Conference PaperPDF Available

Abstract and Figures

This paper proposes a novel architecture for the pilot stage of electro-hydraulic two-stage servovalves that does not need a quiescent flow and a torque motor as well as a flexure tube to operate. The architecture consists of two small piezoelectric valves, coupled with two fixed orifices, which allow variation of the differential pressure at the main stage spool extremities in order to move it with high response speed and accuracy. Each piezoelectric valve is actuated by a piezoelectric ring bender, which exhibits much greater displacement than a stack actuator of the same mass, and greater force than a rectangular bender. The concept is intended to reduce the influence of piezoelectric hysteresis. In order to assess the validity of the proposed configuration and its controller in terms of spool positioning accuracy and dynamic response, detailed simulations are performed by using the software Simscape Fluids. At 50% amplitude the −90° bandwidth is about 150Hz.
Content may be subject to copyright.
1 Copyright © 2018 by ASME
Proceedings of the 2018 Bath/ASME Symposium on Fluid Power and Motion Control
FPMC 2018
September 12-14, 2018, University of Bath, Bath, United Kingdom
FPMC2018-8864
A NOVEL PIEZOELECTRIC DOUBLE-FLAPPER SERVOVALVE PILOT STAGE:
OPERATING PRINCIPLE AND PERFORMANCE PREDICTION
ABSTRACT
This paper proposes a novel architecture for the pilot stage of
electro-hydraulic two-stage servovalves that does not need a
quiescent flow and a torque motor as well as a flexure tube to
operate. The architecture consists of two small piezoelectric
valves, coupled with two fixed orifices, which allow variation of
the differential pressure at the main stage spool extremities in
order to move it with high response speed and accuracy. Each
piezoelectric valve is actuated by a piezoelectric ring bender,
which exhibits much greater displacement than a stack actuator
of the same mass, and greater force than a rectangular bender.
The concept is intended to reduce the influence of piezoelectric
hysteresis. In order to assess the validity of the proposed
configuration and its controller in terms of spool positioning
accuracy and dynamic response, detailed simulations are
performed by using the software Simscape Fluids. At 50%
amplitude the -90 bandwidth is about 150Hz.
Keywords: Servovalve, Piezoelectric, Ring bender, Simscape
NOMENCLATURE
A
Spool end area [mm2]
Aleak
Leakage area [mm2]
Ar
Restricted area [mm2]
Ar,0
Restricted area fixed orifice [mm2]
b
Width of the slots [mm]
C
Damping coefficient of the main spool [Ns/m]
Caprb
Capacitance of the ring bender [nF]
CD
Discharge coefficient of the main valve
CD,0
Discharge coefficient of the fixed orifice
CD,P
Discharge coefficient of the piezovalve
Crb
Damping coefficient of the ring bender [Ns/m]
Cstop
Damping coefficient hard stop [Ns/m]
D
Main spool diameter [mm]
Dint
Diameter of the hydraulic chamber [mm]
d
Diameter of the piezo valve orifice [mm]
E
Bulk modulus [N/m2]
E0
Pure liquid bulk modulus [N/m2]
e
Error
Fb,rb
Blocking force [N]
Fflow
Flow force [N]
Imax
Maximum current [A]
Ka
Gain of the amplifier
Kd,v
Max. blocking force over max. voltage [N/V]
KI
Integral gain
Kp
Proportional gain
Krb
Spring stiffness of the ring bender [N/m]
Ks
Additional spring stiffness [N/m]
Kstop
Stiffness hard stop [N/m]
Lint
Length of the hydraulic chamber [mm]
M
Main spool mass [kg]
m
Ring bender mass [kg]
n
Hysteresis non-linear term [V]
p
Pressure [N/m2]
pa
Ambient pressure [N/m2]
Q
Flow rate through the main valve[m3/s]
qc
Flow rate through the hydraulic chamber [m3/s]
Paolo Tamburrano
Department of Mechanics, Mathematics and
Management (DMMM), Polytechnic University of
Bari, Via Orabona 4, 70125, Bari, Italy
Riccardo Amirante
Department of Mechanics, Mathematics and
Management (DMMM), Polytechnic University of
Bari, Via Orabona 4, 70125, Bari, Italy
Elia Distaso
Department of Mechanics, Mathematics and
Management (DMMM), Polytechnic University
of Bari, Via Orabona 4, 70125, Bari, Italy
Andrew R. Plummer
Centre for Power Transmission and Motion
Control (PTMC), Department of Mechanical
Engineering, University of Bath, Claverton
Down, BA2 7AY, Bath, UK
2 Copyright © 2018 by ASME
q0
Flow rate through the fixed orifice [m3/s]
qs
Flow rate through the external chamber[m3/s]
qv
Flow rate through the piezo valve[m3/s]
Recr
Critical Reynolds number
t
Time [s]
u
Control output [V]
Vamp
Voltage from the amplifier [V]
Vc
Control Voltage [V]
Vcham
Oil volume in the hydraulic chamber [mm3]
Vdead
Dead volume [mm3]
Vleft
Oil volume in the left spool chamber [mm3]
Vo
Volume of the hydraulic chamber [mm3]
Vright
Oil volume in the right spool chamber [mm3]
X
Main spool displacement [mm]
x
Ring bender displacement [mm]
x0
Pre-compression of the additional spring [mm]
α
Parameter for the hysteresis formula
β
Parameter for the hysteresis formula
γ
Ratio of the specific heats
δ
Parameter for the hysteresis formula
ε
Relative gas content at atmospheric pressure
θ
Flow angle [rad]
ξ
Damping factor of the amplifier
Fluid kinematic viscosity [m2/s]
ρg0
Gas density at ambient pressure [kg/m3]
ρl
Actual density of the oil [kg/m3]
ρl0
Density of the oil at ambient pressure [kg/m3]
ωn
Natural frequency of the amplifier [rad/s]
INTRODUCTION
Electro-hydraulic two-stage servovalves adopt a hydraulic
amplification system serving as a pilot stage to move a main-
stage fluid-metering spool; the amplification system is usually a
flapper nozzle, a deflector jet or a jet pipe [1]. These valves are
widely used in aerospace and industrial sectors because of their
reliability and high performance in terms of step response speed
and frequency response [2]. However, they present a few
disadvantages that are still unsolved at the state of the art. One
of these disadvantages is the necessity for the pilot stage to have
a quiescent flow rate to work; although small compared to the
flow rate through the main stage, this internal leakage causes
power consumption [3]. Another disadvantage is given by the
electromagnetic torque motor assembly, which is necessary to
generate the hydraulic amplification, because it is composed of
a large number of mechanical and electrical parts that penalize
simplicity, set-up, and manufacturing costs [4]. Among these
parts, the most critical one is the flexure tube. The flexure tube
is a key component in the operation of a two-stage servovalve,
since it acts as a seal between the torque motor and the hydraulic
section of the valve. However, the presence of this component
increases the complexity of the valve, costs and duration of
manufacture because it needs to be manufactured very accurately
to ensure the stiffness required [5].
The use of piezoelectric actuators can be instrumental in
removing both the torque motor and the flexure tube from a
servovalve design, thus reducing complexity and manufacturing
costs. A piezoelectric actuator produces mechanical strain and
actuation force when a voltage is applied. A piezoelectric
actuator can provide very fast response times, but the drawbacks
are the high hysteresis (which can be as high as 20% of the
maximum displacement), the high dependence on temperature
variations, and creep. Common commercially available
piezoelectric actuators are the stack-type, amplified stacks,
rectangular benders and ring benders, as shown in Fig.1.
A stack actuator provides very high forces but low
displacement. To overcome this issue, amplification systems are
adopted to increase the displacement of a stack actuator but
producing lower forces. As an alternative to amplified stack
actuators, bimorph rectangular benders use a bending
deformation to exhibit high displacement but with very low
actuation forces. Finally, a ring bender is a flat annular disc
which deforms in a concave or convex fashion depending on the
polarity of the applied voltage. A ring bender exhibits greater
displacement than a stack actuator of the same mass, and an
increase in stiffness in comparison to similar size rectangular
bimorph type actuators.
(a) (b) (c) (d)
Figure 1. Piezo-stack actuator (a); amplified piezo-stack
actuator (b); bimorph actuator (c); ring bender actuator (d)
All these typologies of piezo-actuators have been used in the
literature as actuators for novel concepts of servovalves.
A stack actuator was used in [6] to control the flapper in a nozzle-
flapper pilot stage. In [7], four poppet valves as variable
restrictors were proportionally driven by a piezoelectric stack for
the actuation of a main stage valve.
In [8] and [9], a piezoelectric valve was directly actuated by
an amplified piezo stack actuator, with a lever element being
used to amplify the motion of the piezoelectric stack.
In [10], a commercially available stack actuator with a
flexure amplification system was used in place of a torque motor
to move a flapper in a flapper nozzle pilot stage.
In [11], [12] and [13], a piezoelectric bimorph actuator was
used to actuate the flapper in place of a torque motor in a two-
stage nozzleflapper servovalve.
Sangiah et al. [5] used a bimorph piezoelectric actuator for
the control of the first stage of a deflector jet valve.
Bertin et al. [14] developed a pilot stage of a nozzle flapper
valve by using ring bender actuators.
In [3], [15] and [16], the first stage of a two stage servovalve
was realized by employing a four-way three-position small spool
controlled by a piezoeletric ring bender. The flow through the
small spool was capable of controlling a larger spool of the main
stage valve. The servovalve body was constructed using additive
manufacture (AM). This can provide significant benefits in
3 Copyright © 2018 by ASME
weight and manufacturing labour cost, as well as providing
additional design freedom.
In [7], a pilot operated piezo valve developed by
Hagemeister is described. The valve uses two adjustable
restrictors to change the pressure at the extremities of a main
stage spool. The pilot stage uses a flapper-nozzle system with a
piezoelectric bender realising the variable orifice. Since the
benders are only capable of very small forces, the bending
elements are equipped with compensating pistons for pressure
compensation.
DEVELOPMENT OF THE VALVE CONCEPT
The architecture proposed in this paper is shown in Fig. 2
and is based on the use of two small piezo valves which have the
task of changing the pressure at the extremities of a main spool.
The piezo valves are two way two position (2/2) valves which
are both actuated by a piezoelectric actuator. Each valve is
hydraulically connected both to one of the extremities of a main
spool and to one of two fixed orifices, which in turn are
connected with the high pressure port P. The main spool can be
a typical main spool of a four way three position (4/3) main stage
valve. The sliding spool is moved directly by opening the left
piezovalve or the right piezovalve depending on the required
hydraulic connections (P-A and B-T, or P-B and A-T).This
architecture can provide the following advantages compared to a
typical two stage valve:
The two piezo valves are normally closed, therefore ideally
there is no internal flow through the small piezo valves when
they are closed, which results in a notable reduction of the
internal leakage through the system.
The implementation of small piezo valves has the potential
to improve the response time of the main stage, because of
the high dynamics of piezo-actuators, the low inertia of the
components of the small piezo valves and the possibility of
using control systems that act on both piezovalves
simultaneously.
The need for a torque motor pilot stage with its associated
disadvantages is also avoided.
Figure 2. Architecture proposed
A piezoelectric actuator exhibits mechanical strain and/or
actuation force in response to an applied voltage. For a given
input voltage, the higher the mechanical strain, the lower the
actuation force exerted by the piezo-material. Fig. 3 shows the
relationship among displacement, force and voltage. The
maximum actuation force, called the blocking force, is obtained
at null strain.
For this specific application, the piezoelectric actuator that
best fits the displacement and force requirement is the ring
bender (see Fig. 1d). Table 1 reports the characteristics of the
ring benders produced by the manufacturer Noliac [17], showing
that different models are available according to the force and
displacement required. Note that all piezoelectric actuators
driven by voltage amplifiers exhibit hysteresis, typically of up to
20%. The presence of hysteresis can cause errors in the control
of the opening degrees of the piezo valves; it can also cause the
piezo valves to remain slightly open in the case of zero voltage
applied to the ring bender. These effects will be translated to the
main stage spool, causing errors in the spool positioning. For
these reasons, it is important to take measures to compensate for
hysteresis.
Figure 3. Displacement-Force-Voltage relationship for a
piezoelectric actuator
Figure 4. Possible designs for the piezo valves
P
T
A
B
P
P
T
T
Piezo
actuator
Piezo
actuator
Force
Displacement
Voltage
Blocking
Force
Free
stroke
0
4 Copyright © 2018 by ASME
The piezo valve can be designed either as a spool valve (see
fig.4a) or as a poppet valve (see fig.4b to 4e). Among possible
designs for the poppet valve type, the architectures shown in Fig.
4b, 4c, and 4d provide pressure compensation. The architecture
shown in Fig. 4a is very similar to that shown in Fig. 4d, apart
from the fact that the latter presents a mechanical stop for the
inner spool. The mechanical stop present in a poppet valve can
help to reduce the effect of hysteresis, as the ring bender can be
provided with a negative voltage to come back into a closed
position, thus avoiding internal leakage. Instead, in the case of
no mechanical stop (fig. 4a), hysteresis can be compensated by
providing the spool with a significant overlap, which must be as
high as 20 % of the spool travel.
Table 1. Characteristics of the ring benders produced by the
manufacturer Noliac
Product
type
Length /
outer
diameter
Width /
inner
diameter
Height
Operating
voltage,
max.
Free
stroke,
max.
Blocking
force,
max.
CMBR02
20 mm
4 mm
1.25
mm
200 V
± 28
µm
16 N
CMBR03
20 mm
4 mm
1.8 mm
200 V
± 20
µm
22 N
CMBR04
30 mm
6 mm
0.7 mm
200 V
± 108
µm
11 N
CMBR05
30 mm
6 mm
1.25
mm
200 V
± 70
µm
29 N
CMBR07
40 mm
8 mm
0.7 mm
200 V
± 185
µm
13 N
CMBR08
40 mm
8 mm
1.25
mm
200 V
± 115
µm
39 N
In the following, the nozzle-flapper type architecture of Fig.
4(e) is selected as the design to be investigated, by virtue of its
highest simplicity of construction. To provide null adjustment,
the ring bender is provided with an additional spring, as shown
in Fig.5, but there is no pressure compensation (i.e. the null will
shift with supply pressure changes).
The effects of cavitation can negatively affect the
performance of the nozzles and they can be evaluated by means
of numerical investigations [18].
Figure 5. Selected architecture of piezovalve with additional
spring for null adjustment
NUMERICAL MODEL
Fig. 6 shows the implementation of the two selected nozzle-
flapper piezo valves into a two stage valve design. In order to
analyze the feasibility of the proposed architecture, a simulation
model was developed using Simscape Fluids, which provides
component libraries for simulating fluid systems including
hydraulic pumps, valves, actuators, pipelines, and heat
exchangers [19]. The dynamic system was solved by computing
its states at successive time steps over a specified time span. The
time step was taken equal to 10-6 s, in order to increase the
accuracy. In the developed model, port A and port B of the main
valve are hydraulically connected, port P is connected to a pump
providing a constant pressure pp, whereas port T is connected to
a tank having constant pressure pT. The developed model is very
accurate, since it accounts for real phenomena such as fluid
viscosity, compressibility effects and presence of air within the
hydraulic oil.
Figure 6. Schematization of the main stage spool connected
with the two piezovalves
Main spool model
Assuming that the main spool (having mass M) moves to
the right with a displacement denoted by X, the following
equations were applied to analytically study the main stage
valve. The overall actuation force acting on the main spool is to
be calculated as the product of the pressure difference between
the left and right control chambers (pl-pr) and the spool area A.
The overall actuation force is counteracted by the flow forces in
the control chambers (FflowPB, FflowAT), the damping force (C)
and the inertia force ():
   (1)
The damping coefficient C allows the fluid viscosity to be
taken into account. The flow rates qs,l and qs,r flowing through the
external chambers of the main spool are calculated as follows:
T
B
p
p
Spool
pl
P
T
A
pr
q
s,l
qs,r
p
p
q
o,l
q
o,r
Q
Q
T
T
P
X
5 Copyright © 2018 by ASME
 

 
 
 (2a)
 

 
 
 (2b)
where Vleft and Vright are the oil volumes present in the left and
right external chambers of the main spool, which can be
calculated as a function of the dead volume, Vdead, the spool
position, X, and the spool surface A:
    (3a)
   (3b)
In equations 2, ρl0 is the oil density at ambient pressure and
ρl is the actual density of the oil, calculated as follows:



(4)
where accounts for the quantity of air present in the oil, E0 is
the bulk modulus at atmospheric conditions, while and are
the gauge pressure of the oil and the atmospheric pressure; γ is
the ratio of the specific heat at constant pressure to the specific
heat at constant volume, with ρg0 denoting the gas density at
atmospheric pressure.
Each metering chamber of the main spool is approximated
by an orifice with variable area, applying the following
equations:
  

 (5)
  (6)



(7)
where Q is the flow rate through the main spool, is the
discharge coefficient,  is the pressure drop across the metering
section, is the restricted area, b is the width of the slots;
denotes the minimum pressure for turbulent flow, is the
critical Reynolds number and is the fluid kinematic viscosity,
 is the leakage area, which can be calculated as the product
of the spool perimeter and the clearance between the spool and
the bushing.
The flow forces must be taken into account for a precise
analysis of the spool dynamics; their value is usually very high
and cannot be neglected. Because of this, numerical and
experimental investigations are present in the literature to reduce
the flow forces [20] [21]. The flow force in each metering section
was calculated by using the equation:


where θ is the flow angle, which is calculated by using the
relation [22]:    
) (9)
The position of the main spool is measured by an ideal
translational motion sensor, which does not account for inertia,
friction, delays and energy consumption.
Piezovalve model
Two hydraulic chambers are considered in order to evaluate
the effects of the fluid compressibility in the lines connecting the
piezovalves with the main stage:

(10)



 (11)
where is the geometrical volume of the chamber (equal to the
product of an internal diameter Dint and an overall internal length
Lint),  is the oil volume in the chamber at the gauge
pressure p, with denoting the flow rate through the chamber.
If the pressure in the chamber reaches the cavitation limit, the
above equations are enhanced by representing the fluid in the
chamber as a mixture of liquid and a small amount of entrained,
non-dissolved gas, and by calculating the mixture bulk modulus
as:
  

 
 (12)
In this way it is possible to introduce an approximate model of
cavitation, which takes place in a chamber if the pressure drops
below the fluid vapor saturation level. In fact, the bulk modulus
of a mixture decreases at p→pa, thus considerably slowing down
further pressure change. At high pressure, p>>pa, a small
amount of non-dissolved gas has practically no effect on the
system behavior.
Each piezovalve is simulated as an orifice with variable area
through the following equations:
 

 (13)
  (14)



(15)
where CD,p denotes the discharge coefficient of the piezo valves,
is the flow rate through the piezovalve, d is the diameter of
the orifice, x is the ring bender position, is the fluid kinematic
viscosity,  is the pressure drop across the piezovalve.
The position x of the ring bender is determined according to the
actuation force and the resistant forces acting on the ring bender
(having the mass denoted by m), as follows:
        (16)
where  denotes the blocking force exerted by the ring bender
having a stiffness denoted by . The additional spring has
6 Copyright © 2018 by ASME
stiffness denoted by and pre-compression denoted by x0; is
the damping coefficient of the ring bender which allows taking
into account the effects of viscosity upon the piezo valve
performance. Considering that the flow exiting the piezo valve
can be assumed radial, the flow forces acting on the left and on
the right ring bender (Fflow,pl and Fflow,pr) are calculated as follows:
 
(17)
  
(18)
The ring bender stroke is limited by two stops that restrict
its motion between upper and lower bounds. Each stop is
represented as a spring and damper. A force Fstop acts on the ring
bender when the maximum or minimum displacement is
reached:
  
 
for x  (19a)
  
 for
x (19b)
where  and  are the spring stiffness and damping of the
stop, with  and  denoting the maximum and minimum
displacement of the ring bender.
The flow through the fixed orifices is calculated as follows:
  

 (20a)
 

 (20b)



(21)
where CD,0 denotes the discharge coefficient of the two fixed
orifices, and  denotes the orifice area.
Piezoelectric hysteresis was considered by implementing
the Bouc-Wen hysteresis model, described and used in [15]. The
Bouc-Wen model is represented by equation 22, where is the
hysteresis nonlinear term:

  
 
 
 (22)
where α, β and δ are tuning parameters used to match the
hysteresis model to experimental data (the values from [15] are
used), and Vamp is the output voltage from the amplifier. The
hysteresis non-linear term allows the blocking force to be
expressed as a function of the output voltage from the amplifier
as follows: = (23)
where Kd,v is the ring bender maximum blocking force divided
by the maximum operating voltage.
The amplifier was simulated by using a second order
transfer function:

 (24)
where Vc is the control voltage that is supplied to the amplifier,
and is the gain of the amplifier. In addition, to model the
current limit, the rate of change of voltage is limited according
to the following equation:

 
 (25)
where Caprb is the capacitance of the piezoelectric ring bender.
Controller model
The control voltages to the left and to the right ring bender,
Vc,l and Vc,r, were obtained as follows:
=  (26a)
=  (26b)
where ul and ur are the control output, while off is an offset
introduced to cope with hysteresis; ul and ur are obtained through
a Propotional-Integral (PI) controller. The control output from
the PI controller is calculated through a proportional-integral
action employing a clamping anti-wind up method:

for e(t)0 (27a)
 
for e(t)0 (27b)
where e(t) is the error between the actual spool position and the
demand.
PARAMETERS VALUES
The blocking force can be adjusted by changing the voltage
applied to the ring bender, from -100 Volt to 100 Volt, and the
range values for the blocking force depend on the selected ring
bender. In this analysis, the ring bender CMBR08 (see Table 1),
having a maximum blocking force of ±39 N, is used.
Simulations were performed using the values for the operating
parameters as reported in Table 2.
The parameters of the main stage were assumed in order to
consider the simulation of a medium-size valve having a main
stage spool characterized by a diameter of 7 mm and mass of 20
g. The width of the slots of the bushing were assumed equal to
10 mm, namely about one half of the spool perimeter. A very
small value for the leakage area was considered, since in
servovalves the clearance between spool and bushing is usually
very small (lower than 3µm [1]). The dead volume, Vdead, was
7 Copyright © 2018 by ASME
calculated as the product of the main spool lateral surface and
the double of the maximum spool displacement (assumed equal
to 1 mm).
Table 2. Parameters assumed for the simulations
Component
Parameter
Symbol
Value
Main Valve
Main spool lateral surface
38.5 mm2
Main spool diameter
D
7 mm
Discharge coefficient
CD
0.7
Width of the slots
b
10 mm
Leakage area

1e-9 m2
Main spool mass
M
20 g
Dead volume
Vdead
77 mm2
Damping coefficient
C
10 Ns/m
Piezo valves
Diameter of the orifice
d
1 mm
Discharge coefficient
CD,p
0.7
Maximum displacement of
the ring bender
xmax
0.115 mm
Minimum displacement of
the ring bender
xmin
0 mm
Mass
m
6 g
Damping coefficient
Crb
412 Ns/m
Ring bender stiffness
krb
340000 N/m
Additional spring stiffness
Ks
340000 N/m
Pre-compression
additional spring
X0
0.06 mm
Stop damping coefficient

150 Ns/m
Stop stiffness

106 N/m
Fixed orifices
Restricted area
Ar,0
0.08 mm2
Discharge coefficient
CD,0
0.7
Hydraulic
chambers 1 & 2
Diameter
Dint
3 mm
Length
Lint
40 mm
PI parameters
Proportional gain
12
Integral gain
700
Offset
off
1 V
Saturation limits
ul,max
ur,max
6V
Pump
Pressure
pP
210 bar
Reservoir
Pressure
pT
1 bar
Oil
Density
ρl0
966 kg/m3
Relative gas content
ε
0.005
Amplifier
Natural frequency
ωn
12000 rad/s
Damping factor
ξ
0.8
Maximum current
Imax
1A
Ring bender capacitance

2x1740 nF
Gain of the amplifier
20
The dimensions of the additional chambers (Dint, Lint) were
calculated by estimating a possible hydraulic volume comprised
between the piezovalves and the fixed orifices (note that the
overall hydraulic volume at the left and right of the main spool
is given by the sum of the dead volume and the volume of the
additional chamber). The parameters of the ring bender and
amplifier are the same as those reported in a previous paper [3].
With regard to the stiffness of the additional spring, it was
assumed equal to that of the ring bender. The pre-compression of
the additional spring was calculated by equating the maximum
flow force acting on the ring bender with the pre-compression
force. The resulting pre-compression is then increased by 20%
to account for the hysteresis of the ring bender:

 (28)
In this way, the additional spring is capable of keeping the
piezovalve closed in case of a null voltage applied to the ring
bender.
The results of the simulations, analyzed in the following
section, have been obtained for the fixed parameters reported in
Table 2. Future investigations will deal with the numerical
optimization of the main parameters, such as the area of the fixed
orifices, the diameter of the orifice of the piezo valves and the
PID parameters, in order to obtain the maximum response speed
of the valve.
RESULTS
The effectiveness of the proposed valve architecture has
been studied using step responses for the position of the main
spool. Fig. 7a, 7b, 7c and 7d show the time history of four step
tests, for step sizes of 0.1 mm, 0.4mm, 0.7mm and 1 mm,
respectively.
Similarly, Fig. 8a, Fig 8b, Fig. 8c and Fig. 8d show the time
history of the ring bender positions for steps of 0.1mm, 0.4mm,
0.7mm, and 1mm respectively. Table 3 reports the flow rate and
response times achieved for the four step tests.
The response rate is given by the rising time interval
required to reach 90% of the imposed set point. The valve
responds rapidly, with the response time increasing with
increasing step amplitude. Very small overshoots are seen.To
reach the desired target position, the right piezovalve is opened,
while the left piezovalve is maintained closed. The opposite
happens in the case of a negative step (i.e. the main spool moving
from the right to the left). The right piezovalve reaches its
maximum opening for the steps of 0.4mm, 0.7mm and 1mm, and
just remains open longer for the larger steps. When the set point
has been reached, the right piezo valve is maintained slightly
opened in order to have a pressure difference at the spool
extremities that is capable of counteracting the flow forces. It is
noteworthy that, thanks to the proposed control strategy, both
ring benders come back to the zero position when X=0mm in
spite of hysteresis. This is a notable result, as this ensures that no
internal leakage is present in the small piezo valves.
8 Copyright © 2018 by ASME
(a)
(b)
(c)
(d)
Figure 7. Main spool position vs time; (a): step=0.1mm,
(b): step=0.4mm, (c): step=0.7mm, (d): step=1mm.
(a)
(b)
(c)
(d)
Figure 8. Ring bender position vs time;
(a): step=0.1mm, (b): step=0.4mm, (c): step=0.7mm, (d):
step=1mm.
Table 3. Results of the step tests
Step test
X=0.1mm
X=0.4mm
X=0.7mm
X=1mm
Flow rate
6 l/min
25 l/min
43 l/min
62 l/min
Time to reach 90% of
the output
1.11 ms
2.73 ms
4.45 ms
6.2ms
9 Copyright © 2018 by ASME
The frequency analysis has been performed using a step-sine
signal. Figure 9 shows the time history of the main spool position
predicted for an amplitude of X=1mm and an input frequency of
50 Hz (Fig.9a) and 100 Hz (Fig. 9b). Figure 10 shows the
corresponding ring bender positions.
(a)
(b)
Figure 9. Frequency response: main spool position vs time; (a):
amplitude=1mm, frequency=50 Hz; (b): amplitude=1mm,
frequency=100 Hz.
(a)
(b)
Figure 10. Frequency response: ring bender position vs time;
(a): amplitude=1mm, frequency=50 Hz; (b): amplitude=1mm,
frequency=100 Hz.
(a)
(b)
Figure 11. Bode Plot: magnitude diagram (a) and phase
diagram (b)
As visible in both graphs of Fig.9a and 9b, the response to
a high frequency sinusoidal input is more like a triangle wave,
with an amplitude that decreases with increasing frequency. This
is because at these amplitudes and frequencies, the ring bender
valves are saturating, i.e. reaching their maximum opening.
The Bode Plot has also been plotted for an amplitude of 1
mm, 0.5 mm and 0.1 mm, as shown in Fig. 11.
When the amplitude is 1 mm, the phase shift is around -90
deg for a frequency of 100 Hz; when the amplitude is 0.5mm, the
phase shift is about -90 deg for 150 Hz frequency; when the
amplitude is 0.1mm, the phase shift is about -90 deg for 275 Hz
frequency.
These results confirm the effectiveness of the proposed
valve architecture: the main spool displacement reaches the set
point in a very short time and with negligible overshoot. The
dynamic characteristics of the system are also very good, with
the frequency response being comparable to that of conventional
two stage servovalves. The main advantage is the lack of internal
leakage in the small piezo valves, which will result in a more
energy efficiency for the system compared to conventional two
stage servovalves.
-8
-6
-4
-2
0
2
4
6
8
10 100
Amplitude (db)
Frequency-Hertz
500
Magnitude plot
1 mm amplitude 0.5 mm amplitude
0.1 mm amplitude
-120
-100
-80
-60
-40
-20
0
10 100
Phase-degrees
Frequency-Hertz
500
Phase plot
1 mm amplitude 0.5 mm amplitude
0.1 mm amplitude
10 Copyright © 2018 by ASME
CONCLUSIONS
This paper has presented a novel architecture for
servovalves that is based on the use of two small piezovalves
capable of changing the pressure at the extremities of a main
stage spool. Each piezovalve is actuated by a piezoelectric ring
bender, which can provide a good level of force and
displacement. The main advantage of this architecture is the
reduction of internal leakage, which results in power savings. A
SimScape model was developed to study this architecture. Both
step tests and frequency analysis have been carried out. It has
been shown that the response time is very fast, with negligible
overshoots being registered. When the spool position amplitude
is 1 mm, which is its maximum design value, the phase shift is
around -90 deg for a frequency of 100 Hz. Also, when the
amplitude is 50%, the phase shift is about -90 deg at 150 Hz
frequency.
REFERENCES
[1] H. Merritt, Hydraulic Control System, John Wiley &
Sons Inc, 1967.
[2] J. F. Blackburn, G. Reethof, and J. L. Shearer, Fluid
power control, The Mit press, 1960.
[3] J. Persson, A. R. Plummer, C. R. Bowen, and I. Brooks,
“Design and Modelling of a Novel Servovalve Actuated
by a Piezoelectric Ring Bender,” in ASME/BATH 2015
Symposium on Fluid Power and Motion Control, 2015.
[4] A. Plummer, “Electrohydraulic servovalves past,
present, and future”, Proc. 10th Int. Fluid Power Conf.,
pp. 405424, 2016.
[5] D. K. Sangiah, A. R. Plummer, C. R. Bowen, and P.
Guerrier, “A novel piezohydraulic aerospace servovalve.
Part 1: Design and modelling”, Proc. Inst. Mech. Eng.
Part I J. Syst. Control Eng., vol. 227, no. 4, pp. 371389,
2013.
[6] Y. B. Bang, C. S. Joo, K. I. Lee, J. W. Hur, and W. K.
Lim, “Development of a two-stage high speed
electrohydraulic servovalve systems using stack-type
piezoelectric elements”, in IEEE/ASME International
Conference on Advanced Intelligent Mechatronics, AIM,
2003, vol. 1, pp. 131136.
[7] F. Bauer and M. Reichert, “The use of piezo-actuators
for high dynamic servovalves”, Olhydraulik Pneum.,
vol. 49, no. 6, 2005.
[8] J. E. Lindler and E. H. Anderson, “Piezoelectric Direct
Drive Servovalve", Industrial and Commercial
Applications of Smart Structures Technologies, San
Diego, March 2002.
[9] J. Jeon, C. Han, Y. M. Han, and S. B. Choi, “A new type
of a direct-drive valve system driven by a piezostack
actuator and sliding spool,” Smart Mater. Struct., vol. 23,
no. 7, 2014.
[10] S. Karunanidhi and M. Singaperumal, “Mathematical
modelling and experimental characterization of a high
dynamic servo valve integrated with piezoelectric
actuator”, Proc. Inst. Mech. Eng. Part I J. Syst. Control
Eng., vol. 224, no. 4, pp. 419435, 2010.
[11] A. Milecki, “Modelling and investigations of
electrohydraulic servo valve with piezo element,” Masz.
i Autom. Technol., vol. 26, no. 2, pp. 177184, 2006.
[12] L. Zhu, E. Shiju, X. Zhu, and C. Gao, “Development of
Hydroelectric Servo-Valve Based on Piezoelectric
Elements,” in 2010 Int. Conf. Mech. Autom. Control
Eng. MACE2010, 2010, pp. 3330 3333.
[13] G.-M. Cheng, P. Li, Z.-G. Yang, S.-J. E, and J.-F. Liu,
“Double-nozzle piezoelectric servovalve,” Guangxue
Jingmi Gongcheng/Optics Precis. Eng., vol. 13, no. 3,
2005.
[14] M. J. F. Bertin, C. R. Bowen, A. R. Plummer, and D. N.
Johnston, “An Investigation of Piezoelectric Ring
Benders and Their Potential for Actuating Servo
Valves,” in Proceedings of the Bath/ASME Symposium
on Fluid Power and Motion Control, 2014, p. 6.
[15] J. Persson, A. R. Plummer, C. R. Bowen, and P. L.
Elliott, “Dynamic Modelling and Performance of a Two
Stage Piezoelectric Servovalve,” in 9th FPNI Ph. D.
Symposium on Fluid Power. American Society of
Mechanical Engineers, 2016.
[16] J. Persson, A. Plummer, C. Bowen, and P. Elliott, "Non-
linear Control of a Piezoelectric Two Stage Servovalve",
in Proceedings of 15th Scandinavian International
Conference on Fluid Power, June 7-9, 2017, Linköping,
Sweden.
[17] Noliac, http://www.noliac.com/products/actuators/plate-
stacks/. Accessed September 2017.
[18] Amirante R, Distaso E, Tamburrano P. Experimental and
numerical analysis of cavitation in hydraulic
proportional directional valves. Energy Convers Manag
2014;87:20819.
[19] Matlab & Simulink. SimscapeTM user’s guide R2018a.
Mathworks.
[20] Amirante R, Catalano LA, Poloni C, Tamburrano P.
Fluid-dynamic design optimization of hydraulic
proportional directional valves. Eng Optim 2014;46(10),
1295-1314.
[21] Amirante R, Distaso E, Tamburrano P. Sliding spool
design for reducing the actuation forces in direct
operated proportional directional valves: Experimental
validation. Energy Convers Manag 2016;119:399410.
[22] Simscape Fluids documentation.
https://www.mathworks.com/help/physmod/hydro/ref/s
poolorificehydraulicforce.html.
... The feasibility of the proposed solution was already demonstrated in reference [28], employing a simulation model of the full valve concept. The simulations showed that this valve architecture has a high potential in terms of response speed, in addition to minimising the internal leakage. ...
... The good results achieved in [28] prompted the authors of the present work to construct a prototype of one of the two 2/2 normally closed piezo-valves representing the novel pilot stage concept. The prototype was tested in a hydraulic test rig located at the Fluid Power Laboratory of the Centre of Power Transmission and Motion Control (PTMC) of the University of Bath. ...
... where E0 is the pure liquid bulk modulus, is the gas-specific heat ratio ( = 1.4), is the relative gas content at atmospheric pressure and is the atmospheric pressure [28,30]. The hydraulic part of the piezo-valve is simulated using the orifice Equation (1), in which the orifice area is evaluated as , where d is the diameter of the nozzle (10) and is the displacement of the ring bender with respect to the nozzle tip. ...
Article
Full-text available
Electrohydraulic servovalves are widely used for precise motion control in aerospace and other industries due to their high accuracy and speed of response. However, commercial two-stage servovalves have several undesirable characteristics, such as the power consumption caused by the quiescent flow (internal leakage) in the pilot stage, and the complexity and high number of parts of the torque motor assembly, which affect the cost and the speed of manufacture. The solution to these problems can help to reduce costs, weight and power consumption, and enhance the reliability and reproducibility as well as the performance of these valves. For these reasons, this paper proposes a novel configuration for the pilot stage: it is composed of two normally closed two-way two-position (2/2) valves actuated by two piezo-electric ring benders; the opening and closing of the two piezo-valves can generate a differential pressure to be used to control the displacement of the main spool. In this way, there is negligible quiescent flow when the main stage is at rest; in addition, the torque motor and all its components are removed. To assess the performance of this novel pilot stage concept, a prototype of the piezo-valve has been constructed and tested. The experimental results indicate that the response speed of the new piezo-valve is very high. Furthermore, a numerical model is employed to show that, by adjusting specific parameters, the performance of the piezo-valve can be further improved, so that the valve can be fully opened or closed in less than 5 ms.
... The proposed architecture was preliminarily studied in [26], in which a simulation model was developed to predict the performance of this architecture for a medium size valve, showing that the response time can be very short, with an interval time of about 7 ms being predicted to change the main spool position from 0% to 90% of the maximum opening (X = 1 mm). That simulation work assumed some estimated values for the damping factor of the ring bender and for the amplifier parameters, which were taken from other studies [21]. ...
... That simulation work assumed some estimated values for the damping factor of the ring bender and for the amplifier parameters, which were taken from other studies [21]. The preliminary results of [26], being promising, encouraged us to construct a piezo-valve prototype, which was tested in a test rig, reproducing only the pilot stage of the architecture shown in Figure 2, including the fixed restriction Ar,0 and the chamber of volume V0 between the fixed restriction and the piezo-valve [27]. This research activity was described thoroughly in [27], which must be regarded as the first part of the present study. ...
... The valve prototype and the test rig, which, for simplicity, were constructed with non-optimized parameters (for example, the volume V0 and the mass m0 of the moving parts were quite large), were instrumental in validating a numerical model for the pilot stage [27]. The damping factor of the ring bender, the amplifier parameters, and the hysteresis parameters were evaluated with good accuracy in [27], thus obtaining more precise values compared to the previous study [26]. Despite using large values for V0, m0, and for the maximum opening of the ring bender (xmax), the experimental results confirmed that the piezo-valve has high potential in terms of step response speed, since the interval time required for the displacement of the ring bender to change from 0% to 90% of its final value considered in the tests (xmax = 0.15 mm) was less than 5 ms [27]. ...
Article
Full-text available
In part I of this study, we experimentally and numerically investigated the pilot stage of a novel two-stage servovalve architecture. The novelty of the proposed configuration is the torque motor being removed and replaced with two small two-way two-position (2/2) valves actuated by piezoelectric ring benders, which can effectively control the opening degree of a main spool valve. With this novel architecture, the typical drawbacks of two-stage servovalves can be overcome, such as the high complexity of the torque motor and the high internal leakage in the pilot stage when the main valve is at rest in the neutral position (null). The low complexity and the negligible internal leakage of the piezo-valves are accompanied by the high response speed typical of piezoelectric actuators. The valve assessment is completed in the present study, since the entire valve architecture (main stage + pilot stage) is investigated. In particular, a simplified numerical model is developed to provide a design tool that allows, for a given main stage spool, the values of the geometrical parameters of the pilot stage to be chosen along with the characteristics of the ring bender. This design procedure is applied to a 7 mm diameter main spool; afterward, a detailed numerical model of the entire valve, solved by SimScape Fluids software, is employed to demonstrate that the response of the main stage valve is very rapid while ensuring negligible internal leakage through the piezo-valves when the main stage is closed (resulting in lower power consumption). For this reason, the proposed valve can be regarded as a “clean” component for energy conversion, having lower energy consumption than commercially available servovalves.
... In [4,58,59], a novel pilot stage configuration composed of two normally closed twoway two-position (2/2) valves actuated by two piezoelectric ring benders was developed, as shown in Figure 25. Again, an LVDT was used to achieve closed-loop control. ...
... The results are very satisfactory in terms of performance levels. A novel configuration has also been studied: making use of both two-ring benders and a different configuration for the hydraulic bridge, this novel valve can eliminate internal leakage in the pilot stage while obtaining a good step response and bandwidth [4,58,59]. ...
Article
Full-text available
This paper is a thorough review of innovative architectures of electro-hydraulic servovalves that exploit actuation systems based on piezo-electric materials. The use of commercially available piezo-electric actuators, namely, piezo stacks, amplified piezo stacks, rectangular benders, and ring benders, is very promising for the actuation of the main stages and of the pilot stages of servovalves given the fast response and low weight of piezoelectric materials. The use of these actuators can also allow novel designs to be developed, thus helping manufacturers to overcome the typical drawbacks of commercial servovalves, such as the high complexity and the high internal leakage of the pilot stages of two-stage servovalves as well as the large size and weight of direct-drive servovalves. First, the piezoelectric actuators that can be used for driving servovalves are presented in the paper, and their characteristics are thoroughly discussed. The main novel architectures present in the literature are then explained and compared with the commercial ones, and their performance parameters are discussed to draw conclusions on the prospect that some of these architectures can be used by manufacturers as future designs.
... Concerning the former, current research studies have investigated the effects of geometrical imperfections, due to wear and manufacturing processes, showing that such imperfections have a huge effect on the overall leakage [3][4][5][6]. Concerning the latter, novel configurations for the pilot stage have been proposed and are currently being studied taking advantage of piezo-electric actuators, showing that the internal leakage of the pilot stage can be reduced using such configurations [7][8][9]. ...
Article
Full-text available
Despite being widely used in several applications, commercially available spool valves, both servovalves and proportional valves, are inefficient components because they cause high power consumption due to the large pressure drops across the metering orifices. A recent research field aims at substituting spool valves with on/off valves having high switching frequency (changing state between open and closed in a few milliseconds or less) and producing low pressure drops, in order to make the so-called digital hydraulics possible. In spite of the advantages that it could provide, digital hydraulics does not have significative industrial applications yet, because of the difficulty in manufacturing such high frequency on/off valves. Hence, this paper performs a feasibility study of an on/off poppet-type valve actuated directly by a commercially available ring stack, which is a multilayer piezo-actuator capable of generating very high actuation forces needed for this application. Modulation of the average flow can be achieved by changing the duty cycle of the pulse width modulation (PWM) signal driving the piezostack. An inertance tube could also be used to smooth flow pulsation. The simulations obtained using a detailed Simulink model show that high switching frequency and very effective flow modulation can be obtained with this valve architecture along with low pressure drops and high flow rates, thus making it potentially suitable for digital hydraulics. The disadvantages of this single stage architecture are the large dimensions of the piezo stacks, and the high current generated because of both the high capacitance of the piezo stack and the high frequency switching. However, large-scale production of these components could help to reduce the costs, and the simulations show that limiting the maximum current to 10 A still provides good regulation.
... In [4,58,59], a novel pilot stage configuration composed of two normally closed twoway two-position (2/2) valves actuated by two piezo-electric ring benders was developed, as shown in Fig. 25. Again, an LVDT was used to achieve closed loop control. ...
Preprint
Full-text available
This paper is a thorough review of innovative architectures of electro-hydraulic servovalves that exploit actuation systems based on piezo-electric materials. The use of commercially available piezo-electric actuators, namely, piezo-stacks, amplified piezo-stacks, rectangular benders and ring benders, is very promising for the actuation of the main stages and of the pilot stages of servovalves, given the fast response and low weight of piezoelectric materials. The use of these actuators can also allow novel designs to be developed, thus helping manufacturers to overcome the typical drawbacks of commercial servovalves, such as the high complexity and the high internal leakage of the pilot stages of two stage servovalves, as well as the large size and weight of direct drive servovalves. Firstly, the piezoelectric actuators that can be used for driving servovalves are presented in the paper and their characteristics are thoroughly discussed. Then, the main novel architectures present in the literature are explained and compared with the commercial ones, and their performance parameters are discussed to draw conclusions on the prospect that some of these architectures can be used by manufacturers as future designs.
... In addition, lightly damped flapper resonance makes the valve main spool position very sensitive to external noise. To date, research studies in the scientific literature have mainly been focused on reduction of complexity of servovalves; in particular, a promising research field aims to replace the electromagnetic torque motor assembly with piezoelectric actuators, thus reducing complexity and manufacturing costs [7][8][9][10][11][12][13]. However, the scientific literature lacks proposals to reduce the undesired effects of external noise upon a servovalve. ...
Conference Paper
A fundamental component of two-stage servovalves is the flexure tube, since it both constitutes a low friction pivot for the inner flapper and allows the torque motor to be separated from the hydraulic fluid, thus avoiding contamination particles being trapped inside the torque motor. The inertia of the torque motor armature interacting with the flexure tube stiffness gives lightly damped resonances, which may lead to fatigue failure due to excessive bending under vibration, as well as limiting the position control bandwidth of the main spool. This effect is counteracted by the film of liquid interposed between the flapper and the flexure tube, which is “squeezed” during the flapper motion providing damping. However, the underlying physics of the damping mechanism caused by the squeeze film inside the flapper-flexure tube system is not well-understood, and to date the scientific literature has lacked analyses and investigations aimed at providing insights into this phenomenon. Because of this, the aim of this paper is to develop a reliable CFD model which can help to understand where and how the damping forces are generated during the flapper motion because of the squeeze film. The developed model could be used in further investigations, aimed, for example, at studying the effects of fluid properties and geometric parameters upon the damping factor, in order to achieve more effective designs which can enhance the damping factor in the flapper-flexure tube system of new generation servovalves. This work has been carried out as a collaboration between the University of Bath and the Polytechnic University of Bari, and Moog Controls ltd (Tewkesbury, UK), a world leading manufacturer of servovalves.
... The pilot stage is always actuated by a torque motor, while the hydraulic amplification is obtained through a double-flapper nozzle, a deflector jet or a jet pipe. There can be either a mechanical feedback or an electrical one depending on the application (mechanical feedbacks are more used in aircraft) [1][2][3][4]. ...
Conference Paper
The internal leakage in two stage servovalves causes unwanted power consumption; it is the sum of two contributions: the internal leakage in the main stage and the internal leakage in the pilot stage. While the latter can be assumed almost constant regardless of the spool position, the former is maximum at null and decreases with increasing opening degree of a given valve. Because of this, the power consumption is significant when a valve is at rest, namely, when it is not modulating flow. Despite being a very important feature of these valves, the internal leakage occurring in the main stage around null and its associated issues are not properly addressed in the scientific literature. Because of this, this paper aims at providing a deep investigation into this phenomenon. In particular, it will be discussed how it can be studied using analytical equations. In addition, a CFD analysis is carried out in this paper in order to obtain a simple CFD model that has general validity and that can be used to predict the internal leakage around null in the main stage. The developed model can be easily reproduced by manufacturers, and it can be used to understand the effects of geometrical imperfections and tolerances as well as fluid properties upon the internal leakage around null. The present paper has been realized in collaboration with Moog controls ltd, a world leading manufacturer of servovalves.
... However, they present a few disadvantages that are still unresolved. One of these disadvantages is the necessity for the pilot stage to have a quiescent flow rate to work (Hunt & Vaughan, 1996); although small compared to the maximum flow rate through the main stage, this internal leakage can consume a considerable proportion of the input power if the valve is at rest for long periods (Tamburrano et al., 2018a). The electromagnetic torque motor assembly is also a major issue with these valves because it is composed of many sensitive mechanical and electrical parts that penalize simplicity, set-up and manufacturing costs. ...
Article
Full-text available
This paper provides a review of the state of the art of electro-hydraulic servovalves, which are widely used valves in industrial applications and aerospace, being key components for closed loop electrohydraulic motion control systems. The paper discusses their operating principles and the analytical models used to study these valves. Commercially available units are also analysed in detail, reporting the performance levels achieved by current servovalves in addition to discussing their advantages and drawbacks. A detailed analysis of research that investigates these valves via computational fluid dynamic analysis is also provided. Research studies on novel control systems and novel configurations based on the use of smart materials, which aim to improve performance or reduce cost, are also analysed in detail.
Article
Full-text available
Enhancing the dynamic bandwidth of flow control valves based on piezoelectric actuators has attracted much attention in the field of precision fluid control. This paper reports a high-speed piezoelectric direct-drive flow control valve with enhanced flow rate by introducing a new push-pull complementary compliant mechanism. An improved semi-rhombus compliant amplifying mechanism is designed to amplify the micro stroke of piezo-stacks with enhanced resonance frequency. To facilitate design, the dynamic stiffness model of the compliant amplifying mechanism is derived and the structural parameters are optimized using the Pareto multi-objective optimization strategy. In addition, a polyvinylidene fluoride (PVDF) based high-response displacement sensor with an improved differential charge amplifier circuit is developed and integrated into the valve to measure the spool displacement real-timely. A proof-of-concept prototype is fabricated and the flow characteristics are experimentally tested in a closed-loop control with the PVDF sensor. The flow rate and dynamic bandwidth of the presented piezo-valve are evidently enhanced, reaching the dynamic bandwidth in excess of 920 Hz (-3 dB) and the flow rate of ±6 L/min (corresponding stroke is 0.2 mm) under the supply pressure of 70 bar.
Article
To acquire a linear load pressure with the potential of energy-saving compared with conventional throttling hydraulic systems, a switched hydraulic circuit (referred to as a flow-dependent inertia hydraulic converter) featuring a flow-dependent inertial effect was proposed. In this device, the flowing fluid modulated by an equivalent fast switching valve allows the fluid inertance at the connection pipe and rotation inertia incorporated in the flywheel to be combined. The theoretical and simulation models of the flow-dependent inertia hydraulic converter were established, and then the flow-dependent inertia was discussed and quantified. The predictions of the load pressure, flywheel speed and system efficiency were confirmed by experimental measurements. Results indicated that the theoretical and simulation predictions match well with the experiments for this configuration. The pressure wave propagation effect within the connection pipe is responsible for the variation in the transient load pressure and flywheel speed that is initially modulated by the rotation inertia theoretically, which seriously questions the predecessor's hypothesis of using a compressible volume to model the upstream connection pipe. The parabolic mean suction flow induced by the flow-dependent inertia draw our attention to the available supply flowrate and duty cycle. Finally, the throttling loss and system efficiency that varies with the desired characteristics is superior to that of conventional throttling hydraulic systems theoretically, and an acceptable system efficiency was obtained under the premise of ignoring friction experimentally.
Article
Full-text available
This article proposes an effective methodology for the fluid-dynamic design optimization of the sliding spool of a hydraulic proportional directional valve: the goal is the minimization of the flow force at a prescribed flow rate, so as to reduce the required opening force while keeping the operation features unchanged. A full three-dimensional model of the flow field within the valve is employed to accurately predict the flow force acting on the spool. A theoretical analysis, based on both the axial momentum equation and flow simulations, is conducted to define the design parameters, which need to be properly selected in order to reduce the flow force without significantly affecting the flow rate. A genetic algorithm, coupled with a computational fluid dynamics flow solver, is employed to minimize the flow force acting on the valve spool at the maximum opening. A comparison with a typical single-objective optimization algorithm is performed to evaluate performance and effectiveness of the employed genetic algorithm. The optimized spool develops a maximum flow force which is smaller than that produced by the commercially available valve, mainly due to some major modifications occurring in the discharge section. Reducing the flow force and thus the electromagnetic force exerted by the solenoid actuators allows the operational range of direct (single-stage) driven valves to be enlarged.
Conference Paper
This paper describes the design, simulation and testing of a piezoelectric spool valve. An actuator has been connected to the valve and tested under closed loop control. A mathematical model of the valve was produced and a prototype of the valve was tested. The mathematical model is validated against the experimental data. Step and frequency responses for both the valve and actuator are presented. It was found that displacement of the hydraulic fluid by the ring bender had an impact on the valve performance. To reduce the effect of the piezoelectric hysteresis, closed loop spool position control was evaluated. A noticeable difference can be observed between open loop and closed loop performance. Copyright © 2015 by ASME Country-Specific Mortality and Growth Failure in Infancy and Yound Children and Association With Material Stature Use interactive graphics and maps to view and sort country-specific infant and early dhildhood mortality and growth failure data and their association with maternal
Conference Paper
This paper describes the design, simulation and testing of a piezoelectrically-actuated two stage spool valve. The first stage is a small spool controlled by a piezoelectric ring bender and the second stage is a closed loop position controlled main spool. A dynamic mathematical model of the valve has been derived. Step response tests for both the first stage and second stage are presented, and experimental second stage frequency response results are shown. The mathematical models for the amplifier, first stage and second stage spool behaviour are validated against experimental data. The second stage spool behaviour is found to be very sensitive to the first stage flow characteristics. In addition, the displacement of the hydraulic fluid by the piezoelectric ring bender, from one side of the piezoelectric ring bender to the other, had a significant impact on the valve characteristics. Copyright © 2016 by ASME Country-Specific Mortality and Growth Failure in Infancy and Yound Children and Association With Material Stature Use interactive graphics and maps to view and sort country-specific infant and early dhildhood mortality and growth failure data and their association with maternal
Article
The high demands on precision and dynamics of hydraulic drives result in new requirements for valves. The performance of valves depends primarily on the dynamics of electromechanical actuators, which drive the valve spool directly or indirectly. The use of fast and stiff piezoelectric actuators as valve drives offers advantages for valve characteristics concerning disturbance and reference input at high dynamics.
Article
To improve the static and dynamic characteristics of traditional electrohydraulic servovalves, a new type electrohydraulic servovalve was designed by using piezoelectric actuator based on bimorph instead of the moment motor of conventional servovalves. The results show that the output distance of designed piezoelectric bimorph actuator is 82.5 μm, fluctuation in full scale is less than 0.7 μm, and resonance frequency is 1.2 kHz. It can meet the requests of servovalve. And the designed servovalve owns perfect linearity when the frequency is less than 100 Hz. The lacks of traditional electrohydraulic servovalves in bandwidth and anti-electromagnetic ability have been overcome.
Article
Servovalves are compact, accurate and high-bandwidth modulating valves widely used in aerospace, defence, industrial and marine applications. However, manufacturing costs are high due to high part count and tight tolerances required, particularly in the first stage of the valve, and due to manual adjustments required as part of the set-up process. In this research, a novel servovalve concept is investigated that has the potential to be more cost-effective. In particular, for the first time, a piezoelectric first-stage actuator is developed to move a servovalve spool using the deflector jet principle; this is especially suited to aerospace actuation requirements. In the new valve, the conventional electromagnetic torque motor is replaced by a multilayer bimorph piezoelectric actuator. The bimorph deflects a jet of fluid to create a pressure differential across the valve spool; hence the spool moves. A feedback wire is used to facilitate proportional spool position control via mechanical feedback. The bimorph is directly coupled to the feedback wire and is immersed in hydraulic fluid. A high-order non-linear model of the valve has been developed and used to predict valve static and dynamic characteristics and is described in this article. This makes use of stiffness constants derived analytically for the bimorph-feedback wire assembly and cross-referenced to finite element analysis predictions. The model of the flow force acting on the deflector is an approximation of the force characteristic found from computation fluid dynamic analysis. The measured characteristics of the prototype valve are in good agreement with the simulation results and prove that the operational concept is viable.
Article
A direct-drive valve (DDV) system is a kind of electrohydraulic servo valve system, in which the actuator directly drives the spool of the valve. In conventional DDV systems, the spool is generally driven by an electromagnetic actuator. Performance characteristics such as frequency bandwidth of DDV systems driven by the electromagnetic actuator are limited due to the actuator response property. In order to improve the performance characteristics of conventional DDV systems, in this work a new configuration for a direct-drive valve system actuated by a piezostack actuator with a flexible beam mechanism is proposed (in short, a piezo-driven DDV system). Its benefits are demonstrated through both simulation and experiment. After describing the geometric configuration and operational principle of the proposed valve system, a governing equation of the whole system is obtained by combining the dynamic equations of the fluid part and the structural parts: the piezostack, the flexible beam, and the spool. In the structural parts of the piezostack and flexible beam, a lumped parameter modeling method is used, while the conventional rule of the fluid momentum is used for the fluid part. In order to evaluate valve performances of the proposed system, an experimental apparatus consisting of a hydraulic circuit and the piezo-driven DDV system is established. The performance characteristics are evaluated in terms of maximum spool displacement, flow rate, frequency characteristics, and step response. In addition, in order to advocate the feasibility of the proposed dynamic model, a comparison between simulation and experiment is undertaken.