4 Actuator Control
4.3 Valve Speed Control
4.3.4 Negative and static load control
η efficiency of the by pass speed control QL flow into the actuator, [volume/time]
ΔpL load dependant pressure drop across the actuator, [pressure]
QP pump flow, [volume/time]
ΔpP pressure rise across the pump, [pressure]
ΔpC pressure drop across the 3-way flow control valve, [pressure]
Qset set flow value of the 3-way flow control valve, [volume/time]
4.3.4 Negative and static load control
A negative load is defined as a load that tries to move an actuator in the same direction as the flow. Negative loads always represent a cavitation threat to the hydraulic system.
They typically appear when a load has to be lowered, see Figure 4.12 to the left.
pP
Figure 4.12 Two systems holding a load and trying to lower it.
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To the left a system without any load holding capability. To the right a system with perfect load holding capability but without proper functionality.
For the simple system shown above, the only pressure build up below the cylinder piston comes from pushing flow through the tank connection of the directional control valve as well as other restrictions (piping, filter, coolers, etc.) in the flow path before it reaches the tank reservoir. A load holding pressure, pL, that will maintain equilibrium with the load and the pump pressure, pP, is necessary. However, if the flow required to build up this pressure is larger than what the pump is capable of delivering (including the flow gearing of the differential cylinder in Figure 4.12) then the pressure line of the pump will cavitate, i.e., in order to obtain equilibrium, pP will try to become negative.
This is impossible and the load runs away.
At the same time, any static load always represents a load drop threat. This may be caused by pipe/hose bursting and in less critical cases (load dropping slowly) by leakage from the pressurized regions to the tank reservoir. The latter is especially a problem when using spool based directional control valves (which is the typical case), as they cannot be made leakage proof.
These 2 problems: Load drop and runaway loads, may be dealt with in several ways.
Basically, the load drop due to pipe/hose bursting is dealt with by mounting a seat valve directly on the actuator, see Figure 4.13 to the right. This gives a leak proof load holding capability, however, it is necessary to lift the poppet/ball from the seat when the load is supposed to be lowered. The opening pressure may be picked up in 3 different ways.
pP
pL
⋅A ϕ
A F
pL
A Arv
ρp
ρp
pP
F
⋅A ϕ
Figure 4.13 To the left is shown part of a system using a pilot operated check valve to lower a load. To the right is shown the same system but with a restrictor-check valve added.
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The pilot operated check valve, described in section 3.2, uses pP as opening pressure, see Figure 4.13 to the left. By increasing the pilot area relative to the seat are, see Figure 3.10, a relatively low pumping pressure (low power consumption) can pilot the check valve open when the load is lowered. The check valve has a relatively weak spring so that it does not offer too much resistance when the load is raised. Ignoring its contribution, the necessary pump pressure to pilot open the valve may be determined from:
pP pump pressure required to pilot open the valve, [pressure]
pL pressure on the piston side of the cylinder, [pressure]
ρp ratio of the pilot area to the seat area, (rx/r)2 , see Figure 3.10 F load, [force]
A piston area of the cylinder, [area]
ϕ cylinder area ratio
The weak spring does, however, also mean that the pilot operated check valve cannot work as a metering (pressure build up) unit to control a runaway load. It would be much too unstable as minor pressure fluctuations would cause consistently closing and opening of the valve. Hence, a pilot operated check valve only solves the load holding problem. Therefore a pilot operated check valve will normally be mounted in series with a restrictor valve, see Figure 4.13 to the right. The restrictor valve is adjusted to offer the flow restriction necessary to build up a sufficient load holding pressure, pL. At the same time it will have a dampening effect on piston speed and pressure fluctuations in the system, and hence improve stability. If the pressure drop across the pilot operated check valve is disregarded and it is assumed that it is piloted fully open, i.e., the pilot piston is resting against a mechanical stop, then the governing equations for the system shown in Figure 4.13 are:
P
pL pressure on the piston side of the cylinder, [pressure]
F load, [force]
A piston area of the cylinder, [area]
ϕ cylinder area ratio
pP pump pressure required to pilot open the valve, [pressure]
QP pump flow, [volume/time]
CD discharge coefficient of the restrictor valve Arv discharge area of the restrictor valve, [area]
ρ mass density of the fluid, [mass/volume]
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Combining Equation (4.8) and (4.9) yields an expression for the correct setting of the discharge area of the restrictor valve for a given pump flow, a desired pump pressure and a maximum load, Fmax:
⎟⎠
⎜ ⎞
⎝
⎛ + ⋅
⋅
⋅
⋅
=
P max
D
P rv
A p F C 2
A Q
ρ ϕ
ϕ (4.10)
The so-called counter balance valve, see Figure 4.14 to the left, which principally corresponds to the pressure relief valve discussed in section 3.3 uses pL as opening pressure.
pP
pL
⋅A ϕ
A F
pL
A
ρp
pP
F
⋅A ϕ
pcr pcr
Figure 4.14 To the left is shown part of a system lowering a load by means of a counterbalance valve. To the right is shown a system lowering a load by means of an over centre valve. Both
systems have a check valve in parallel with the load holding valve.
The crack pressure is set a certain percentage above what is required to maintain equilibrium with the load. This way both the load holding problem as well as the load runaway problem is taken care of. A separate check valve that allow by pass of the flow when lifting the load is necessary.
Neither the pilot operated check valve nor the counter balance valve are well suited for varying loads because they are set to handle the maximum load, thereby causing very high pumping pressures when lowering smaller loads. For systems with greatly varying loads a so-called over centre valve may be used. It is a combination of the pilot operated check valve and the counter balance valve as it uses both pP and pL as opening pressure, see Figure 4.14 to the right. The pump pressure will typically act on an area 3-10 times larger than the area acted upon by the load holding pressure. The ratio between the area
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acted upon by the pump pressure and the area acted upon by the load holding pressure is referred to as the pilot area ratio. Ignoring the pressure in the return line the governing equations for the system shown in Figure 4.14 to the right are:
P
L p
A
p = F +ϕ⋅ (4.11)
P p L
cr p p
p = +ρ ⋅ (4.12)
where
pL pressure on the piston side of the cylinder, [pressure]
F load, [force]
A piston area of the cylinder, [area]
ϕ cylinder area ratio
pP pump pressure required to pilot open the valve, [pressure]
pcr crack pressure of the over centre valve, [pressure]
ρp pilot area ratio of the over centre valve
The crack pressure is defined as the load holding pressure required to open the over centre valve without the aid of any pump pressure. The crack pressure is adjusted by adjusting the initial compression of the spring. Combining Equation (4.11) and (4.12) yields an expression for the correct setting of the crack pressure for a given pump flow, a desired pump pressure and a maximum load, Fmax:
(
p)
Pmax
cr p
A
p = F + ρ +ϕ ⋅ (4.13)
In general, the effect of using both pL and pP as opening pressures tends to stabilize the system. However, too high pilot area ratios will cause instabilities even for over centre valves. The effect of changes in the load is strongly reduced, see Figure 4.15.
0 0.6 1.2 1.8 2.4
0 0.2 0.4 0.6 0.8 1
Fmax
/ F A
/ F
p
max P
Counterbalance valve
Over center valve
Pilot operated check valve w. restrictor
Figure 4.15 The pump pressure dependency on the load for the different types of load holding valves
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4.3.4 Braking
In general, an actuator is brought to a stand still by moving a directional control valve into neutral, see Figure 4.16.
Direction of motion Direction of motion
Figure 4.16 Simple hydraulic system shown lowering a load and breaking.
When the valve goes to neutral the load will, due to inertia, continue its motion. This will lead to a compression of the fluid in branch A and a decompression of the fluid in branch B. Hence pL increases and pP decreases. For large inertia loads and systems with insufficient damping this may easily cause severe problems, both with respect to overloading in branch A due to pressure peaks and cavitation in branch B. These problems are typically handled by inserting a shock valve and a suction valve, see Figure 4.17.
The shock is simply a pressure safety valve, see section 3.3, dimensioned to a rather small flow. It is set a certain percentage above the maximum expected static load pressure. The suction valve is a check valve, see section 3.2, with a very weak spring, so that the back pressure can crack it open and refill the branch in danger of cavitating. If the back pressure/tank reservoir pressure is not high enough, an extra check valve may be inserted in the return line with a somewhat stiffer spring, ensuring sufficient suction pressure.
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Figure 4.17 Simple system shown breaking a load. The system is equipped with a shock valve, a suction valve and a check valve in the return line to maintain a certain back pressure.
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