Hydraulic actuators translate hydraulic fluid flow into linear mechanical motion. A “balanced” linear actuator is shown in Figure 1 below. The actuator is considered balanced because the piston pressure areas in Chamber A and Chamber B are equal. Balanced actuators are used on primary flight control surfaces and other locations where loads are the same in both directions. Normally one or more lugs are located on the housing for connection to fixed structure. The moving end is connected directly to the load, such as a landing gear, or through a linkage to provide rotation movement, such as a primary flight control surface.
Figure 1 Balanced Actuator Schematic
Figure 2 shows an unbalanced actuator where extend and retract piston areas are different (due to the piston rod on the retract side). Unbalanced actuators are used where loading is different in two directions, such as a spoiler panel actuator, thrust reverser or landing gear actuator, or where space constraints prohibit a balanced actuator. The actuator in Figure 2 also has a Linear Variable Differential Transducer (LVDT) installed in the center of the piston rod and attached to the fix housing. This transducer provides actuator stroke position information for the control electronics (used for closed loop position control and fault detection).
Figure 2 Unbalanced Actuator Schematic
Figure 3 shows a dual tandem actuator. This actuator is controlled by two pistons and four chamber pressures. Normally, both pistons are controlled by separate, but with identical control valves. The two valves would be supplied by separate hydraulic systems for redundancy. In event one control valve fails or hydraulic system fails, the actuator still operates but at ½ of the force capability.
Figure 3 Balanced Dual Tandem Actuator Schematic
The most important characteristics for a hydraulic actuator are
Load Capability – load capability is the primary design criteria in sizing an actuator. Accurate load determination is critical and must include the external force as well as friction, inertial and disturbance forces. Also, in many applications load varies with stroke. For example, the load force on a landing gear actuator usually varies greatly as a function of gear position. As another example, “bucket” type thrust reverser actuators have an aiding load on extend and an opposing load on retract. In fact, aero forces can pull the buckets open, extending the actuator, and the possibility of actuator extend chamber cavitation must be investigated (use of accumulators are common in thrust reversers systems). Loads for flight control surfaces vary widely as function of flight condition.
Piston areas in both chambers – this determines the amount of force the actuator can output
Flow port restrictors (or orifices) – these determine the amount of flow into and out of the actuator chambers, and hence determine the velocity of the actuator piston. Note, flow port restrictors are not always used but are included here for discussion. Generally, the flow pipe leading up to an actuator is sized for fluid velocity and pressure drop, not to control flow into and out of actuator chambers. Actuator velocity may be set either by the inflowing or the outflowing orifice (or valve), depending on the orifice diameters and pressure drops across both orifices. Also, on the chamber with increasing volume it is important that the sufficient flow into the chamber is possible without cavitation. A secondary effect of orifice is to damp upstream pressure fluctuations to reduce effects on piston movement.
Housing geometry – affects actuator installation characteristics and clearances to structure and surrounding components. The envelope requirements often dictate housing geometry and mounting lug locations.
Piston rod diameter – From a structural point of view, an actuator behaves like a column loaded in tension and compression. Column loading characteristics are determined by the housing and piston rod area. Maximum actuator static load capability is set by the lesser of the column load capability and piston area hydraulic (Force = Pressure x Piston Area).
Piston seals – Contribute to overall actuator friction and hydraulic flow leakage between chambers. Ideally friction is low (minimizes heat generation) and leakage is zero. Seals must also hold up over the life of the actuator (100,000 cycles or more) in all environmental conditions. Seals may be dual redundant providing greater reliability but higher friction. Some actuators do not use seals on the piston head but instead maintain close machine tolerances on the piston outer diameter and inner bore diameter. This type of design will have some amount of hydraulic laminar leakage across the piston head proportional to the differential pressure across the piston. Additionally, this type of design requires pressure equalization grooves on the piston head outer diameter in order to prevent the piston head from hydraulically locking against the actuator bore.
Housing Seals - Contribute to overall actuator friction and any external leakage. Ideally, actuator friction is low (minimizes heat generation) and external leakage is negligible. Housing seals may also be dual redundant. In addition, scraper seals may be installed at the external housing / piston rod interface. Scraper seals help clean frost/ice/debris off of the piston rod to help prevent damage to the pressure seals and helps prevent foreign particles from entering into the actuator.
Chamber Volumes – Set by actuator travel requirements and required piston area. Piston area and chamber volume are used to compute frequency response characteristics of the actuator.
Static Load & Fatigue Load Characteristics – As discussed previously, actuator load and rate requirements determine necessary piston areas. However, the load bearing structural elements (parts) in an actuator must be capable of standing limit and ultimate loads and therefore the actuator structural elements must be designed to these loads. Furthermore, fatigue loads (usually defined by a duty cycle spectrum which contains #of cycles, load for each cycle, stroke and rate) can also be significant and must be taken into account when designing an actuator. Actuator piston, rod, and housing must be designed for static and fatigue loads. For actuators which have low cycles over their life, static loads may be the dominant structural design factor. However, for actuators with high cycles over their design life, fatigue often becomes the dominant structural design factor and ends up sizing components within the actuator.
Other features may be installed in an actuator. Locks hold the actuator in either the extend or retract position. Locks are pin or cam that physically locks an actuator in position. Full hydraulic system pressure is normally required to “unlock” an actuator. Position switches (often called “down and locked” or “up and locked” switches for landing gear actuators or “extend” / “retract” switches for other actuators) to provide indication when an actuator has reached a full travel in 1 direction. If they are to signify that the actuator lock is in place, then ideally, the switch is activated by the locking pin and not the piston. Figure 4 shows a switch and plunger that activate off of piston position. As the piston nears the extend position, the switch plunger will ride up the ramp and, at the point of full extension, will activate the switch. The switch is spring loaded to the open position. Normally these switch installations have some means for rigging the switch setting.
A last feature that may be installed in an actuator is a snubber. A snubber is a typically a coil or belleville spring used to slow the piston down prior to the piston bottoming out on the housing (see Figure 4). Snubbers may be required when an actuator has a high load and fast velocity to prevent damage to the load or the housing. Snubbing may also be done by reducing the outlet flow area as an actuator gets to the end of its travel. The smaller flow area reduces outlet flow and damps/slows down the actuator at the end of its travel.
In some applications a hydraulic orifice actuator snubber is tuned to minimize the peak deceleration of the load should the actuator be driven into the end stop. In this case, a hydraulic snubber would be tuned to exhibit a square wave snubbing pressure when the cushion is engage at the actuator full speed condition. This will tend to generate a square wave deceleration of the load thereby minimizing the peak deceleration experienced – dissipates the energy over a longer distance. The snubber is tuned by controlling the snubber orifice area as the actuator travels in the snubber. A full six degree of freedom flight simulator motion base would be one application where a tuned snubber would be incorporated. Here the motion base playload will house training crews along with expensive electronic equipment.
Actuator Friction Characteristics
Friction always opposes actuator motion. So the direction of the friction force vector switches direction with velocity. There are 3 types of friction present in a hydraulic actuator:
Coulomb friction – Coulomb friction is a constant friction force (does not vary with velocity). Coulomb friction represents friction associated with mechanical surfaces rubbing together and includes bearing friction, friction in flight control surface hinges, and so on.
Viscous friction – Viscous friction represents the force required to push hydraulic (viscous) fluid through restrictive passages. This would include the force required (or used) for leakage or to push fluid through any small passages that may exist in the actuator. Viscous friction is small at low velocity and increases linearly with piston velocity.
Seal (Stribeck) friction – Seal friction is the friction of the seals against the housing or piston. Seal friction is high at low velocity and drops off exponentially as actuator velocity increases.
Notionally, the three types of friction are shown in the Figure 5. The total actuator friction is the sum of coulomb, viscous and seal friction. Note how the total friction curve has a minimum value and is dominated by the seal friction at low velocities and by viscous friction at high velocities. Coulomb friction simply raises or lowers the total friction curve.
Figure 5 Seal Friction Characteristics
Friction represents wasted energy that is dissipated as heat within the actuator. Therefore, friction reduces overall actuator efficiency. Normally, it is desirable to keep friction low. However, there are tradeoffs in the potential for leakage and seals to wear prematurely if friction levels are pushed too low. In some cases, a minimum friction is required to provide flutter damping under loss of hydraulic power, or to reduce an actuator’s sensitivity to external disturbances (such as gust loads).
Hydraulic Natural Frequency
Hydraulic natural frequency for an actuator sets bounds on actuator performance. Generally for a bang-bang type actuator (such as a landing gear or thrust reverser actuator), hydraulic natural frequency is not a concern. However, for actuators where cycling occurs at fast rates (such as a servoactuator), hydraulic natural frequency should be evaluated in the design process.
Hydraulic fluid in an actuator chamber is compressible and thus acts like a very stiff spring. The stiffness of the spring is a function of the fluid bulk modulus, piston area and chamber volume. The piston in a hydraulic actuator can then be viewed as a mass with a spring connected on each side of the piston. The basic equation for computing the natural frequency of the actuator is
where the spring rate is based on the two hydraulic springs connected in series. The specific equation to compute the hydraulic natural frequency is provided in the actuator modeling section.
It is not practical to operate an actuator above the actuator’s natural frequency as the actuator will essentially be unstable and uncontrollable. Typically hydraulic actuators have very low damping (ξ = 0.05 to 0.10); hence, when the actuator is commanded to cycle at its natural frequency the output is extremely large and uncontrollable. See Figure 6 for a typical second order response with a damping ratio of 0.05. This would approximate a hydraulic actuator output position divided by input sine wave position command. At the natural frequency the actuator stroke would be amplified ten times the amplitude of the input. Therefore, maximum duty cycles for a hydraulic actuator should be limited to a value much less than the hydraulic natural frequency.
Figure 6 Variation of Output/Input Ratio with Frequency for Damping Ratio of 0.05