Stress regulator for pulp grinding apparatus and method

The present invention provides a regulator for a refiner or grinding apparatus having opposed relatively rotatable grinding discs. Each grinding disc carries a grinding segment, and a predetermined gap or grinding space is defined between the grinding segments. The regulator, which is provided to compensate for axial variations in the predetermined grinding space due to variations of the load, includes a piston axially movable in a cylinder and a calibrated resilient element acting on the piston. The regulator is coupled to a hydraulic servo motor for the grinding discs, and the dimensions of the piston and cylinder of the regulator are selected to correspond to that of the servo motor, but in reduced scale. By calibrating the resilient element of the regulator to correspond to the axial load applied to the grinding discs, the displacement of the regulator piston corresponds to the relative displacement of the grinding disc and thus to a variation in the grinding space between the grinding discs. Variations in the grinding space are instantaneously and continuously counteracted by the regulator to maintain the grinding space at its predetermined value.

BACKGROUND OF THE INVENTION 
The present invention relates to grinding apparatus with grinding discs 
which rotate relative to one another, defining therebetween a grinding 
space in which the material is ground under atmospheric or 
superatmospheric pressure and under corresponding temperature. The 
grinding discs are supported by an axially displaceable shaft or stator 
disc for adjustment of the spacing between the discs and which axial 
displacement is controlled by means of one or more servo motors. The 
grinding apparatus is principally intended for grinding 
lignocellulose-containing material in the form of chips or fiber products. 
In order to achieve optimal grinding results, it is of great importance 
that the predetermined distance between the grinding discs is maintained 
constant during the grinding process, even in the event of variations in 
the amount of material to be ground. For example, the axial load on the 
grinding members can vary, for example, from zero tons at stoppage of the 
supply of material to 100 tons at full load of 25,000 kw. 
During these axial load variations and with a fixed distance between the 
grinding elements in the range of 0.05 to 0.2 mm, depending upon the 
desired grinding result, it will be understood that extensions and 
retractions of the machine components which support the grinding elements 
may cause variations in the grinding space which exceed the pre-set value. 
Variations in the grinding space defined between the two grinding elements 
under normal operation of a refiner are substantially linear with the 
axial load to which the grinding elements are subject. 
This means that the space between the grinding elements can not be adjusted 
to the desired value during idling, but must be adjusted to the desired 
value during the actual grinding operation and at each change in the load 
factor. 
In the event of sudden interruption in supply of material, the axial load 
is reduced to zero, with consequent neutralization of the extensions and 
retractions which are desired to be maintained in the apparatus, causing 
the spacing between the grinding elements to be immediately decreased to a 
degree where there will be frictional contact between the grinding discs. 
Such frictional contact at a rotational speed on the order of 1,000 to 
3,600 r.p.m. will cuase an immediate dry generated temperature increase up 
to the melting point of the grinding elements, with consequent destruction 
of the apparatus. 
Several methods have been used heretofore in an attempt to prevent such 
destruction of the grinding elements. An example of these heretofore known 
methods is a load or feed sensor means which, for example, at decreased 
material supply or load, returns the grinding elements mechanically or 
hydraulically to a pre-determined position free of contact between the 
grinding discs. Several such systems are described in Swedish Patent No. 
214,707, corresponding to U.S. Pat. No. 3,212,721 and in Swedish Patent 
No. 395,372 and corresponding U.S. Pat. No. 4,073,442 which describe 
single disc refiners having sensor means including an electrically 
controlled extension metering system or an electrically controlled 
resistance measuring system, or a mechanically controlled sensor or a 
hydraulically actuated wedge shaped member by means of which the spacing 
between the grinding elements is controlled. 
The sensor means of the aforementioned prior art, although useful, may 
exhibit certain disadvantages. The electrical metering systems may not 
react in sufficient time to prevent contact between grinding elements in 
the event of a sudden interruption of supply material which unexpectedly 
reduces the axial load to zero, particularly where the refiner is already 
operating at relatively small pre-set grinding space between the grinding 
elements. As discussed in U.S. Pat. No. 4,073,442, the electrical sensor 
means first separate the grinding elements only after initial metallic 
contact between the two occurs. Additionally, the use of an elecrical 
metering increases the overall cost of the refiner apparatus as a result 
of the necessary electrical components and labor required to install the 
same, and increases the possibility of malfunction of the apparatus as a 
result of failure of the electrical sensor system. 
The use of mechanical control means, such as the wedge-shaped element 92 
described in U.S. Pat. No. 3,212,721, provides mechanical control means to 
displace a piston a distance corresponding to the relative displacement of 
the grinding elements which displacement corresponds to variations or 
deviations of the grinding space from its preset value. However, movement 
of the wedge element 92 is subject to frictional and inertial constraints. 
Accordingly, in operation, the displacement of the piston controlled by 
the wedge may not be of a continuous and dynamic nature, may not precisely 
correspond to the actual variation from the preset grinding space between 
grinding elements of the refiner, and may not react quickly enough to 
cause the necessary corrective action to be taken to return the grinding 
space to its preset value. 
SUMMARY OF THE INVENTION 
According to the present invention, which is applicable to single disc 
refiners as well as double disc refiners, the regulator means which 
actuates the hydraulic adjusting means has been replaced by a 
hydraulically actuated axially displaceable piston having the same area 
relationship as on the hydraulic servo motor for the grinding elements. 
The piston works against a resilient element which is calibrated to 
operate with the same resilient constant as the sum of the extension and 
retraction of the axially loaded components of the grinding apparatus. 
By hydraulically coupling this piston to the hydraulic servo motor for the 
grinding means, a variation in the length of this regulator system is 
achieved which corresponds fully to the extensions in the different 
apparatus components in the refiner. This regulator can be placed between 
the adjusting means for the refiner and its set-screw or between the 
adjusting means and its mounting in the apparatus frame, and will thereby 
always control the adjusting means by its variations in extension so that 
the servo motor piston will always be displaced in proportion to the 
changes in extensions arising in the apparatus and thereby will always 
maintain the set distance between the grinding elements, regardless of 
variations in axial load. 
The regulator of the present invention may be employed in grinding 
apparatus having only one rotatable grinding element and one stationary 
grinding element, or in apparatus having two opposed rotatable grinding 
elements. 
By application of the above described control technology to grinding 
apparatus having two relatively rotating grinding elements, both of which 
are controlled by a hydraulic servo motor, both rotating discs can be 
provided with separate control means according to above, or only one of 
the rotating discs and in which the resilient device included in the 
system is calibrated and constructed to correspond to the total extension 
in both sides of the apparatus. The same holds true if only one of the 
rotating discs is controlled by a hydraulic servo motor while the opposing 
disc is anchored mechanically. Other details and advantages of the 
invention will be described in conjunction with the following description 
of the drawings.

DESCRIPTION OF THE BEST MODE FOR CARRYING OUT THE INVENTION 
Referring to FIG. 1 of the drawing, reference numeral 110 generally 
illustrates the control or regulator means of the present invention. The 
regulator 110 includes an hydraulic axially displaceable piston 100 
enclosed in a cylindrical housing 102 forming two pressure chambers 105 
and 107 with supply openings 104, 106 for the hydraulic pressure medium 
from a servo motor (described below) by means of which the piston 100 can 
be forced against a resilient element, as, for example, a spring device 
108 enclosed in the cylindrical housing 102. The spring device is formed 
by calibrated spring plates having a resiliency constant proportional to 
the piston area of the control device 110 and respectively to the 
regulated combined extension and retraction of the axially loaded 
components of the grinding apparatus, as will be more fully described 
herein. The control means 110 is journalled in bearing 114,which is 
mounted by means of a console 48 to a housing for the servo motor of the 
grinding apparatus. 
FIG. 2 shows a single disc refiner with the control device 110 placed 
between the control means for the refiner and its set-screw. The apparatus 
comprises a frame 10 in which shaft 12 is journalled into bearings 14, 16. 
The bearing 14 is housed within an inner bearing housing 18 and together 
with the latter is axially displaceable within an outer bearing housing 
20. 
In the same manner, the bearing 16, which is a combined axial and radial 
thrust bearing, is axially displaceable together with an inner bearing 
casing 22 within an outer bearing casing 24. The shaft 12 carries a rotor 
26, onto which a grinding disc 28 is rigidly secured and thus is rotated 
together with the shaft. A stator 31 carrying a stationary grinding disc 
30 is fastened by means of bolts to a casing 32, divided at a horizontal 
level above the shaft. The material to be ground is fed into the apparatus 
through a central channel 34 formed in the casing 32 and conveyed in an 
outward direction between the grinding discs 28 and 30, where it is 
disintegrated. Disposed in the base part of the casing 32 is a discharge 
opening 36 for removal of the ground fibrous material. 
A hydraulic servomotor, generally designated by reference numeral 38, is 
provided around the shaft 12. The servomotor comprises a casing 40 which 
may be integrally formed with the bearing casing 24, and a piston 42, 
which is concentric with and, with play, surrounds the shaft 12 and bears 
against the inner casing 22. The piston 42 has a central flange 44, 
axially movable within the casing 40. 
A positive pressure chamber 68 is defined on the left hand side of the 
servomotor flange 44, while a negative pressure chamber 69 is defined on 
the right hand side of the servomotor glange 44, as shown in FIG. 2. The 
expression "positive" means that in chamber 68 a hydraulic pressure is 
maintained which generates an axial pressure force component which is 
directed towards the stationary grinding disc 30. The expression 
"negative" means that in chamber 69 a hydraulic pressure is maintained 
which generates a force component acting in the direction opposite to that 
of the positive pressure. 
The axial movement of a servomotor 38 is achieved and controlled by means 
of a pilot valve 45 and the extension regulator 110 or control means 
operatively coupled thereto. The pilot valve 45 is fixedly mounted to the 
servomotor housing 40, 24 by means of the console 48, while the extension 
regulator at control means 110 is journalled in a bearing 114 for axial 
displacement between the pilot valve 45 and a set screw 76. The set screw 
76 is supported by a bracket 74 fixedly mounted to the servomotor piston 
42 and is thus displaced along with the axial displacements of the 
servomotor piston 42 and shaft 12. 
The servomotor 38, via the pilot valve 45 and the extension regulator 110 
controls the predetermined spacing between the grinding discs 28, 30 and 
thus during the passage of the ground material through the grinding space 
between the discs counteracts the axial forces generated. The 
counteracting forces are generated by means of a hydraulic pressure medium 
which is supplied to a central chamber 52 from an oil sump 60 by means of 
pump 62 and conduit 65. 
The pump 62 is controlled by a spring loaded valve 64. The central chamber 
52 is located between a pressure chamber 56 (positive) and a pressure 
chamber 54 (negative). A conduit 67 connects the positive pressure chamber 
56 and the positive pressure chamber 105 of the extension regulator or 
control means 110 to the positive pressure side 68 of the servomotor. A 
conduit 66 connects the negative pressure chamber 54 of the pilot valve 45 
and the negative pressure chamber 107 of the extension regulator 110 to 
the negative pressure side 69 of the servomotor. 
Operation of the apparatus illustrated by FIG. 2, including the control 
means illustrated by FIG. 1, which is employed in the refiner of FIG. 2, 
will now be described as follows. It is initially noted that the structure 
and operation of a basic disc refiner such as that illustrated by FIG. 2, 
but without the control means 110, is fully described in U.S. Pat. No. 
3,212,721, issued Oct. 19, 1965, and U.S. Pat. No. 4,073,442, issued on 
Feb. 14, 1978. The disclosure in each of these two patents is expressly 
incorporated by reference herein for the purpose of further illustrating 
the structure and operation of basic refiners of the type with which the 
control means of the present invention is used. 
Referring again to FIG. 2, oil of constant pressure is supplied from the 
pump 62 to the central chamber 52 of the pilot valve 45 through conduit 
65. FIG. 2 shows the piston 46 in a neutral middle position in which the 
hydraulic pressure medium is distributed equally to the chambers 54 and 56 
so that the same pressure will prevail in these two chambers as well as in 
the two chambers 69 and 68 in the servomotor 38. If the piston 46 should 
now move to the left in FIG. 2, the pressure in the space 56 will 
increase, while the pressure in the space 54 will decrease. This is due to 
the fact that the middle flange 47 opens up a greater connecting area 
between the central chamber 52 and the space 56, while at the same time, 
the area between the chamber 54 and the central chamber 52 is reduced. 
Consequently, a higher pressure will act on the piston flange 44 of the 
servomotor in the positive chamber 68 than in the chamber 69. If the 
piston 46 moves in the opposite direction, the result will be the reverse, 
i.e., the pressure in the servomotor chamber 69 will increase, and the 
pressure in chamber 68 will decrease. The material fed between the 
grinding discs 28 and 30 is thus subjected to a pressure, the magnitude of 
which depends upon the position of the piston 46 of the pilot valve and 
which is adjusted by the set screw 76 via the extension regulator 110. 
The piston 46 is pressed constantly against the extension regulator 110 and 
the set screw 76 by a spring 55 of the pilot valve 45. Thus, the piston 46 
follows the set screw as it moves in an axial direction. If the pressure 
between the grinding discs 28 and 30 increases, due to the accumulation of 
grist in the grinding space between the grinding discs, with consequent 
displacement of the rotating grinding disc 28 and servomotor piston 42 
towards the left, the set screw 76 will move a corresponding distance in 
the same directions, since it is fixed to the bracket 74. The piston 46 
will be similarly displaced under the resilient pressure exerted by the 
spring 55. During this displacement of the piston 46, the hydraulic 
pressure will increase in the pressure chamber 56, and, consequently, in 
the chamber 68 in the servomotor. Conversely, the hydraulic pressure in 
the pressure chamber 69 in the servomotor will decrease a corresponding 
degree, The increased pressure generates a counteracting force on the 
servomotor piston 42, in order to return the rotating grinding disc to its 
original position, and thus to restore the grinding space to its 
predetermined width. The grinding space should have a width preferably in 
the range of 0.01 mm, and 0.2 mm, depending upon the type of material to 
be refined. 
On the other hand, in the event of interruption of feed of grist, the 
grinding discs will move towards one another as a result of the decreased 
load. The servomotor piston 42 and the set screw 76 will follow, causing 
the piston 46 of the pilot valve 45 to move toward the right. This 
movement of the piston 46 will in turn cause an increase in pressure in 
the pressure chamber 54 as well as in the servomotor chamber 69, and, 
conversely, a decrease in pressure in the chamber 56 and the servomotor 
chamber 68. By adjusting the set screw 76, the desired space between the 
grinding discs 28,29 can be increased or decreased. Therefore, the 
servomotor piston and the pilot valve are alternately actuated in response 
to momentary variations in the grinding space. As illustrated in FIG. 2, 
the piston 100 of the regulator 110 is in axial alignement with the piston 
46 of the pilot valve 45. As also shown in FIG. 2, the forward end of the 
regulator piston 100 abuts directly against the opposed forward end of the 
pilot valve piston 46. As a result of this relationship, the pistons 46 
and 100 are conjointly linearly movable or displaceable in the same 
direction along a common plane. 
The extension regulator 110 shown in FIG. 1 is axially displaceable and 
located between the set screw 76 and the piston 46, as shown in FIG. 2. 
The regulator 110 is designed to compensate by changes in its longitudinal 
extension for stresses which are generated in the machine components which 
transmit the axial loads (grinding pressures) from the grinding members 
28, 30 to the servomotor piston 44. The piston area of the stress 
regulator 110 has the same relationship as that of the servomotor, but in 
reduced scale. For example, if a hydraulic force of 10 tons is exerted on 
the piston of the servomotor in a certain direction, the regulator 110 may 
be designed so that 1/10th of the force (i.e., 1 ton of hydraulic 
pressure) is applied in the same direction as the regulator piston. The 
spring device 108 which forms part of the stress regulator 110, and 
against which the piston 100 abuts, is calibrated according to the 
regulator's piston area and to the elasticity or displacement of the 
apparatus during axial loads to produce an axial change in length of the 
stress regulator so as to counteract entirely the elastic stress changes 
at each load or stress level in the apparatus. That is, as a result of the 
relationship of the calibrated resilient element 108 to the 
pressure/surface area of the regulator 110 (which corresponds in scale to 
the displacement of the servomotor piston), the displacement of the 
regulator piston corresponds precisely to the displacement of the shaft 12 
and the resulting deviation of the grinding space width from the preset 
value. 
The position of the servomotor piston is adjusted by variation in the 
spacing between the end surface of the set screw 76 and the piston 46 of 
the pilot valve. If this spacing should be changed by changing the 
position of the set screw or by hydraulic adjustment of the total length 
of the stress regulator, the servomotor piston 44 and the shaft 12 are 
displaced to a corresponding degree. 
By intercoupling the chamber 105 of the stress regulator 110 with the 
positive pressure chamber 56 of the pilot valve and the positive pressure 
chamber 68 of the servomotor and the chamber 106 of the stress regulator 
with the negative pressure chamber 54 of the pilot valve and the negative 
pressure chamber of the servomotor, the piston 100 of the stress regulator 
110 is thus loaded, with the resultant load being the net pressure of the 
two pressure chambers 105 and 107. 
This resultant force is entirely proportional to the axial pressure 
components applied to the servomotor piston 44, i.e., of the grinding 
members 28 and 30 and the axial force components generated by the 
superatmospheric pressure in the refiner housing 32, and thus causes an 
extension or retraction at each load moment of the spacing between the set 
screw 76 and the piston 46 of the pilot valve, which in turn causes the 
pilot valve to automatically adjust the servomotor piston so that 
variations in the spacing between the grinding members 28, 30 will be 
constantly counteracted and entirely eliminated, i, e., the predetermined 
grinding space will be maintained constant regardless of load variation. 
It is therefore apparant that the regulator control means 110 of the 
present invention assures continuous and precise control and compensation 
for variations and deviations of the grinding space between the refiner 
discs from its preset value. This is accomplished by design of the 
pressure/surface area ratio of the regulator to correspond (in scale) to 
the pressure exerted on the grinding discs, and by calibrating the 
resilient element so that displacement of the regulator piston corresponds 
to displacement of the shaft carrying the rotatable grinding disc. The use 
of a resilient element, such as the calibrated regulator spring 108, 
eliminates the problems of frictional resistance encountered by the 
aforementioned mechanical wedge regulator elements to provide more precise 
and continuous monitoring and control of shaft displacement, and decrease 
the reaction time of the regulator to compensate for shaft displacement 
resulting from variations of the load. Preferably, the regulator has 
polished surfaces to further reduce any frictional resistance to the 
movement of the regulator spring 108 and the regulator piston 100. 
Although the preferred embodiment of the invention employs a calibrated 
spring 108 as the resilient element of the regulator 110, other calibrated 
resilient elements may also be employed. 
By hydraulically coupling the regulator 110 to the pilot valve 45, the 
pilot valve operates, by hydraulic forces as described above, to stop 
displacement of the disc shaft within about a 0.01 mm movement. 
Displacement of the shaft simultaneously actuates the regulator element, 
which is hydraulically coupled to the pilot valve, to return the displaced 
shaft to its original preset position. The cooperation between the 
regulator and pilot valve results in an automatic, immediate and precise 
response to deviations in the preset grinding space and provides the 
immediate corrective action necessary to restore the grinding space to its 
preset value. Immediate, precise and automatic response is critical to 
preventing destruction of the grinding elements when a sudden and 
unexpected interruption in feed of material occurs, as previously 
discussed. 
FIG. 4 of the drawings is a chart illustrating how the resilient element or 
spring 108 of the regulator 110 is calibrated. The chart compares the 
axial load applied to the rotatable disc of a refiner, such as that 
illustrated by FIG. 2, to the displacement of the disc resultant from the 
applied axial load, without the regulator of the present invention. The 
chart also compares displacement of the resilient element to corresponding 
axial load on the disc to determine the corresponding values of spring 
displacement to shaft displacement. Using the test data, the proper 
calibrated values of the spring 108 may be determined so that spring 
displacement corresponds to disc displacement. As an example, the graph of 
FIG. 4 illustrates that a spring which is displaced about 0.03 mm at an 
applied axial load of 100 tons on the disc of the refiner corresponds to a 
disc displacement of about 0.26 mm. As apparent from the graph, both 
spring and disc displacement vary linearly with applied load. 
FIG. 3 of the drawings illustrates a double rotatable disc refiner 
employing the regulator means of the present invention. The basic 
structure and operation of the refiner illustrated by FIG. 3 is 
substantially similar to that of the refiner illustrated by FIG. 2, and 
corresponding reference numerals have been used in FIG. 3 where 
applicable. The basic difference between the refiners illustrated by FIGS. 
2 and 3 is that the FIG. 3 refiner is a double disc refiner in which the 
grinding segment 30 is mounted to a rotor 140, and not a stator. The rotor 
140 is mounted to and rotatable with a shaft 142. A servo motor 38 is 
operatively associated with each of the shaft 12 for monitoring and 
controlling the displacement of the shaft. Material to be refined is 
introduced into the grinding space between the two counter rotating 
grinding elements through a chute 144. In a double rotating refiner 
apparatus, the opposed grinding surfaces rotate in opposite directions, 
and the shafts carrying each of the rotatable grinding surfaces are 
individually axially displaceable as a result to the load between the 
grinding segments. An example of the basic operation of a double rotating 
disc refiner is described in my U.S. Pat. No. 4,378,092, the disclosure of 
which is incorporated by reference herein. 
Still referring to FIG. 3, a sump, a pump, a pilot valve, a regulator 
device in accordance with the present invention and a set screw, are 
coupled only to the servo mechanism 38 which controls the displacement of 
the shaft 12. The operation of the regulator 110 is identical to that 
described with respect to FIG. 2, except that the resilient element or 
spring 108 of the regulator for the refiner of FIG. 3 is calibrated to 
compensate for twice the displacement of the shaft 12 and the grinding 
element carried by that shaft. As a result of the load in the grinding 
space between the two grinding elements, displacement of the shaft 12 
represents only one-half of the deviation to the grinding space because a 
corresponding displacement of shaft 142 will also occur as a result of 
load variations. Accordingly, by calibrating the resilient element 108 of 
the single regulator element 110 to compensate for the displacement of 
both shafts 12 and 142, a single regulator element 110 may be used in the 
double rotating disc refiner. 
It is, of course, possible to provide each of the rotatable discs and their 
supporting shafts with regulator elements, pilot valves, and set screws. 
This embodiment of the invention is not preferred, because it requires the 
provision of duplicate elements (i.e., sump, pump, pilot valve, regulator 
element, and set screw) and will therefore increase the cost of the 
overall refiner. However, an embodiment of the invention including 
separate regulator systems for each shaft would be preferable under 
circumstances in which a refiner is designed in a manner in which the 
opposed rotatable shafts would not be displaced the same distance as a 
result of variations of the load between the grinding segments. 
Other advantages of the invention described herein will become apparent to 
those skilled in the art. Accordingly, the description of the preferred 
embodiment of the invention is intended to be illustrative only, and not 
restrictive of the scope of the invention, that scope being defined by the 
following claims and all equivalents thereto.