enin
gar
ea[-]
(b) The pressure drop across the di-rectional valves will
decrease whenthe opening areas are increased.
Figure 3.6 Flow sharing system characteristics. Both directional
valveopening areas are increased without aecting the actuator
velocities. The
pressure drop across both directional valves will decrease.
This system characteristic is dierent from most other valve
controlledsystems. Instead of controlling the ow, the valves will
serve as owdividers. One control approach is to open the valve
section at the loadwith the highest ow demand to its maximum [34]
[35]. Other activefunctions must always be opened in proportion to
its ow request. Thisapproach will minimize the pressure drop across
the directional valvesand thus save energy. This is further
discussed in section 3.3.
Another control approach might be to use the valves to increase
thedamping of the system. There is an optimal valve opening where
the
27
-
Fluid Power Systems for Mobile Applications
damping is maximized. For example, when a function is
oscillating thevalve opening could be reduced temporarily in order
to dampen theoscillations. When no oscillations are present, a more
energy ecientcontrol approach can be used. This is further
discussed in section 4.3.1.
3.3 Energy Eciency
The energy eciency of ow control systems is similar to load
sensingsystems. The pump pressure is adjusted according to the
highest loadand high losses might occur when loads with dierent
pressure demandsare operated simultaneously. However, instead of a
prescribed pressuremargin, as in load sensing systems, the pressure
drop between pump andload is given by the resistance in the hoses
and in the valves. Further-more, it is also possible to lower the
pressure drop across the directionalvalve by means of a more energy
ecient control strategy.
In load sensing systems, the pump pressure margin is set to
overcomethe losses in the pump hose, the compensator and the
directional valve.These losses are system dependent and will change
with internal andexternal conditions such as temperature, oil
properties, hose length, etc.The pressure margin is set according
to the worst case to ensure it ishigh enough at all operating
points.
The pressure drop between pump and load can be divided into
threedierent losses:
Losses between pump and valve There will be a pressure drop
be-tween the pump and the valve. The magnitude will depend onthe
internal and external properties mentioned above, but
mostimportantly the ow rate. A simplied model is that the
lossesincrease with the square of the ow rate.
Losses across the compensator There will be a pressure drop
acrossthe compensator. High losses occur if the supply pressure is
muchhigher than the load pressure. This is the case at partial
loadingconditions. The smallest possible loss occurs when the
compen-sator is fully opened. In that case, the required pressure
dropincreases with the square of the ow rate.
Losses across the directional valve Typically, the
compensatormakes sure that the pressure drop across the directional
valveis constant. However, the smallest possible pressure drop
occurs
28
-
The Flow Control Concept
if the valve is fully open. The pressure drop will then follow
theow equation, similar to the compensator pressure drop.
In gure 3.7a, these three dierent losses are shown. If the
pressuremargin is set perfectly, there would be no unnecessary
losses at maxi-mum ow rate in load sensing systems. However, at
lower ow rates,unnecessary losses will occur. In ow control
systems, these losses willbe eliminated since the pump pressure is
set by the resistance in thehose and the valve.
It is possible to further reduce the losses in ow control
systems. Thisis done by opening the valve section with the highest
ow demand to itsmaximum, as described in section 3.2.2, in which
case the pressure dropacross the directional valve is minimized and
additional energy savingsare possible, see gure 3.7b.
A ow control system without pressure compensators would
increasethe eciency even further. In that case, the valve section
at the highestload pressure might be opened completely. However,
its functionalityrequires closed loop control and is therefore
sensor dependent [21].
Pum
ppre
ssure
marg
in[-]
Flow [-]
unneces
saryloss
es
directional valve losses
hose lo
sses
com
pen
sato
rlo
sses
(a) The pump pressure marginis xed in load sensing
systems.Therefore, unnecessary losses occurat lower ow rates.
Pum
ppre
ssure
marg
in[-]
Flow [-]
ecienc
y improv
ements
fully opened directional valve
hose lo
sses
com
pen
sato
rlo
sses
(b) The pump pressure margin isgiven by the system resistances
inow control systems. Eciency im-provements are therefore
possible.
Figure 3.7 Classication of the losses between pump and load.
Threedierent losses occur; hose, compensator and directional valve
losses. At
lower ow rates, unnecessary losses occur in load sensing
systems. No
unnecessary losses occur in ow control systems.
As can be seen in gure 3.7, the two system layouts have the same
e-ciency at maximum ow rate if the pump pressure margin is set
perfectlyin the load sensing system. Flow control systems have
higher eciencyfor smaller ow rates. However, it is important to
consider the power
29
-
Fluid Power Systems for Mobile Applications
losses rather than the pressure losses. For low ow rates, the
power losswill be small even for high pressure drops. Figure 3.8
shows the powersaving opportunities for ow control systems. The
largest power savingsoccur in the medium ow rate area. If the
directional valve is openedcompletely, even more power can be
saved.
Pow
er[-]
Flow [-]
fully openeddirectional valve
power savings
Figure 3.8 Power savings in ow control systems compared to
loadsensing systems. More power can be saved if the directional
valve is com-
pletely opened. No power is saved at maximum ow rate.
Flow control systems have no unnecessary losses for the highest
load.All losses that occur are necessary and limited by, for
example, the diam-eter of the hoses and the maximum opening areas
in the valve. However,ow control systems still have high losses at
partial loading conditions.To increase eciency even further,
individual metering valves or addi-tional hydraulic machines are
required.
A ow control system with two hydraulic pumps has been studiedin
[36] and [37]. The aim is to reduce the losses at partial
loadingconditions without increasing the total installed
displacement. This isachieved by connecting the two pumps when high
ow rates are requiredby one load. Connecting several pumps at high
ow rates is a commonsolution for more simple systems, for example,
in excavators.
30
-
4Dynamic Analysis
The dynamic analyses in this thesis were made to show the
fundamentaldierences between load sensing systems and ow control
systems. Lin-ear models are used and dierent types of compensators
are consideredin the analysis. The only dierence between the load
sensing systemmodel and the ow control system model is the absence
of the feed-back to the pump controller in the ow control system,
see gures 4.1and 4.2. Nevertheless, there are fundamental dynamic
dierences be-tween the two system layouts.
Qa AcVa,Pa
Vb,Pb
Qb
mL
U
Kcb
Qp
Vp,Pp
GpLS
Kca
Ppref
Xp
Figure 4.1 Dynamic load sensing system model.
31
-
Fluid Power Systems for Mobile Applications
Qa AcVa,Pa
Vb,Pb
Qb
mL
U
Kcb
Qp
Vp,Pp
GpFC
Kca
Qpref
Xp
Figure 4.2 Dynamic ow control system model.
4.1 Mathematical Model
A linear mathematical model is constructed to perform the
dynamicanalyses. The derivation of the equations is shown in
[38].
The pump controller can be described in two dierent ways. In
loadsensing systems, the controller consists of a pressure
controlled valvethat controls the displacement piston. If the
pressure balance, Pp =Pp Pa, is disturbed, the valve is displaced
and the pump setting is thenproportional to the integrated valve
ow. Here, the pump is modelledas a pure inductance, see equation
(4.1).
GpLS =Qp
Ppref Pp=
1
Lps(4.1)
The pump controller in ow control systems controls the
displacement,and thereby the ow, directly instead of maintaining a
certain pressuremargin above the highest load pressure. Such a pump
controller has noexternal feedback from the system, similar to the
load sensing feedback.Here, the transfer function describing the
displacement controlled pumpdynamics is called GpFC , see equation
(4.2).
GpFC =Qp
Qpref(4.2)
The continuity equation of the pump volume yields the transfer
func-tion in equation (4.3).
Hs =Pp
Qp Qa =eVps
(4.3)
32
-
Dynamic Analysis
The model for the inlet orice in the directional valve will be
dierentdepending on the compensator design. A non-compensated valve
willhave a ow-pressure dependency according to equation (4.4). In
thisanalysis, the valve is considered to be much faster than the
rest of thesystem. The valve dynamics is therefore ignored. The
dynamics ofpressure compensated valves have been studied in, for
example, [39]and [40].
Gva =Qa
Pp Pa = Kca (4.4)
A traditionally compensated valve will have no ow-pressure
depen-dency since the pressure drop across the directional valve is
constant,see equation (4.5).
Gva =Qa
Pp Pa = 0 (4.5)
A ow sharing pressure compensated valve will have a
ow-pressuredependency, similar to a non-compensated valve, for the
highest load.Lighter loads have no ow-pressure dependency, like
traditional com-pensated valves. However, lighter loads will be
disturbed by the highestload due to cross-coupling of the highest
load pressure to all compen-sators [41].
Gva =Qa
Pp Pa = Kca , Pa = Pamax
Gva =Qa
Pp Pa = 0, Pa < Pamax (4.6)
Gvea =Qa
Pp Pamax= Kca , Pa < Pamax
A detailed investigation of valve models using dierent
compensationtechniques can be found in [41] and paper [II].
A mass load with a gear ratio is considered to act on a
cylinder. Thecontinuity equation for the cylinder chambers together
with the forceequilibrium for the piston is shown in equations
(4.7), (4.8) and (4.9).
Qa =Vae
sPa + AcsXp (4.7)
U2mLs2Xp + BpsXp = AcPa AcPb (4.8)
AcsXp Qb =Vbe
sPb (4.9)
33
-
Fluid Power Systems for Mobile Applications
It is also possible to describe a load which consists of a
hydraulic motorby similar equations [II].
The outlet orice in the directional valve is considered to have
a ow-pressure dependency according to equation (4.10).
Gvb =QbPb
= Kcb (4.10)
4.2 Pump Stability
Due to the absence of the load pressure feedback to the pump
con-troller in ow control systems, there is a fundamental dynamic
dier-ence between load sensing and ow control systems. To show
this, themathematical model in section 4.1 can be simplied. A
ow-pressuredependency at the inlet side of the valve is assumed and
the outlet ori-ce is ignored. The simplications will not inuence
the fundamentaldierences but is important to bear in mind when
making other dynamicanalyses.
A transfer function from inlet ow to pressure in the cylinder
canbe derived using equations (4.7) and (4.8). Ignoring the outlet
oriceresults in a constant pressure on the piston rod side.
ZL
=PaQa
=U2m
Ls + Bp
Vae
U2mLs2 + Vae Bps + A
2c
(4.11)
4.2.1 Load Sensing Systems
The dynamic behaviour of load sensing systems can be described
byequations (4.1), (4.3), (4.4) and (4.11). By reducing the block
diagramin gure 4.3a, the open loop transfer function from desired
pump press-ure margin, Ppref , to actual pressure dierence, Pp = Pp
Pa, canbe derived according to equation (4.12). A complete
investigation ofload sensing systems and their dynamic properties,
including pump con-trollers, can be found in [42].
GpLSGo = GpLSHs
1 + Gva (ZL + Hs)(4.12)
By closing the control loop, the pump controller, GpLS , is a
part of theloop gain, GpLSGo, as shown in gure 4.3b. To achieve a
stable systemthe loop gain must be kept lower than unity when the
phase crosses
34
-
Dynamic Analysis
+
GpLS+
Hs+
Gva
ZL
1Gva
Pp,ref Qp Pp Pp Qa
Pa
(a) Block diagram of a load sensing system derived from the
transferfunctions (4.1) (pump controller), (4.3) (pump volume),
(4.4) (inlet valve)and (4.11) (load).
+
GpLS Go Pp,ref Pp
(b) Rearranged block diagram with theloop gain GpLS Go.
Figure 4.3 Linear model of a load sensing system.
-180. On the other hand, it would be feasible to increase the
gain ofthe pump and its controller to achieve a system that meets
the responserequirements. To achieve a system, with desired
response, the gain ofthe pump controller is increased, but at the
same time the system isapproaching its stability limit. One should
bear in mind that stabilityat one operational point will not
guarantee stability at another, seegure 4.4.
4.2.2 Flow Control Systems
The dynamic behaviour of ow control systems can be described
byequations (4.2), (4.3), (4.4) and (4.11). This results in almost
the sameblock diagram as in gure 4.3. The only dierence is the
absence ofthe feedback to the pump controller, see gure 4.5b. This
results in afundamental dynamic dierence between load sensing
systems and owcontrol systems. Since there is no closed loop for
the pump controller,the stability issues described in section 4.2.1
are eliminated. The pumpand its controller can thereby be designed
to meet the response require-
35
-
Fluid Power Systems for Mobile Applications
100 101 102 103 1046
5
4
3
2
1
0
1
100 101 102 103 104
270
180
90
log10
(Gp
LSG
o)
[-]
Phas
e[]
Frequency [rad/s]
mL
increasing
mL
increasing
Figure 4.4 Bode plot of the open loop gain in gure 4.3b,
GpLSGo.
Table 4.1 Parameter values used in gure 4.4.
Parameter Value Unity
Ac 0.008 m2
Bp 10000 Ns/mVa 4103 m3Vp 5103 m3Kca 1109 m5/NsLp 5108 Pa
s2/m3m
L[6000 12000 30000] kg
U 1 -e 1109 Pa
ments without considering system stability. This has been veried
byexperiments in [22] and [34].
36
-
Dynamic Analysis
GpFC+
Hs+
Gva
ZL
Qpref Qp Pp Pp Qa
Pa
(a) Block diagram of a ow control system derived from the
transferfunctions (4.2) (pump controller), (4.3) (pump volume),
(4.4) (inletvalve) and (4.11) (load).
GpFC Go
Qpref Pp
(b) Rearranged block diagramwith no feedback present.
Figure 4.5 Linear model of a ow control system.
4.3 Damping
Hydraulic systems by themselves are normally poorly damped and
needsome additional damping from the valves to prevent, or at least
reduce,the tendency to oscillate. To obtain damping from a valve,
the owshould increase when the pressure drop across the valve
increases andvice versa. Andersson [43] gives an overview of the
valves contributionto damping in mobile hydraulic systems. An
overview of active oscil-lation damping of mobile machine structure
is given by Rahmfeld andIvantysynova in [44].
Open-centre and constant pressure systems have a high damping
asdescribed in section 2.1. Load sensing systems are poorly damped,
es-pecially if pressure compensators are used. Valveless systems
are ideallyundamped since no valves are present in those kinds of
systems.
4.3.1 Active Control of the Inlet Orice
In this section, the damping contribution of the inlet orice in
owcontrol systems is analysed. The cylinder friction and the outlet
oriceare ignored to simplify the analysis, see gure 4.6. The inlet
valve isassumed to have a ow-pressure dependency, which means that
it could
37
-
Fluid Power Systems for Mobile Applications
be a non-compensated valve or a ow sharing valve at the highest
loadaccording to equation (4.6).
Qa AcVa,Pa
mL
U
Qp
Vp,PpKca
Xp
Figure 4.6 Dynamic model of a ow control system with a mass
load.The outlet orice and the cylinder friction have been
ignored.
The system can be described by equations (4.2), (4.3), (4.4) and
(4.11).An expression for the ow-pressure coecient of the inlet
orice thatgives the highest damping has been derived in paper [II]
according toequation (4.13).
Kcaopt =Ac
3/4i
Vp (i 1)eU2mL
(4.13)
where
i= 1 +
VpVa
(4.14)
The maximum damping in the system can be calculated using
equa-tion (4.15).
hmax =1
2
(
i 1) (4.15)
Equation (4.15) shows that the maximum damping given by the
inletorice only depends on the value of
i, which includes the pump hose
volume and the volume at the inlet side of the cylinder
according toequation (4.14). To get a high damping contribution
from the inletorice, the pump hose volume should be large compared
to the volumeon the inlet side of the cylinder. However, this
relationship will change
38
-
Dynamic Analysis
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
Dam
pin
g[-]
Increased opening area
iincreasing
Figure 4.7 System damping as a function of the opening area of
theinlet orice. A small value of
igives a low damping regardless of the
opening area. The damping will increase with higher values of
i.
during the cylinder stroke. The damping as a function of the
inlet oriceopening area for dierent values of
iis shown in gure 4.7.
To get the highest possible damping for a given value of i, the
inlet
orice opening area has to be small. During certain points of
opera-tion this might result in substantial power losses [II]. To
avoid this itis possible to use the more energy ecient control
strategy describedin section 3.2.2 while no oscillations are
present. When damping is re-quired, the valve can temporarily be
closed more to reach the peaks ingure 4.7. Finally, when the
oscillations have died out, the energy e-cient control strategy can
be applied again. This is possible to do in owcontrol systems
without aecting the cylinder velocities if ow sharingpressure
compensators are used.
Theoretically, a ow control system using traditional
compensatorsobtains no damping from the inlet orice since the ow is
independentof pressure changes, see equation (4.5). This is also
true for lower loadsusing ow sharing compensators according to
equation (4.6). One wayto obtain damping for such loads is to
implement active damping, usingfor example a dynamic load pressure
feedback.
A special case of this analysis is when the inlet orice opening
areaapproaches innity. This is the case in valveless systems, which
have no
39
-
Fluid Power Systems for Mobile Applications
orices at all. As can be seen in gure 4.7, the damping then
approacheszero. Consequently, a valveless system is ideally
undamped.
4.3.2 Design and Control of the Outlet Orice
In this section, the damping contribution of the outlet orice is
analysed.This analysis is not limited to ow control systems, but is
valid for allpump controller designs. The prerequisite is that the
inlet ow can bemodelled as a perfect ow source, which is true if
the inlet orice has noow-pressure dependency, see gure 4.8. This
can be realized with, forexample, a traditional pressure
compensator.
Pa
Qa
AcVa
Vb
Qb
Pbm
L
U
Kcb
Figure 4.8 Dynamic model of a ow controlled cylinder with a
massload and an outlet orice. The pump controller can be of any
design; it
does not aect the analysis.
The system can be described by equations (4.7)-(4.10). Similar
to theanalysis in section 4.3.1, the viscous friction in the
cylinder has beenignored to simplify the analysis. An expression
for the ow-pressurecoecient of the outlet orice that gives the
highest damping has beenderived in paper [III] according to
equation (4.16).
Kcbopt = Ac
Vb
eU2mL (o 1)3/4
o(4.16)
where
o = 1 + 2VaVb
(4.17)
40
-
Dynamic Analysis
The maximum damping in the system can be calculated using
equa-tion (4.18).
hmax =1
2
(o 1
)(4.18)
Equation (4.18) shows that the maximum damping of the system
de-pends only on the value of o , which includes the volume at each
side ofthe cylinder and the cylinder area ratio according to
equation (4.17). Toget a high damping contribution from the outlet
orice the volume onthe inlet side of the cylinder should be large
compared to the volume onthe outlet side. However, this
relationship will change during the cylin-der stroke. A high value
of the cylinder area ratio increases the damping,which means that a
symmetrical cylinder gives higher damping than anasymmetrical. The
damping as a function of the outlet orice openingarea for dierent
values of o is shown in gure 4.9.
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
Dam
pin
g[-]
Increased opening area
o increasing
Figure 4.9 System damping as a function of the opening area of
theoutlet orice. Small values of
ogive a low damping regardless of the
opening area. The damping will increase with higher values of
o.
A valve design that is suggested in paper [III] is to optimize
the damp-ing when the piston is at its lower end position. While
the piston movesupwards, the damping will increase. If a higher
damping is required, itis possible to design the valve with a
smaller orice area. The drawbackswith such a design are, however,
that the damping will be slightly lowerat the pistons lower end
position and that the losses across the outlet
41
-
Fluid Power Systems for Mobile Applications
orice will be higher. If lower losses are required it is
possible to designthe valve with a larger opening area. However,
this is at the expense ofa lower damping. There is no point in
designing the valve with a toosmall orice area. The damping will
then be low and the losses high.
In case of the inlet and outlet orices being decoupled, as in
indi-vidual metering systems, it would be possible to optimize the
dampingduring the cylinder stroke. While the piston is moving, the
outlet oricecould be controlled in order to achieve the highest
possible damping. Itwould also be possible to use a similar control
approach as described insection 4.3.1. When no oscillations are
present, the outlet orice couldbe fully opened, minimizing the
losses. When damping is required, thecontroller could shift to
optimize the damping and temporarily allowhigher losses.
42
-
5Experimental
Results
The energy eciency improvements described in section 3.3 have
beenvalidated using a wheel loader application. Also, the theories
concerningthe design and control of the outlet orice described in
section 4.3.2 havebeen validated in a test rig.
5.1 Energy Eciency Improvements
5.1.1 Hardware Requirements
The hardware requirements in ow control systems are similar to
loadsensing systems. To achieve the same system capacity, a pump
size ofthe same magnitude is used. Only the pump controller needs
to be dif-ferent. Instead of actively maintaining a certain
pressure margin abovethe highest load pressure, the pump
displacement is controlled directlyfrom the operators demand
signals. This requires an electrically con-trolled displacement
controller for the pump. However, the load sensinghose to the pump
controller can be removed.
Flow control systems use the same type of valves as load sensing
sys-tems. Flow sharing pressure compensators are favourable but
work-arounds with traditional compensators also exist, see section
3.2.1. Insome valve designs, a traditional compensator placed
upstream of thedirectional valve can be replaced with its ow
sharing equivalent withouteven replacing the valve housing
[23].
Sensors are not required to achieve the desired functionality in
ow
43
-
Fluid Power Systems for Mobile Applications
control systems if pressure compensators are used. However, it
wouldbe benecial to use sensors to detect if the cylinder end stops
have beenreached. In that case, the valve could be closed and the
pump owadjusted to avoid unnecessary energy losses.
5.1.2 A Demonstrator System
To verify the energy eciency improvements in the ow control
concept,measurements where performed on a wheel loader application,
see g-ure 5.1. The machine was equipped with a pump that can be
operatedin both pressure and ow control modes and a valve prepared
for usewith both traditional and ow sharing compensators, placed
upstreamof the directional valve.
Figure 5.1 The machine used for experiments.
In gure 5.2c, the pump pressure margin for both the load sensing
andthe ow control systems can be seen. The measurements agree with
thetheoretical pressure margin shown in gure 3.7. The ow sent by
thepump is similar in both systems, see gure 5.2a. It can also be
observedin gure 5.2b that the pressure is more oscillative in the
load sensingsystem. This is because the pump controller operates in
a closed loopcontrol mode [42].
A short loading cycle has also been performed to compare load
sensingand ow control. Only the working hydraulics have been taken
into con-sideration, neither the steering nor the transmission.
Figure 5.3a showsthe position of the actuators and gure 5.3b the
energy consumption.The energy consumption was reduced by 14% for
the ow control systemfor this particular application. This is the
same order of magnitude asexperiments performed in [24] and
[34].
44
-
Experimental Results
0 1 2 3 4 5 60
50
100
150
Flow
[l/m
in]
Time [s]
(a) Measured ow for both systems.The ow is increased from zero
tomaximum.
0 1 2 3 4 5 60
5
10
15
20
25
30
35
40
45
50
Time [s]
Pre
ssure
[bar]
(b) Measured pump pressure mar-gin for both systems while the
owis increased.
0 20 40 60 80 1000
5
10
15
20
25
30
Flow [l/min]
Pre
ssure
[bar]
Load sensing
Flow
contro
l
(c) Measured pump pressure mar-gin as a function of measured
ow.Load sensing systems have a con-stant margin while ow control
sys-tems have a margin given by the sys-tem resistances.
Figure 5.2 Experimental results showing the potential of
reducing thepump pressure margin in ow control systems compared to
load sensing
systems.
45
-
Fluid Power Systems for Mobile Applications
0 5 10 15 20 25 30 350
0.1
0.2
0.3
0.4
0.5
0.6
0.7
Pos
itio
n[m
]
Time [s]
Lift
Tilt
(a) Measured positions of the actu-ators during the cycle.
0 5 10 15 20 25 30 350
20
40
60
80
100
120
140
160
Time [s]
Ener
gy[k
J]
Load sensing
Flow control
(b) Measured energy consumptionduring the cycle.
Figure 5.3 Experimental results showing the actuator positions
andthe consumed energy in a short loading cycle. The ow control
system
consumed 14% less energy during the cycle compared to the load
sensing
system.
46
-
Experimental Results
5.2 Improved Damping
A test rig to validate the damping contribution by the outlet
orice hasbeen constructed. It consists of a traditional pressure
compensated valveon the inlet side, a cylinder with a mass load and
a servo valve on theoutlet side, see gure 5.4. Dierent designs of
the outlet orice can beachieved by controlling the opening area of
the servo valve. A constantpressure pump supplies the system.
Pressure sensors are attached onthe supply side and on both
cylinder chambers. The cylinder and theservo valve are equipped
with position sensors. External volumes aremounted on both sides of
the piston. By using either one, it is possibleto manipulate the
dead volumes on either side of the piston.
Figure 5.4 The experimental test stand. The pressure
compensatedvalve can be seen at the lower right and one of the
volumes to the left.
In the experiments, a step is made in the ow by opening the
inletvalve. Oscillations in the cylinder velocity are then studied.
The exper-imental results are presented in gure 5.5. In tests (a)
and (b), thereis a large volume on the inlet side which means that
a relatively highdamping can be expected. In test (a), the outlet
orice area is dimen-sioned close to the maximized damping. As can
be seen in gure 5.5a,there are almost no oscillations in the
cylinder velocity. In test (b), theoutlet orice area is larger than
in test (a) and the damping becomeslower, see gure 5.5b.
In test (c), there is a large volume on the outlet side of the
cylinder,which means that the damping is expected to be low. The
outlet oricearea is dimensioned close to the maximized damping.
Nevertheless, thedamping is still low according to gure 5.5c. This
is consistent with the
47
-
Fluid Power Systems for Mobile Applications
mathematical analysis according to equations (4.17) and
(4.18).In test (d), the outlet orice area is so large that it can
be equated
with having no outlet orice at all. Theoretically, the hydraulic
systemwill not contribute any damping without an outlet orice as
shown insection 4.3.2. This is almost the case in the measurements
as can be seenin gure 5.5d. The damping that is still obtained is
due to secondaryeects ignored in the mathematical analysis, such as
friction and leakage.
0 0.5 1 1.5 2 2.5 3
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Veloc
ity
[-]
Time [s]
(a) A high damping is obtainedwhen there is a large volume on
theinlet side of the cylinder and theoutlet orice is designed close
to itsoptimum.
0 0.5 1 1.5 2 2.5 3
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Veloc
ity
[-]
Time [s]
(b) The damping becomes lowerwhen there is a large volume at
theinlet side and the outlet orice areais too large.
0 0.5 1 1.5 2 2.5 3
0
0.5
1
1.5
2
2.5
3
3.5
Veloc
ity
[-]
Time [s]
(c) When there is a large volume atthe outlet side of the
cylinder, thedamping becomes low even if theoutlet orice is
designed close to itsoptimum.
0 0.5 1 1.5 2 2.5 3
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
1.8
2
Veloc
ity
[-]
Time [s]
(d) Without an outlet orice, onlysecondary eects such as
friction andleakage will contribute to the damp-ing.
Figure 5.5 Experimental results for dierent designs of the
outlet ori-ce.
48
-
6Summary and
Conclusions
This thesis studies a system design where the pump displacement
set-ting is controlled based on the operators command signals
rather thanmaintaining a certain pressure margin above the highest
load pressure.Conventional load sensing systems are
state-of-the-art in industry todayand are therefore used as
comparison base.
The fundamental dierence between ow control and load sensing
isthat the load pressure feedback hose to the pump controller can
beremoved. Instead of controlling the pump in a closed loop control
mode,an open control mode can be used with no feedbacks present.
Thismakes the system design process simpler since the pump can be
designedto meet the response requirements without considering
system stability.As long as the pump is stable as an isolated
component, it will not causeany stability issues in the complete
system. In load sensing systems onthe other hand, an apparently
stable pump can cause instability in thecomplete system.
Flow control systems are more energy ecient compared to load
sens-ing systems. This is because the pressure dierence between
pump andload is given by the system resistance rather than a
prescribed pumppressure margin. The two system layouts have the
same eciency whenthe pump is saturated. However, in all other
operational points, owcontrol systems have a higher eciency than
load sensing systems. Thereare also potential energy savings tied
to the absence of active control ofthe pump.
Pressure compensators are key components in ow control
systems.
49
-
Fluid Power Systems for Mobile Applications
Traditional compensators control the absolute ow through the
direc-tional valves. If the pump also controls the ow, the ow
situation inthe system is over-determined and ow matching problems
occur. Thiscan be solved by introducing sensors or a bleed-o valve.
Another solu-tion is to use ow sharing pressure compensators.
Instead of controllingthe ow, the valves will then serve as ow
dividers, eliminating the owmatching problem. Both traditional and
ow sharing compensators canbe realized by placing the compensators
either upstream or downstreamof the directional valves. A drawback
with ow sharing compensators isthat the highest load dynamically
will disturb all lighter loads.
Damping is a desired property in uid power systems. Low
dampingmakes the system oscillative, which has a negative impact on
both theproductivity and the operator. Two dierent ways to obtain
dampingin ow control systems are analysed in this thesis, either by
the inletorice or by the outlet orice in the directional valve. If
ow sharingcompensators are used, it is possible to control the
inlet orice, with-out aecting the actuator velocity, with the
objective to optimize thedamping. Active damping measures are
required to obtain damping bythe inlet orice if traditional
compensators are used. Some design rulesto obtain a high damping by
the outlet orice are proposed. From anenergy eciency point of view,
it is often better to obtain damping bythe outlet orice since a
smaller pressure drop is required.
It is possible to combine ow control with other working
hydraulicsystems. For example, ow control could be used as a
complement topump controlled actuators. Some high power consumers
could have onededicated pump while other, low power, consumers
share one commonpump. In that case, the total installed
displacement could be kept ata reasonable level while all pumps
could be displacement controlled.Another possibility could be to
use an electric motor in combinationwith a xed displacement pump,
like in eha systems, but share it withseveral loads. This solution
is more favourable in ow control systemsthan in load sensing
systems because a lower bandwidth is required.
Experiments have been performed to show the capability of the
owcontrol approach. For example, reduction of the superuous pump
press-ure margin and energy saving potentials in a short loading
cycle for awheel loader application have been demonstrated.
50
-
7Outlook
Today, both academia and industry devote a lot of eort to the
area ofenergy ecient uid power systems. This will most likely
continue inthe future. In the short term, ow control systems are a
complement,or alternative, to load sensing systems in particular
applications. Thechallenge is how, and to what extent, sensors
should be used. It ispossible to design a ow control system without
the need for sensors,but it might be desirable to use sensors in
some operational cases. Oneexample is when a cylinder reaches its
end stop. From an eciency pointof view, substantial power savings
are possible if the pump controller hassuch information feedback
from the consumers.
In the longer term, independent metering and valveless systems
willprobably gain market shares. Those systems are more energy
ecient,especially during partial loading conditions. Furthermore,
they also havethe possibility to recuperate energy from the loads.
This energy caneither be used to run the system pump as a motor or
to be stored in, forexample, an accumulator. One interesting system
layout in the futurecould be to use a xed displacement
bidirectional pump powered by anelectric motor in ow control
systems. This is a hybrid of eha and owcontrol.
Which hydraulic system to choose will always be a compromise
be-tween, for example, eciency, dynamic characteristics, complexity
andcost. It is possible that ow control systems will be the optimal
solutionfor certain functions in some particular applications.
51
-
Fluid Power Systems for Mobile Applications
52
-
8Review of Papers
In this chapter, the three appended papers in this thesis are
brieysummarized. Papers [I] and [II] analyse dierent aspects of ow
controlsystems and compare the ndings with load sensing systems.
Paper [III]is not limited to ow control systems; the conclusions
are valid for severalsystem designs.
Paper I
Energy Ecient Load Adapting System Without Load Sens-
ing Design and Evaluation
This paper studies dierent pressure compensator techniques in
owcontrol systems. The fundamental dierence between load sen