Fig. 3 — With 4 sec to complete a motion cycle, very little error exists between the command and actual positions.
Fig. 4 — Increasing the work cycle frequency creates a large error between the command and the measured position profiles — unacceptable for most applications.
Fig. 5 — Demanding peak flow that a pump is incapable of delivering causes supply pressure reductions that decimate servoloop performance.
As explained last month, electrohydraulic control using servo or proportional valves usually is achieved with a power unit that delivers constant pressure to the valve. Last month described a test system that used only a pressure-compensated pump with no accumulator. It was controlled by a command profile requiring four seconds to complete one cycle.
We continue the topic by discussing a second test, which also will be conducted without a supplementary accumulator. However, the motion command profile will be speeded up so its cycle will be completed in only one second. It has the same general shape as last month's Figure 2, but everything will be commanded to happen four times faster.
When the first command profile is used — that is, the one that takes four seconds to complete — the results are shown in Figure 3. Four different traces are shown. The command profile and the measured position (piston motion) lie almost atop each other. This occurs because a very good servo loop has been established. Very little error exists between the command and actual positions because servo amplifier gain is very high.
Figure 3 also shows the actuator speed. This value is calculated by differentiating the measured position with respect to time. It clearly shows the dwell times at extension and retraction, the acceleration and deceleration periods, and the nearly flat-topped "constant speed" portions.
Rather than being caused by actual speed variations, the squiggles at upper and lower extremes of the speed trace are caused more by the inevitable consequence of measurement noise that is enhanced by the differentiation of measured data.
The supply (P-port) pressure is also shown as a slightly cusped curve with some small squiggles caused by the dynamic response of the pump compensator. It is traced out near the top of the graph axes.
The pump's compensator initially was adjusted to produce a deadhead pressure of about 1400 psi. as expected, the average dynamic value of supply pressure is slightly less than 1400 psi. But more importantly, the pressure variation is only about 100 psi between the highest and lowest measured values. This performance probably is completely acceptable for all but the very critical applications, and it was accomplished without an accumulator.
Speeding it up
Figures 4 and 5 show the results of reducing the cycle time from one cycle every four seconds to one complete cycle each second. This is a rather severe increase, but not unreasonable in today's industrial environment. Again, no accumulator is used in this test.
Note in Figure 4 the poor compliance between the command and measured position profiles. A large error exists between them, which is very poor performance of the servo loop and probably unacceptable in any application. The cylinder velocity no longer is a clean trapezoid, as in Figure 3. The velocity has been corrupted by the pump's inability to compensate for the rapid flow demand variations.
The reason for the poor following performance in the servo loop is revealed in Figure 5 — the supply pressure. It varies from about 1900 psi to a low of about 200 psi. At the same time, the servovalve current swings from -70 mA to +70mA — the saturation limits — as the servo loop tries to keep the cylinder in synch with the command profile.
Two major issues are at play here: First, the flow demand profile is requesting flow increases and decreases that the pump's pressure compensating mechanism cannot keep up with. This contributes significantly to the wild fluctuations in pump supply pressure. Second, the reduction in the command profile time was "constructed" with the same total cylinder travel distance. This quadruples the peak flow demand in this faster cycle.
Although not shown, the ideal peak flow is about 56 in. 3 /sec, as compared to 14 in. 3 /sec in the first test. Clearly, the pump is being asked to increase and decrease its displacement by an amount four times as great, thereby "penalizing" the pump — that is, imposing a much more severe test condition in this second case. But this is realistic for an industrial application that must provide variable production rates. The cylinder must travel the same amount, and the application requires that it do so four times faster. Now, that revelation having been made, it also follows that the average flow must be four times greater in the second case — not just the peak flow. Clearly, then, the higher cycle rate would benefit from an augmenting accumulator connected directly to the inlet port of the control valve.
Summing it up
The pump used in these tests is advertised to have about a 145 msec response time. That is, it can go from zero to full stroke in that amount of time. So a simple-to-use rule-of-thumb is needed to help designers know when an accumulator is absolutely necessary, and when they can get by without one. We need only compare the published pump response time to the time required to change the flow in the application.
Recall from the first test (with the four second profile) that 0.6 sec were expended in changing the demand flow from zero to maximum. This is four times longer than the pump's response time. In contrast, the higher cycle rate profile reduced the flow increase time to only 150 msec, which caused a clear failure in performance.
Here is the rule-of-thumb to decide if you can get by without an augmenting accumulator in conjunction with a pressure compensated pump: No accumulator is required if the flow change time in going from zero to maximum flow, or from maximum to zero flow, is three or more times longer than the full stroke response time of the pump.
This application, with the shorter cycle time, could benefit from an accumulator. With an accumulator installed, the opening of the valve at flow-increase time would use flow from the accumulator to fulfill the sudden flow demand. This applies not only to discharging the accumulator, but for recharging as well. The pump leisurely changes displacement to meet the average flow demands., and the accumulator reacts to rapid changes in demand.
Book teaches servohydraulics
Designer's Handbook for Electrohydraulic Servo and Proportional Systems covers electrohydraulic system design, control, and analysis by teaching principles and reviewing practical problems from the real world. Much of the information updates and expands on important concepts and calculations from such groundbreaking works as George Keller's Hydraulic System Analysis and Jean Thoma's Hydrostatic Power Transmission.
Covers everything from fundamental hydraulic circuit analysis to dynamic testing of closed-loop electro-hydraulic motion control systems.
Packed with more than 500 pages of useful formulas and dozens of fully worked examples.
Twenty chapters cover hundreds of topics, including components, systems, circuit analysis, modeling, servomechanisms and control, simulation, and testing.
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