A hydraulic servovalve is a servo (see Servo, Hydraulic – Description) with a device (either flapper nozzle or jet pipe) used to position the servo. The term electrohydraulic servovalve is often used because servovalves are controlled through an electrical signal. Servovalves are normally used when accurate position control is required, such as control of a primary flight control surface. Position control is achieved through a closed loop control system, consisting of command sensor, feedback sensor, digital or analog controller, and the servovalve. Servovalves can be used to control hydraulic actuators or hydraulic motors. When a servoactuator is used to control an actuator, the servovalve and actuator combination are often referred to as a servoactuator. The main advantage of a servovalve is that a low power electrical signal can be used to accurately position an actuator or motor. The disadvantage is complexity and cost which results from a component consisting of many detail parts manufactured to very tight tolerances. Therefore, servovalves should only be used when accurate position (or rate) control is required. For more information on closed loop position control using electrohydraulic servoactuators, see Servovalve, Hydraulic – Position Control. Standard servoactuator terminology is provided at the end of this section.
A schematic of a servoactuator is shown in Figure 1. The actuator is included to show how the servovalve and actuator components work together. The primary components in a servovalve are a torque motor, flapper nozzle or jet pipe, and one or more spools. The flapper/nozzle (alternatively jet pipe) and the spool valve are considered “stages”. A stage provides hydraulic force amplification: flapper/nozzle or jet pipe goes from low power electrical signal to spool Δp and the spool valve amplifies Δp on the actuator. The servovalve shown in Figure 1 is a 2 stage servovalve. Almost all servovalves are 2 stage, but some 3 stage designs exist. A 3 stage servo has an additional spool valve between the 1st spool valve and the actuator. The 1st spool valve provides a spool Δp to the 2nd spool valve.
The servovalve shown in Figure 1 uses a flapper nozzle. A servovalve has a hydraulic pressure inlet and an electrical input for the torque motor. The input current controls the flapper position. The flapper position controls the pressure in Chambers A & B of the servo. So, a current (+ or -) will position the flapper, leading to a delta pressure on the servo, which cause the servo to move in one direction or the other. Movement of the servo ports hydraulic pressure to one side of the actuator or the other, while porting the opposite side of the actuator to return. Operation of a servovalve is described in more detail below.
Figure 1 Flapper Nozzle Servoactuator
Flapper Nozzle System
Flapper position is controlled by the electromagnetic torque motor (see top portion of Figure 1). The torque developed by the torque motor is proportional to the applied current. Currents are generally small, in the milliamp range. A torque motor consists of two permanent magnets with a coil winding attached to a magnetically permeable armature. The armature is part of the flapper piece. When a current is applied to the coils, magnetic flux acting on the ends of the armature is developed. The direction of the magnetic flux (force) depends on the sign (direction) of the current. The magnetic flux will cause the armature tips to be attracted to the ends of the permanent magnets (current direction determines which magnetic pole is attracting and which one is repelling). This magnetic force creates an applied torque on the flapper assembly, which is proportional to applied current. In the absence of any other forces, the magnetic force would cause the armature to contact the permanent magnet and effectively lock in this position. However, other forces are acting on the nozzle, such that flapper position is determined through a torque balance consisting of magnetic flux (force), hydraulic flow forces through each nozzle, friction on the flapper hinge point, and any spring (wire) connecting the flapper to the spool (which is almost always installed used in servovalves to improve performance and stability).
As the applied current is increased, the armature and flapper will rotate. As the flapper moves closer to one nozzle, the flow area through this nozzle is decreased while the flow area through the other nozzle increases. The flapper generally rotates over very small angles (~ 0.01 rad) and the gap (G in the figure) is around 0.002 – 0.003 inches. If the gap, G, between the magnet and the flapper end gets too large, the torque motor may latch and become inoperative due to limited available torque from the torque motor.
The flapper nozzle consists of the flapper, two inlet orifices (O1 and O2), two outlet nozzles (n1 and n2), nozzle backpressure nozzle (n3) and usually a feedback spring. As described above, the torque motor positions the flapper, which in turns controls the flow through the nozzles. The inlet orfices, O1 and O2, are important as they create a pressure volume whose pressure is controlled by the flapper.
Figure 2 Flapper Nozzle Geometry
Referring to figure 2, for the flapper nozzle to control flow in a linear manner, the relationship
must be maintained. This relationship implies that the circumferential area created by the flapper distance to the nozzle must be smaller than the nozzle diameter, such that the circumferential area controls flow and not the nozzle diameter. In this way, the flow area varies linearly with flapper position. Also, the torque motor materials, windings and overall design features lead to accurate control of torque such that small movements of the flapper are possible. This leads to accurate control of the pilot spool, which in turns provides accurate control of the actuator.
The goal of the flapper and nozzles is to control the pressure acting on both sides of the pilot spool. When the flapper is in the neutral position, the nozzle flow areas are equal and the pressures Pn1 and Pn2 are equal. When the flow areas and inlet nozzle pressures are equal, the flow forces through each nozzle keep the flapper centered in the neutral position. For a zero lapped pilot spool valve, there would be no flow into or out of the actuator chambers. As the flapper moves towards one of the nozzles, the outlet flow area is reduced for this nozzle. Outlet flow area increases for the other nozzle. For example, looking at Figure 1 let the flapper move towards the n1 nozzle. This will reduce the outlet flow area and the pressure Pn1 will increase. At the same time, the outlet flow area at the n2 nozzle will increase and the pressure Pn2 will decrease. A delta pressure Δp = Pn1 – Pn2 will occur across the pilot spool piston and the pilot spool will displace to the right. High pressure fluid will then flow to the PA actuator chamber while the PB actuator chamber is ported to return. Depending on the size of the flapper and nozzles, the Δp across the pilot spool is limited in magnitude (200-300 lb range for medium size aerospace applications).
Most servovalves incorporate a feedback spring (wire) between the pilot spool and the flapper. This wire is shown as a dotted blue line in Figure 1. Examining Figure 1, if the flapper moves to the left, the Δp on the pilot spool moves the spool to the right. The feedback wire will then pull the flapper back towards the neutral position. Hence the feedback wire provides a stabilizing force to the flapper and helps improve stability and response of the flapper system. This same affect can be done electronically by putting a feedback sensor (usually a linear variable differential transducer) on the pilot spool. The output of the sensor is fed back electronically to reduce the current command and allow the flapper to move back to the neutral position.
Jet Pipe
Another method to control the pilot spool is to use a jet pipe configuration. The jet pipe is an alternative to the flapper nozzle system; however, a similar torque motor is used to control the jet pipe position. A schematic of a jet pipe servoactuator is shown in Figure 3.
Figure 3 Jet Pipe Servoactuator
The jet pipe converts kinetic energy of the moving fluid into static pressure. When the jet pipe is centered between the 2 receiver holes in a receiver block, the pressure on the servo is equal. However, when the jet pipe is rotated toward one of the receiver holes, the pressure at this receiver hole is greater than the other receiver hole, thus creating a load imbalance on the servo. Figure 4 shows a schematic of the jet pipe illustrating how pressure varies between the receiver holes as the jet pipe is rotated.
Figure 4 Jet Pipe Operation
The stagnation pressure at the tip of the jet pipe is given by Bernoulli’s equation as
This is the stagnation pressure at the midstream of the flow and would represent the maximum pressure given by the jet pipe. From the center of the jet stream, the pressure drops off as shown in the Figure 4(a). In the Figure 4(a) configuration, the pressures on both sides of the servo are equal (Δp=0). In Figure 4(b), the jet pipe has been rotated to the right. This has the effect of increasing the pressure on the right side of the servo and reducing pressure on the left side. The servo will then move to the left. As a general rule, movement of the jet pipe is sufficiently small such that the differential pressure will vary linearly over the range of jet pipe travel.
Optimization of nozzle performance is done by experimentation. There is a relationship between the nozzle diameter and receiver hole diameters, which usually must be developed through testing. Also, the distance from the nozzle exit to the receiver is important (L = 2 Dn has been suggested in literature) as well as the distance between the receiver holes. In general, the receiver holes should be as close together as possible, to keep P1 and P2 as high as possible. It is desirable to keep receiver holes as large as possible to prevent contamination issues. The goal of a jet pipe design is to achieve the necessary maximum Δp across the pilot spool and maintain tight position control (no different from a flapper nozzle design).
The biggest advantage of the jet pipe valve over the flapper valve is less sensitivity to contamination. Jet pipe orifices are generally larger than flapper nozzle orfices at the expense of more leakage flow. A clogged flapper nozzle orifice will cause a servovalve to go hardover in one direction. A jet pipe valve will generally fail neutral or operate sluggish if the inlet nozzle plugs. However, both configurations are still used today and both have proven to be reliable and accurate in service.
Servo
The servo part of the valve is exactly the same as any servo or spool valve. The function of the servo is the same for either a flapper nozzle or a jet pipe servoactuator. The relationship between flow and Δp through the servo valve is governed by the orifice flow equation. Servo position is determined by a force balance on the spool, which includes the Δp created from the flapper nozzle or jet pipe, friction forces, spring forces and flow forces acting on the spool. For a complete description of a servo, see Servo, Hydraulic – Description.
When the spool is in the neutral position, the servovalve is in the null position. In some applications, a compression spring is installed on each side of the servo to help keep the servo centered. In other applications (spoiler panel servovalves, for example), a spring is installed in one side only which will push the servo in one direction. For flight spoilers the spring would bias the actuator to the retract position. So, in the absence of electrical commands, the spring pushes the servo towards the retract command position allowing hydraulic fluid to flow to the retract chamber. The applied current required to overcome the spring force and return the servo to the null (no flow) position is referred to as the null bias. The null bias current will drift in service due to changes in supply pressure, operating temperatures, wear and other factors. Good servovalve design practice is to keep long term null bias shifts to within ±3% of rated current.
Flow characteristics and the affects of load flow and load pressure drop are determined by the servo or spool portion of the valve. The equations that describe the relationship between these parameters are given in Servovalve, Hydraulic – Equations. To understand the theoretical performance of a servovalve, these mathematical relationships must be understood.
Actuator
The actuator part of the servoactuator has the same characteristics as any hydraulic actuator. See Actuator, Hydraulic – Description for more information on actuator designs and characteristics.
Servovalve Flow Characteristics
Plots of typical flow characteristics of a servoactuator are shown below. As stated above, flow characteristics are determined by the servo. Therefore, the orifice flow equation describes the flow characteristics. Several figures are shown below to highlight the behavior of servovalves. These figures represent a zero lapped servovalve, which are the most common in aerospace applications. Servovalves with overlapped or underlapped spools will have different flow characteristics around null (for further explanation see Servo, Hydraulic – Description).
Figure 5 shows the relationship between control flow and load pressure. The shape of the curves is determined by the orifice flow equation,
(1)
where i / imax is the applied current expressed as a ratio of maximum current, ρ is the density, Δp is the pressure drop through the servo flow ports and q is the flow rate through the servo. The servo valve will have a higher gain around null then at the end of the spool travel.
Figure 5
Servo Flow Characteristics – Control Flow vs. Load Pressure Figure 6 shows the how
control flow varies with input current. The relationship is linear
for a constant Δp across the servo, as
can be seen from equation (1). Also shown in Figure 6 is the effect
of pressure drop through the servo. Similar to Figure 5, flow
increases for higher Δp. Figure 6
represents an ideal servo, with no hysteresis or friction effects.
Figure 6
Servo Flow Characteristics – Control Flow vs. Input Current Figure 7 shows the
effects of hysteresis on the ideal flow curve. As current is cycled
from 0 to maximum positive value, then to maximum negative value,
then back to zero, the control flow will follow the blue curve.
Hysteresis results from electromagnetic affects in the torque motor
as well as friction in the valve. The electromagnetic effects have a
nonlinear characteristic where the width of the hysteresis loop
varies with the input signal. Friction behaves more like a backlash
nonlinearity which tends to be more constant over the range of
movement (when moving at constant speed). Servovalve hysteresis is
measured at slow speeds where valve dynamic (inertia) effects are
negligible.
Figure 7
Servo Flow Characteristics – Flow Curve with Hysteresis The last curve, Figure
8, repeats the curve in Figure 7, but adds in a null bias offset. A
null bias is created with an unbalanced spring on the servo. The
spring will force the servo to not be centered (offset in one
direction). A certain amount of torque motor current is required to
offset the spring force, which has the affect of shifting the flow
curve. In the example of Figure 8, 1 milliamp (positive) is required
to hold the valve in the no flow (actuator stationary) position. The
main reason for a null bias offset is when current is lost the
actuator will be powered in a desired direction (usually to retract).
Figure 8
Servo Flow Characteristics – Flow Curve with Hysteresis &
Null Bias Servoactuator
Terminology In
the aerospace industry, the document that governs servoactuator
design is SAE Aerospace Recommend Practice (ARP) 490. A copy of this
document is available form Society of Automotive and Aerospace
Engineers (SAE). Important definitions are provided below.
Control Flow - Flow through
valve ports (to actuator). Control flow is referred to as no load
flow when there is zero load pressure drop (i.e., pressure due to
load on actuator piston is 0).
Flow Gain - Slope of the
control flow vs. input current curve in any specified operating
region (see previous figures). Typically this refers to the normal
flow region which is outside of the null region and before saturation
near the end of travel.
Hysteresis - Difference in
the valve input currents required to produce the same valve output
during a single cycle of valve input current, expressed as a percent
of rated current (see previous figures). A nominal value for
hysteresis is ±3% of rated current.
Linearity - Degree to which
normal flow curve conforms to normal flow gain line with other
operational variables held constant.
Load Pressure
Drop - Differential pressure between control ports. Control ports are
the ports to the actuator.
Null - Condition where the
valve supplies zero control flow at zero load pressure drop (actuator
is held stationary at the null position).
Null Bias - Input current
required to bring the valve to null, excluding the effects of valve
hysteresis.
Null Region - Region about
null where effects of lap in the output stage predominate.
Null Shift - Change in null
bias due to changes in supply pressure, temperature and other
operating conditions, including wear.
Polarity - Relationship
between direction of control flow and direction of input current.
Pressure Gain - Rate of
change of load pressure with input current at zero control flow
(psi/ma, for example).
Rated Current - Specified
input current (milliamp) of either polarity to produce rated flow.
Rated Flow - Specified
control flow corresponding to rated current and specified load
pressure drop.
Servovalve - A electrical
input, flow control valve, which can provide continuous control.
Output stage flow is a direct function of current.
Stage - Any device where
hydraulic amplification occurs in a servovalve. Three most common
stages are flapper valve, jet pipe and spool valve. Servovalves can
be single-stage or two-stage. In rare cases, a servovalve may have 3
stages. Threshold - Increment
of input current required to produce a change in valve output,
expresses as a percent of rated current. A normal value of threshold
is ±1% of rated current. Selecting a
Servovalve or ServoActuator Servoactuators are
inherently complicated with nonlinear behavior. In addition,
servoactuators are associated with electronic controllers and closed
loop position control loops. When selecting a servovalve or
servoactuator a large number of factors need to be examined. From a
performance perspective, the critical items to look at are covered in
Servovalve, Hydraulic – Sizing.
Beyond performance factors covered in Servovalve,
Hydraulic – Sizing, other items would include Pressure
Rating – make sure valve is rated for your system pressure Temperature
Rating – valve should be rated for fluid temperatures and
applicable environmental temperatures Rated
Flow – catalog data for servovalves will provide data for
single operating condition which consists of flow rate, inlet
pressure and applied current. This data is a design operating point
for the valve and can be used for initial sizing studies. Rated
Pressure - catalog data for servovalves will provide data for
single operating condition which consists of flow rate, inlet
pressure and applied current. This data is a design operating point
for the valve and can be used for initial sizing studies. Rated
Current - catalog data for servovalves will provide data for a
single operating condition which consists of flow rate, inlet
pressure and applied current. This data is a design operating point
for the valve and can be used for initial sizing studies. Closed
Loop System Interactions – when using a servo in a closed loop
system, the closed loop system interactions (both stability and
closed loop performance) should be evaluated (see Servovalve,
Hydraulic – Position Control) Valve
Materials – should be sufficient to pass proof and burst
testing, not be susceptible to corrosion and other environmental
considerations, and not cause any problems under temperature extremes Failure
Modes – the main failure modes are servo valve jamming in any
position from full closed to full open and a hardover command.
Servovalves are normally closely monitored through the electronic
controller for any abnormal behavior.
Electrical
Interfaces – electronic controller output signals need to be
compatible with torque motor requirements. Electrical connectors
should be high standard (such as military standard) to prevent
deterioration in service (which can lead to hardover failures).
Envelope/Mounting
– servoactuator must fit in the available space and have
mounting features that allow the unit to be mounted to primary
support structure Hydraulic
Fittings – hydraulic connections must be compatible. Inlet and
outlet ports will often have different size connectors to prevent
cross connecting hydraulic lines during installation. Applications of
Servoactuators Servoactuators will be
used for any hydraulically actuated control where precise position or
rate control is required. The most common application for
servoactuators is on primary flight control surfaces. A
servoactuator will be used on any primary flight control surface that
is hydraulically actuated. On primary flight control surfaces,
servoactuators will be part of an overall power control unit (PCU).
More information on power control units can be found in Power
Control Units, Hydraulic – Description. Other applications
for servovalves in commercial aircraft would be flap or slat control,
spoiler panels and nosewheel steering. On military applications,
servovalves can be used to control many other systems on the aircraft
such as thrust vectoring.
Servoactuator/Servovalve
Qualification See Qualification
- Hydraulic Components for discussion on servovalve and
servoactuator qualification and required certification testing.
