Using simulation software to simplify DSP-based Electro-Hydraulic Servo Actuator Designs: Part 1

Richard Poley, Texas Instruments

August 09, 2006

Richard Poley, Texas Instruments August 09, 2006

A basic building block of many embedded control applications, the electro-hydraulic servo system is used in a wide range of environments. These include manufacturing systems, materials test machines, active suspension systems, mining machinery, fatigue testing, flight simulation, paper machines, ships and electromagnetic marine engineering, injection moulding machines, robotics, and steel and aluminum mill equipment.

Hydraulic systems are also common in aircraft, where their high power-to-weight ratio and precise control makes them an ideal choice for actuation of flight surfaces.

Hydraulic actuators are characterized by their ability to impart large forces at high speeds and are used in many industrial motion systems. In applications where good dynamic performance is important it is common to contain the actuator in a servo loop comprising a feedback transducer and electronic controller. The majority of electronic servo-controllers used in these systems are analog based implementations of the well-known PID (Proportional-Integrative-Derivative) type.

The most basic and obvious requirement of such systems is that the device must be able to compute the control algorithm swiftly enough to keep up with the real-time demands of the system. In many cases, such as with the simple PID controller, the control task is relatively simple. The vast majority of electronic closed loop controllers are based on simple analog circuit designs offering robust, low cost implementations of the well known PID control strategy.

Moving beyond MCU-based PID control
While this approach works well in systems with simple topology and limited bandwidth, in many cases the simple MCU-based PID controller is not enough. To implement more complex control strategies where it is often necessary to perform additional processor tasks, more CPU bandwidth is required. Here it is desirable to select a processor which is optimized to perform real-time computations. A multi-bus architecture and rich instruction set makes DSPs well suited to executing demanding real-time control algorithms.

The use of digital signal processor based controllers in such applications is sometimes avoided because it is assumed that the shift from the well understood MCU-based PID methodologies to one based on DSP will require learning new programming languages, as well as dealing with difficulties in testing and de-bugging the code.

However, recent developments in simulation software such make it possible to bypass such difficulties and automatically generate optimized source code which may be compiled and run on the target processor directly from the simulation environment.

This "hardware-in-the-loop" approach enables the control algorithms generated by the model to be executed on a real target processor during simulation, and increases the level of design confidence. The ease with which control algorithms can be created and modified in this manner can save months of development time and leads to earlier error detection compared with traditional hand coding methods.

In this series of articles we will cover the basics of electro-hydraulic servo systems, pointing out where DSP-based control loops would be useful, and describe the mathematical models for the various plant elements that can be developed using Simulink. Also described here is a case study of a hydraulic control system, fitting real data to the model to validate its behavior.

The processor used to provide the basic control is the TMS320C28x (Figure 1, below), with a 32 bit fixed point DSP core. To optimize it for such applications, several libraries have been made available for the platform, including an extensive library of optimized control algorithms.

Figure 1. TI DSP320C28X block diagram

The basics of electro-hydraulic servo control
Although electrical motors are sometimes used in many of servo control applications, motion control systems requiring either very high force or wide bandwidth are often addressed more efficiently with electro-hydraulic rather than electromagnetic means. In general, applications with bandwidths of greater than about 20 Hz or control power greater than about 15 kW, may be regarded as suitable for servo-hydraulic techniques.

Apart from the ability to deliver higher forces at fast speeds, servo-hydraulic systems offer several other benefits over their electrical counterparts. For example, hydraulic systems are mechanically "stiffer", resulting in higher machine frame resonant frequencies for a given power level, higher loop gain and improved dynamic performance. They also have the important benefit of being self-cooled since the driving fluid effectively acts as a cooling medium carrying heat away from the actuator and flow control components. Unfortunately hydraulic systems also exhibit several inherent non-linear effects which can complicate the control problem.

Typical Hydraulic System. A typical position controlled hydraulic system consists of a power supply, flow control valve, linear actuator, displacement transducer, and electronic servo-controller. The servo controller compares the signal from the feedback displacement transducer with an input demand to determine the position error, and produces a command signal to drive the flow control valve. The control valve adjusts the flow of pressurized oil to move the actuator until the desired position is attained: a condition indicated by the error signal falling to zero. A force controlled hydraulic system operates in a similar way, except that the oil flow is adjusted to achieve an output force, measured by a suitable transducer.

As shown in Figure 2, below, all hydraulic servo systems contain at their core six basic building blocks: the hydraulic power supply, the flow control valve, the liner hydraulic actuator, displacement transducer, and servo controller.

Figure 2. Block Diagram of a Position Controlled Hydraulic Servo System

Hydraulic Power supply. System oil pressure depends on various factors. Low pressure means less leakage, but physically larger components are required to develop a given force. High pressure systems suffer from more leakage, but have better dynamic performance and are both smaller and lighter: significant advantages in mobile and aircraft applications. In many high performance systems 3,000 psi (approximately 210 bar) is a standard choice of system pressure.

Oil is drawn from a reservoir (tank) into a rotary vane or piston pump, driven at constant speed by an electric motor. The oil is driven at constant flow rate into an adjustable pressure relief valve, which regulates system pressure by allowing excess oil to return to the reservoir once a pre-defined pressure threshold has been reached.

Pressurized hydraulic oil is carried to the servo-valve through a system of rigid or flexible piping, possibly fitted with electrically operated shut-off valves to control hydraulic start-up and shut-down sequences. Oil is returned from the valve to the tank through a low pressure return pipe, which is often fitted with an in-line heat exchanger for temperature regulation of the oil.

Flow control valve. The electro-hydraulic flow control valve acts as a high gain electrical to hydraulic transducer, the input to which is an electrical voltage or current, and the output a variable flow of oil. The valve consists of a spool with lands machined into it, moving within a cylindrical sleeve. The lands are aligned with apertures cut in the sleeve such that movement of the spool progressively changes the exposed aperture size and alters differential oil flow between two control ports.

Figure 3: Diagram of Three and Four-way Flow Control Valve Spool

Figure 3 above shows the spool configuration of a typical "3-4" flow control valve. The ports are labeled P (pressure), T (tank), and A and B (load control ports). The spool is shown displaced a small distance (xv) as a result of a command force applied to one end, and arrows at each port indicate the direction of fluid flow which results.

With no command force applied (Fv=0), the spool is centralized and all ports are closed off by the lands resulting in no load flow. In the context of hydraulic servo-systems, flow control valves fall broadly into two main categories: proportional valves and servo-valves.

Proportional valves use direct actuation of the spool from an electrical solenoid or torque motor, whereas servo-valves use at least one intermediate hydraulic amplifier stage between the electrical torque motor and the spool.

A major advantage of proportional valves is that they are largely unaffected by changes in supply pressure and oil viscosity. However, the relatively large armature mass and large time constant associated with the coil means that these valves generally have poorer dynamic performance compared with servo-valves of equivalent flow characteristics.

In recent years, "servo-proportional" valves have begun to appear with shorter spool displacements and lighter spools, giving dynamic performance which approaches that of true servo-valves but at a much lower cost.

The basic servo-valve produces a control flow proportional to input current for a constant load. While the dynamic performance of a servo-valve is influenced somewhat by operating conditions (supply pressure, input signal level, fluid and ambient temperature and so on) a major advantage is that load dynamics do not affect stability, unlike single stage proportional valves.

Servo valves usually have superior dynamic response, although their close internal machining tolerances make them relatively expensive and susceptible to contamination of the hydraulic fluid.

Two stage servo-valves may be further divided into nozzle-flapper and jet pipe types. Both use a similar design of electromagnetic torque motor, but the hydraulic amplifier circuits are radically different. Nozzle-flapper type servo-valves are currently by far the most common in high performance servo applications and the description which follows is based on this type of valve.

Figure 4. Cross Section of Nozzle-flapper Type Servo-valve (illustration courtesy of Moog

A cross sectional view of a typical nozzle-flapper type servo-valve is shown in Figure 4 above. High pressure hydraulic oil is supplied at the inlet pressure port (P), and a low pressure return line to the oil reservoir is connected to the tank port (T). The two hydraulic control ports (A and B) carry the control oil flows to and from the load actuator.

Linear Hydraulic Actuator. A hydraulic actuator is a device which converts hydraulic energy into mechanical force or motion. Actuators may be divided into those with linear movement (sometimes called rams, cylinders or jacks), and those with rotary movement (rotary actuators and motors).

Linear actuators may be further sub-divided into those in which hydraulic pressure is applied to one side of the piston only (single acting) and are capable of movement only in one direction, and those in which pressure is applied to both sides of  the piston (double acting) and are therefore capable of controlled movement in both directions.

Linear actuators may also be classified as single-ended, in which the piston has an extension rod on one end only, or the double-ended type which have rods on both ends. Single-ended actuators are useful in space constrained applications, but unequal areas on each side of the piston results in asymmetrical flow gain which can complicate the control problem.

Double-ended actuators have the advantage that they naturally produce equal force and speed in both directions, and for this reason are sometimes called symmetric or synchronizing cylinders. Hydraulic motors are a separate class of actuator, in which the speed and direction of a rotating output shaft is regulated by the flow control valve.

Figure 5. Linear Servo-Hydraulic Actuator Assemblies (illustration courtesy of Moog)

The description which follows is based on a linear, double-acting, double-ended actuator, a type used in many industrial applications. A cross section of such an actuator is shown in Figure 6. The actuator consists of a rod and central annulus, and incorporates low friction seals fitted to the piston annulus and at each of the cylinder end caps to minimize leakage. Control ports are drilled into each end of the cylinder to allow hydraulic fluid to flow in and out of the two chambers.

Figure 6. Cross-sectional Diagram of Double-ended, Double-acting Linear Actuator

The position of the piston is determined by the hydraulic fluid pressures in the chambers on either side of the central annulus, and may be adjusted by forcing fluid into one control port while allowing it to escape from the other. In the diagram above, hydraulic fluid is shown entering control port A while escaping from port B.

This causes an increase in fluid pressure in the chamber to the left of the piston annulus, and a decrease in pressure in the right chamber. The net pressure difference exerts a force on the active area of the annulus which moves the piston to the right as shown. Adjustment of piston position is therefore a matter of controlling the differential oil flow between the two actuator control ports.

Displacement Transducer. Position transducers are usually collocated with the actuator, and often attached directly to the piston rod. Various types of feedback transducer are in use, including incremental or absolute encoders, inductive linear variable differential transformer (LVDT's) and rotary variable differential transformer (RVDT's), linear and rotary potentiometers, and resolvers. In industrial applications employing linear displacement control, the LVDT is a common choice of feedback transducer due to its accuracy and robustness.

The transducer is usually selected such that its bandwidth is ten times or so higher than that of the servo-valve and actuator, and is often omitted from a first analysis of the system.

Figure 7. Block Diagram of the Analogue Controller

Servo Controller. In the Figure 7 above of the layout of the servo-controller block, the error amplifier continuously monitors the input reference signal (ur) and compares it against the actuator position (Up) measured by a displacement transducer to yield an error signal (Ue).

Ue = ur - up

The error is manipulated by the servo controller according to a pre-defined control law to generate a command signal (uv) to drive the hydraulic flow control valve. Most conventional electro-hydraulic servo-systems use a PID form of control, occasionally enhanced with velocity feedback. The processing of the error signal in such a controller is a function of the proportional, integral, and derivative gain compensation settings according to the control law: 

where Kp, Ki, and Kd are the PID constants, ue is the error signal and uv is the controller output.

With a clear idea of the basic electro-mechanical building blocks and their relationship to one another, we are now ready to develop the basic mathematical models of the hydraulic components and make use of the Simulink modelling software from The Mathworks Inc.

 This will be the subject of Part 2 in this series: "Using modelling tools to simplify hydraulic PID system design."

Richard Poley is Field Application Engineer at Texas Instruments with focus on digital control systems.

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