Oil compression compensation system

A pre-set fluid actuator for setting the roll gap of a rolling mill having a transducer for detecting actuator movement and an electrohydraulic servovalve for controlling its position, a control system for compensating for the compressibility of the actuator fluid caused by the rolling force in a manner that requires no operation of the servovalve by creating a pre-set anticipating compression compensation signal and reducing the output of this signal by the output of a signal representing the actual fluid compression caused by the actual rolling force.

BACKGROUND OF THE INVENTION 
(a) Field of Invention 
In many applications where fluid is employed as a power or energy medium, 
such as oil in a force applying or resisting hydraulic piston cylinder 
assembly, the inherent compressibility of the medium can create a serious 
disadvantage or limitation in the use of the device. Of course different 
mediums have different coefficients of compressibility and while in 
certain applications the design of the device can allow for the 
compressibility factor in others it presents a serious disadvantage 
resulting in curtailed use and/or expensive and inordinate auxiliary 
compensation equipment. For the purpose of explaining the present 
invention and not as a limitation of its application the well known 
problem of hydraulic actuators employed in strip or plate (strip) rolling 
mills to control the roll gap has been selected. 
Many modern day rolling mills incorporate quick acting hydraulic piston 
cylinder assemblies (actuators) to control the gap between the work rolls 
and hence the gauge of the rolled product produced by the mill. Due to the 
fact that the actuator is usually arranged in the direct line of force of 
the rolling force to resist the force the amount or degree of the 
compressibility of the oil results in a "gauge error" since the 
compressibility allows the roll gap to be increased by the amount of the 
compressibility. 
Because the gauge tolerance of rolled strip is generally very critical 
present mill designs have had to adhere to two additional design criteria, 
namely to keep the stroke of the actuator as small as possible and/or 
providing a control system to reposition the moveable element of the 
actuator to compensate for the element's movement due to the 
compressibility of the oil under the rolling force. Both of these criteria 
in many types of mills represent serious, if not prohibitive, limitations. 
Moreover, certain mill designs require long stroke actuators which in the 
past have necessitated some type of control system to compensate for the 
compressibility of the oil. 
Such present day control systems employ some form of a position detector to 
detect fluid compression induced movement of the moveable element of the 
actuator which in turn requires the employment of an electrohydraulic 
servovalve unit. The systems in question usually operate under pressures 
of 4000 PSI in which the rolling force can range between 4,000,000 to 
12,000,000 pounds for reduction mills, which means that considerable 
dynamic fluid energy and response time is inherently involved in effecting 
a correction for the compression of the fluid, more about which will be 
discussed later. The dynamics of this situation can result in the 
production of considerable off gauge rolled strip depending on the 
particular speed of the mill involved. 
(b) Background Printed Information 
Representative of prior publications discussing the concern of actuator 
compliance in rolling mill automatic gauge control systems (AGC Systems) 
may be found in U.S. Pat. Nos. 3,427,839 and 4,102,171 and in articles 
appearing in the "Iron and Steel Engineer" entitled Design and application 
of hydraulic gap control systems by Paul Huzyak et. al. - August, 1984, 
page 13 and Dynamic characteristics of automatic gauge control system with 
hydraulic actuators by Vladimir B. Ginzburg - January, 1984, page 57. 
BRIEF SUMMARY OF THE INVENTION 
The present invention relates to a medium compression compensation system 
for a force applying and/or resisting actuator in which unless compensated 
for the compressibility of the medium creates an adverse effect by 
allowing movement of the moveable element or elements of the actuator. 
Included in the system is a means for detecting and producing a signal 
representing the degree of movement of the moveable element from a pre-set 
no load position and unless otherwise negated effects a repositioning of 
the moveable element to its pre-set position, a means for pre-setting the 
position of the moveable element under a no load condition and providing a 
signal representative thereof including compensation for the compression 
of the medium under an anticipated load, and a means for determining the 
compression of the medium under a loaded condition and modifying the 
detected movement signal by the amount of the actual load determined 
condition. 
In terms of a rolling mill or the like the normal undesired change in the 
roll gap caused by the compression of the oil in the actuator is quickly 
and efficiently compensated for without the operation of the 
electrohydraulic servovalve system by reason of the fact that the signal 
measuring the movement of the actuator due to the compression of the oil 
is prevented from effecting an operation of the servovalve because the 
system has built into it an anticipated pre-set compression factor which 
allows this factor to be modified by a factor representative of the 
compression of the oil caused by the actual rolling force.

DETAIL DESCRIPTION OF THE INVENTION 
With reference to FIG. 1 there is illustrated some of the basic components 
of a well known 4 Hi strip rolling mill 10 for rolling metal strip 
including work rolls 12, backing-up rolls 14, hydraulic piston cylinder 
assembly 16 (actuator), including a position transducer 18 and lastly a 
load cell 20. 
The combined hydraulic electric system that made up the control circuit of 
FIG. 1 and how the components are arranged and operated are also well 
known as indicated in the aforesaid mentioned published articles and for 
which reason will not be described in detail. Starting with the load cell 
20, it feeds an electrical representative signal of the actual rolling 
force to a function generator 24 and to a summing amplifier 26, the latter 
comparing the actual rolling force signal with an anticipated rolling 
force signal produced by an electrical adjusting device 28. The function 
generator 24 also receive a signal 29 of an anticipated rolling force and 
divides alternatively either the actual or anticipated rolling force by 
the value representing the modulus of the mill to produce a signal 
representing the mill stretch (1/M) in accordance with the well known load 
meter AGC formula G=H-F/M where: 
G=rolled gauge. 
H=no load roll setting. 
F/M=actual or anticipated roll force divided by a constant representing the 
mill modulus. 
The summing amplifier 26 sends its signal to the function generator 29 
where it is divided by a factor 1/K representing the actuator compliance 
value to produce a signal equal to the oil compression (Xoc), which signal 
is received by a control amplifier 30. 
This signal will take two different forms i.e. two alternative signals, 
depending on whether or not an actual rolling force is being produced, 
i.e. a signal (Fant-o) representing an anticipated rolling force value or 
(Fant-Fact) representing the difference between actual rolling force value 
and an anticipated rolling force value, discussed later in connection with 
equation 1. As noted in the above referred to published articles - 
actuator compliance is a well known technology and may take into account a 
number of factors depending on the degree of sophistication required, more 
about which will appear below. 
Returning to FIG. 1, the position transducer 18 associated with the 
actuator 16 for detecting movement of its piston produces a signal 
representative thereof which is fed to the control amplifier 30. In 
addition to the three signals already identified being received by the 
amplifier 30, it also receives a signal (H) from an adjusting device 34 
representing the desired gauge of the strip to be rolled by the mill. The 
control amplifier 30 feeds it signal to a power amplifier 36 which in turn 
sends a signal to an electrohydraulic directional servovalve 38 having the 
customary supply inlet at 40 and return at 42. The servovalve may be 
several well known types, for example a Moog Series 72 High Flow Two-stage 
servovalve sold by the Moog Controls Division of Moog Inc., N.Y. U.S.A. 
The servovalve 38 feeds in a controlled manner to the actuator 16. As in 
the case of the servovalve, the other components of the electrohydraulic 
control system, and their operation in an AGC system, are all well known 
as indicated by the above referred to patents and articles. It will also 
be recognized that the control system is essentially a standard load meter 
AGC system which has been modified to incorporate the teaching of the 
present invention. A better appreciation of this incorporation of the 
present invention can be gained by considering the following two 
equations: 
EQU Xoc=(Fant-Fact) / K : equation 1 
where: 
Xoc=oil compression compensation. 
Fant=anticipated rolling force. 
Fact=actual rolling force. 
K=constant for the actuator compliance. 
According to equation 1 the oil compression compensation factor is equal to 
the anticipated rolling force minus the actual rolling force divided by 
the actuator compliance. In referring to FIG. 1, the Fant value is 
produced by the adjusting device 28 which value is employed by the summing 
amplifier 26 and the function generator 29 to produce the Xoc value i.e. 
the function generator 29 in combination with the summing amplifier 26 
solves equation 1. 
The second equation is expressed as: 
EQU G=H-Fant/M-Xoc : equation 2 
where the new terms have the following meaning: 
G=desired gauge to be rolled. 
H=no load roll setting. 
M=constant factor for the mill modulus. 
The function generator 24 produces two alternative signals i.e. Fant/M or 
Fact/M by multiplying the rolling force signals by the constant 1/M 
representing the mill stretch factor. Thus function generator 24 produces 
the factor Fant/M of equation 2 employed for the pre-setting roll gap 
operation. 
The position control amplifier 30 thus solves equation 2 and in cooperation 
with the power amplifier 36 controls the operation of the servovalve 38 to 
set i.e. pre set the initial no load roll gap setting of the mill 10, 
which setting will have two negative values, one representing a computed 
value for the mill stretch for an anticipated rolling force and the other 
representing a computed value for the oil compression compensation for the 
same anticipated rolling force. Without requiring any further operation of 
the hydraulic components to compensate for compression of the oil and of 
equal importance requiring no corrective operation during rolling by such 
components the present invention effects a compensation quickly and 
efficiently by the simple operation of the summing amplifier 26 and the 
function generator 29. Accordingly, when the strip enters the mill bite 
and the rolling force begins to build up, the computed value Xoc will be 
progressively reduced by the reference signal equal to the amount of the 
actual compression of the oil as computed by the units 26 and 29. In 
effect the signal from the position transducer 18 which measures the 
movement of the piston of the actuator 16 and produces a signal requiring 
corrective operation on the part of the control amplifier 30 and 
servovalve 38 is progressively negated by an amount representing the 
progressive increase in compression of the oil as a function of the 
increases in the rolling force. As a result, the position control 
amplifier 30 is not called on to effect an operation of the actuator 16 
thereby eliminating all of the attendant disadvantages. 
In certain applications, the value Xoc during the initial rolling phase 
will be preferably produced between the time period when the strip begins 
to enter the roll bite until it is fully in the roll bite, for example a 
period of approximately 16 milliseconds. Also the Xoc value can be 
recalculated or updated periodically, in the case of the above example 
approximately every 4 milliseconds. 
The following comments have reference to the derivation of the actuator 
compliance (stiffness) used in the oil compression compensation equation. 
Actuator stiffness is a function of the actuator stroke as can be seen in 
the following equation: 
EQU K=(2.times.A.sup.2 .times.B ) / (V+A.times.S ) : equation 3 
The stiffness equation represents two single acting actuators acting in 
parallel where: 
S=actuator stroke, in. 
A=actuator area, in.sup.2. 
V=volume of oil in the line between valve and actuator, in.sup.3. 
B=oil modulus, psi. 
From the oil compression compensation equation: 
EQU S=Z-H+F / M+F / K : equation 4 
where: 
Z=actuator stroke when mill is "zeroed". 
H=strip thickness. 
F=anticipated force. 
M=mill modulus. 
Substituting equation 4 into equation 3 and rearranging yields: 
##EQU1## 
If the valve is mounted on the actuator, then 
EQU A.times.(M.times.(Z-H )+F )&gt;&gt;V.times.M 
and the immediately above equation reduces to: 
##EQU2## 
Relating the immediately above equation to an example, assumed in which 
the following is: 
actuator diameter=38 inch. 
actuator stroke at mill "zero", Z=5.00 inch. 
mill modulus, M=30,000,000 lbs/inch. 
oil modulus, B=180,000 lbs/in.sup.2. 
anticipated force, Fant=8,000,000 lbs. 
product thickness, H=0.5 inch. 
the servovalve is mounted on the actuator, V=0. 
actuator area, A=1134.1 in.sup.2. 
actuator stiffness, K=84,000,000 lbs/inch. 
oil compression factor, Xoc=(Fant-Fact ) / K equation 1. 
when Fact=0 , Xoc=0.095 inches. 
The initial gap reference, G is: 
EQU G=H-Fant/M-Xoc equation 2=0.138 inches. 
The following chart shows the changes in the roll gap and the actuator 
reference as the strip enters the mill and the rolling force, (Fact) 
increases up to the point where the strip is completely in the mill 
employing an eight discrete unit updating program. 
______________________________________ 
Fact Fact/M Fact/K Actual Gap 
Actuator Ref. 
______________________________________ 
0 * 10.sup.6 
.000 .000 .138 4.862 
1 * 10.sup.6 
.033 .012 .183 4.850 
2 * 10.sup.6 
.067 .024 .229 4.838 
3 * 10.sup.6 
.100 .036 .274 4.826 
4 * 10.sup.6 
.133 .048 .319 4.815 
5 * 10.sup.6 
.167 .060 .365 4.802 
6 * 10.sup.6 
.200 .071 .409 4.791 
7 * 10.sup.6 
.233 .083 .454 4.779 
8 * 10.sup.6 
.267 .095 .500 4.767 
______________________________________ 
From the chart it can be readily observed the results of the employment of 
the present invention by noting, (1) actuator ref. 4.862=0.5000-0.138 i.e. 
(Z-G ), (2) that the changes in the actuator reference value of 4.767 
equals the actuator reference 4.862-0.095, the Fact/K value, and (3) that 
the only correction required to be made by the actuator is the 0.267 
Fact/M value since the compression of the oil is self compensated for, it 
being appreciated that the Fact/M term is compensated for in the 
gauge-meter equation. i.e. equation 2. 
The immediately above discussion may be better understood by referring to 
FIG. 2. As noted the curve 2 plots time against changes in the roll gap 
during the initial rolling period between the time period when the strip 
begins to enter the roll bite until it is fully in the roll bite where 
there is no compensation for oil compressibility as practiced by the 
present invention. Without such compensation the servovalve attempts to 
effect a compensation by adding oil to the actuator therefore performance 
is based on the operation of the servovalve and the response time of the 
system as portrayed by curve 2 in FIG. 2. 
In counterdistinction, curve 1, representing the present invention, the 
initial roll gap includes a pre-set precalculated compensation value for 
the compressions of the oil (Xoc), in which the curve illustrates three 
important achieved advantages, (1) the substantial reduction in time to 
effect compensation for oil compression and hence reduction in attendant 
off gauge rolled product, (2) reduction in the time the maximum adverse 
roll gap change is reached also reducing off gauge product and (3) 
requires no need to add oil to the actuator and therefore performance is 
not based on the operation of the servovalve or system response time. 
It will be appreciated by those skilled in the art that the principles of 
the present invention can be practiced in many other applications and 
forms than disclosed and may be modified to adapt to such applications and 
forms without departing from the scope thereof.