Note: Descriptions are shown in the official language in which they were submitted.
1;~5'~
,.
NON-LINEAR ELECTROMAGNETIC VIBR~TION DEVICE
BACKGROUND AND FIELD OF INVENTION
- The subject invention relates to a device for gener-
ating mechanical vibrations. More particularly, it relates
5 to an electromagnetic vibration device utilizing non-linear
sectionalized waveform plate springs.
Utilization of the principle of electromagne}ic shock
excitation for designing drivers has been used in the past
in many forms of vibrating devices. Thus far, however, the
10 application of vibrating machines is still limited and
their energy saving effects and various other potential
advantages have not been fully demonstrated nor made prac-
tical in wide application.
Prior efforts in this field have largely laid emphasis
15 on structural forms and mass combinations, and two crucial
problem areas have not been seriously considered and ad-
dressed, namely: 1) the design, manufacture, and parameter
selection for the vibration component, and 2) the control
of mechanical vibration.
Mechanical vibration is a complex physical phenome-
non. Its amplitude, frequency, energy dissipation, and
limits of stored and released energy depend on numerous
factors including: a) rigidity characteristics of thè
vibration component, b) structural damping of the vibration
component, c) magnitude, frequency, and other characteris-
tics of a shock exciting force, d) magnitude and type of
external damping, and e) expected variations and changes in
load.
If the vibration component is a linear spring (with
constant rigidity), damping is directly proportional to
velocity and a linear vibration system is formed. When
such a system is applied to vibration machines, the disad-
vantages are that the limit of stored energy is relatively
- 2 - ~2~7~21
small and the stability of vibration is rather poor. For
this reason, efforts have been made in the field toward
non-linear vibration devices many of which are available
today.
Non-linear vibration systems of the prior art, how-
ever, have also been less than satisfactory. Certain non-
linear springs may stabilize a systems inherent frequency,
and achieve larger vibration amplitudes at lower cost of
power consumption, but are applicable only in ordinary
operating conditions which do no~ require higher precision
and accuracy in directional vibration, or consistent reso-
nance and vibrator stability during fluctuating load condi-
tions in mass, flexibility, and damping. Furthermore,
conventional non-linear vibration systems are not suitable
for effective operation when a vibration frequency differ-
ent from the exciting frequency is required, or when it is
used to drive a large vibration system. Also, changes in
the mass of the external load, in rigidity, in damping, or
combined random changes of varying kinds, will result in
instability of vibration. For these reasons, the range of
applications of previously known non-linear vibration de-
vices have been limited. Alternatively, complicated and
costly electrical feedback devices must be employed. Con-
sequently, a complex supplemental system is usually re-
quired for providing electrical control or other kinds ofassistance to non-linear vibrators of the prior art in
order to accommodate operating conditions of the type men-
tioned above. These complex supplemental systems invaria-
bly add to the cost of energy consumption as well as re-
stricting the practical adaptability of the vibrator todifferent applications by causing the overall structure to
be bulky and inflexible.
Although research and development of non-linear vibra-
tion systems has had a history of many years; due to the
vastness, richness, and variability of this field, a series
- 3 - 12S7~
of fundamental problems remained unsolved prior to the
instant invention, including those mentioned above. For
example, how many limiting rings of the phase space are
there in an ordinary differential system of nth order (i.e.
5 how many basically different stable vibration types are
there in a non-linear vibration system which may be
described by an ordinary differential equation of nth
order)? Of the twenty three problems proposed by Hilbert
nearly a century ago, this is the sixteenth one and the one
10 for which the least progress has been made. Although in
theory this problem has not been solved thoroughly, the in-
stant invention, in the field of non-linear electromagnetic
vibration, provides a practical solution in utilizing
selective elastic vibration components of selective para-
15 meters and rigidity in combination for effecting a con-
trolled stable vibration, in the absence of complex exter-
nal electrical supplemental feedback controller devices.
The difficulties suggested in the preceding are not
intended to be exhaustive, but rather are among many which
may tend to reduce the effectiveness and satisfaction with
prior vibration devices. Other noteworthy problems may
- also exist; however, those presented above should be suffi-
cient to demonstrate that non-linear electromagnetic vibra-
tion devices appearing in the past will admit to worthwhile
improvement.
OBJECTS OF THE INVENTION
It is therefore a general object of the invention to
provide a novel, non-linear electromagnetic vibration de--
vice which will obviate or minimize difficulties of the
type previously described.
It is a specific object of the instant invention to
; provide a non-linear electromagnetic vibration device with-
~i
1257;~2~
out requiring the use of a supplemental electrical control-
ler for regulating and converting vibration frequency.
It is another object of the invention to provide a
non-linear electromagnetic vibration device yielding con-
sistent oscillations independent of varying loads.
It is still another object of the invention to provide
a non-linear electromagnetic vibration device which is
highly efficient in energy consumption.
It is yet still another object of the invention to
provide a non-linear electromagnetic vibration device which
is highly adaptive for use in varying application environ-
ments.
It is another object of the invention to provide a
non-linear electromagnetic vibration device which will
exhibit precise linear motion without external guides or
bearings.
It is a further object of the invention to provide a
non-linear electromagnetic vibration device capable of
yielding stable and consistent oscillations.
It is yet a further object of the invention to provide
a non-linear electromagnetic vibration device capable of
consistently maintaining a harmonic resonance state.
It is still a further object of the invention to pro-
vide a non-linear electromagnetic vibration device which is
highly compact and easily mountable in varying incorporat-
ing units.
It is yet a further object of the invention to provide
a non-linear electromagnetic vibration device which is
essentially frictionless in operation.
~ 5 1257321
It is a related object of the invention to provide a
non-linear electromagnetic vibration device which is essen-
tially maintenance free and may be sealed.
It is yet still a further object of the invention to
5 effect any combination of the foregoing objects.
BRIEF SUMMARY OF A PREFERRED EMBODIMENT
A non-linear electromagnetic vibration device as dis-
closed in an embodiment of the instant invention utilizes
sectionalized waveform plate springs of various selectable
elastic curvature and overall structural configuration. A
plurality of waveform plate springs are sequentially con-
nected and rigidly fixed in respective first and second
sectionalized waveform plate spring sets. A first and
second vibrating part respectively including an electromag-
netic armature and core of the electromagnetic source arearranged for fixed connection to said first and second
spring sets such that dual cooperatively oscillating struc-
tures each of a parallelogram cross-section are formed and
are effective to provide precise guiding action during
oscillation. Furthermore, convenient mechanical frequency
conversion is made available through selective adjustments
in the respective masses of the vibrating parts, and the
number of plate springs in each sectionalized waveform
plate spring set.
BRIEF DESCRIPTION OF THE DRAWINGS
FIGURE 1 illustrates a perspective view of a non-
linear electromagnetic vibration device according to one
preferred embodiment of the instant invention.
- 6 - lZ57~2~
FIGURE 2 illustrates a section of FIGURE 1, rotated
approximately 180, and discloses an electromagnet and
armature supporting block.
FIGURE 3 illustrates a longitudinal partial cross-
sectional view taken along section line 3-3 in FIGURE 1.
FIGURE 4 illustrates an internal exploded axonometric
view of the elements shown in FIGURE 2, which has been
rotated approximately 180, with respect to the view shown
in FIGURE 2.
FIGURE 5 illustrates in an exploded and axonometric
view of the manner of connection of sectionalized waveform
plate springs.
FIGURE 6 illustrates a perspective view of an alter-
nate preferred embodiment of the instant invention.
FIGURE 7 illustrates an axonometric view of a portion
of a sectionalized waveform plate spring.
FIGUREs 8A, 8B, and 8C illustrate various shapes of
each sectionalized waveform plate springs in accordance
with the invention.
FIGUREs 9A, 9B, 9C, 9D, and 9E illustrate various
elastic waveforms for selectable edge contours of various
sectionalized waveform plate springs.
FIGURE lOA illustrates dual cooperatively oscillating
structures of a parallelogram shaped cross-section
according to the subject invention.
FIGURE lOB illustrates the directional motion of the
dual cooperatively oscillating structure shown in FIGURE
lOA.
7 1Z573~1
FIGURE llA illustrates an electrical circuit for sup-
plying working fre~uency current.
FIGURE llB, note sheet 8, illustrates the waveform of
current supplied by the circuit of FIGURE llA.
FIGURE 12 illustrates a dual vibrator push-pull reso-
nance assembly in accordance with the invention.
FIGURE 13 illustrates a vibrating sieve shaker mec-
hanism according to an embodiment of the instant invention
and including a non-linear electromagnetic vibration de-
vice.
FIGURE 14 illustrates a surface roughening mechanism
according to an embodiment of the instant invention and in-
cluding a non-linear electromagnetic vibration device.
FIGURE 15 illustrates a plate shaking, pounding mec-
hanism according to an embodiment of the instant inventionand including a non-linear electromagnetic vibration de-
vice.
DETAILED DESCRIPTION
TerminologY
Prior to providing a detailed description o~ various
preferred embodiments of the invention it may be worthwhile
to briefly establish operative terminology in connection
with the subject non-linear electromagnetic device.
In the instant invention the electromagnetic oscilla-
tion installation may be regarded as a bipolar oscillating
system. The core of the electromagnet and other structures
fixedly connected thereto in combination constitutes a
'.
'
- 8 - 1257~2~
first vibrating part of a given mass Ml. On the other
hand, the armature of the electromagnet and other struc-
tures fixedly connected thereto in combination constitutes
a second vibrating part of a given mass M2. During oscil-
5 lation, these two oscillating parts make reciprocatingmotion relative to their respective static positions prior
to such oscillations. The corresponding amplitudes of
oscillation of these vibrating parts are inversely propor-
tional to the associated given mass as given by the e~ua-
tion Al/A2 = M2/Ml wherein Al and A2 represent respectiveamplitudes and Ml and M2 represents respective masses of
the vibrating parts. Of these two vibrating parts, the one
of comparatively smaller total mass has comparatively
larger amplitude, and is referenced herein as the main
15 oscillating or main vibrating part. The other vibrating
part of larger mass and smaller oscillation amplitude is
referenced herein as the relative oscillating or relative
vibrating part. The total amplitude of the bipolar system,
A, is the sum of the main vibrating part and the relative
vibrating part amplitudes, Al + A2. If the mass ratio
M2/Ml is greater than ten (10), that is Al is 10 times A2,
then the whole electromagnetic oscillating system can be
regarded as a single proton oscillating system. In this
event, the total amplitude value A may be regarded as being
approximately equal to Al. Similarly, it is possible to
adjust the amplitude ratios of the two vibrating parts by
adjustments in their respective mass. Accordingly, depend-
ing on the manner of connection of various system compo-
nents to the electromagnet core or to the electromagnet
armature, main and relative vibrating parts of various
different mass, configuration, and oscillation amplitude
may be formed.
Structure and OPeration
Referring now to the drawings wherein like numerals
indicate like parts, certain preferred embodiments of the
-
9 ~2573~
instant invention will be described in detail. In refer-
ence to Figures 1-4, a non-linear electromagnetic vibration
device 20 is shown utilizing sectionalized bellows or wave-
form plate springs. A laminated magnetic core 6 and core
seat 7 of an electromagnet are riveted together. Core seat
7 is connected to lateral frame elements 4 through bolts
12. Lateral frames 4 are connected at respective ends via
bolts 14 through sectionali~ed bellows or waveform plate
spring sets 9 to mechanical tuning members lO and 11 which
may serve as balancing weight block type tuning elements
for tuning the resonant point of the oscillating system by
an increase or decrease in mass. In some instances a top
cover 50 may be connected to lateral frames 4 ~note FIGURE
lS). So configured, the core 6, core seat 7, frames 4,
weight block members lO and ll, and an optional top cover
(see FIGURE l~), and all connecting bolt~ and coupled por-
tions of the sectionalized bellows spring sets 9 together
constitute a first oscillation or vibration part or collec-
tive mass. Tuning members lO and ll are primarily effec-
tive near the resonance state, are mounted via screws 14 onthe outside of sectionalized waveform plate spring sets 9
and are operative to enable the entire vibration assembly
to remain persistently near a primary and/or lower order
resonance states.
An exploded view of a central area of Figures 1-3 is
particularly shown in Figure 4. Armature 5 is formed from
riveted and plied metallic material, T-shaped, and fixed to
upper support plate 56 and lower support plate 57. The
plates 56 and 57 are bolted and soldered to front and rear
support members 2 and 3 positioned adjacent opposing sides
of the magnetic core. The plates 56 and 57 extend through
a magnetic coil 8 which is operably disposed within the
magnetic core 6. The armature is externally tightly en-
cased by coil 8. This arrangement is more clearly illus-
trated in Figure 3. Front and rear supports 2 and 3 are
fixedly connected to a base plate l, and also to a portion
~^ -
- lo- 1257321
of a respective sectional spring-set 9 via bolts 14 ~note
Figure 3). So configured, armature 5, coil 8, support
members 2 and 3, base plate 1, upper and lower connecting
plates 56 and 57, all connecting bolts and portions of sec-
tional spring-sets 9 connected thereto together constitute
a second oscillation or vibration part of collective mass.
In the arrangement described above, the first vibrat-
ing part including the magnetic core has a larger mass than
the second vibrating part which includes the magnetic arma-
ture, and therefore has a smaller vibration amplitude rela-
tive thereto. Thus, the second vibrating part including
the armature, and having a larger vibration amplitude, is
referred to as the main vibrating part, and the first vi-
brating part including the core and having a lower vibra-
tion amplitude is referred to as the relative vibratingpart. The terminology as applied is dependent upon the
relative mass and vibration amplitudes of the two vibrating
parts, and may be reversed if the relative mass and ampli-
tudes are different in alternative embodiments. Different
structural connections of various components in distinct
embodiments of the instant invention may yield many combi-
nations of main and relative vibrating parts. For in-
stance, the lateral frames 4, instead of the front and rear
support plates 2 and 3, may be connected to the base plate
1 in an alternative embodiment and may or may not result in
a reversal in relative collective mass and vibration ampli-
tude. It is possible to adjust the size ratio of compara-
tive amplitudes through adjustments in mass ratio, in order
to meet the requirements of varying applications.
A vernier bolt 13, shown in Figures 2 and 3, has front
and rear sections of different screw pitch. The front
screw section is adapted to the bore 7B on armature base
seat 7, and the rear screw section is adapted to the bore
4B on a protruding central part of the bottom of frame 4.
Turning of vernier screw 13 will cause the riveted core 6
25732~
and core base seat 7 to advance or retreat relative to the
frame 4. Because the frame 4 is at a static position when
not oscillating, it does not vary in relation to the arma-
ture 5. In this manner, an adjustment via turning the
vernier bolt 13 effects an adjustment of the size of air
gap between the core 6 and the armature 5. Locking bolts
12 are also shown for fixing the core 6 to the frame 4;
they are symmetrically mounted on opposing sides of the
arrangement. Each bolt 12 passes through a bore 4A on a
corresponding side wall of frame 4, and turns into the bore
7A on core seat 7. It physically connects the frame 4 with
the core seat 7. When it is necessary to adjust the air-
gap of the electromagnet, one must first loosen locking
bolt 12. The presence of elongated bore 4A permits the
above described adjustments in the electromagnet air gap.
Upon completion of air gap adjustments, bolt 12 is re-
tightened in order to fix the core 6 in its new position
relative to the armature and frame.
The respective manner of connection between the main
and relative vibrating parts and the sectional band spring-
sets 9 is illustrated in detail at Figure 5. Nuts and
bolts 52 and 14 fixedly couple respective mid portions of
each sectional band spring set to the main vibrating part,
and respective end portions of each sectional band spring
set to the relative vibrating part. In alternative embodi-
ments, this connection may be reversed such that the mid
portion of the plate spring sets is connected to the rela-
tive vibrating part and the end portions to the main vi-
brating part. In this Figure, a top structure 50, as re-
ferred to above, is also shown. Each sectional band springset includes a series of sectionalized plate springs sepa-
rated by shims 17 which prevent friction during oscillating
motion of the springs. In the disclosed embodiment, three
shims are disposed between each adjacent pair of section-
- 35 alized plate springs.
12573~
- 12 -
In Figure 6, an alternative embodiment is shown where-
in a second sectional band spring set 53 is attached to
each original spring set 9 on each side of the assembly.
Separators 54 are respectively disposed between spring sets
9 and 53. In addition, screws 55 are shown, which may be
used to fix the frame 4, instead of end pieces 2 and 3, to
bed plate 1.
The construction of sectionalized bellows (or wave-
form) plate springs 17 is illustrated in Figure 5, and also
in Figure 7 which shows a half portion of a single section-
alized plate spring. On the center and two sides of each
plate spring 17, a respective mounting plane 15 having two
tap-holes thereon is located (Figure 7 shows only half a
spring). In each tap-hole, a stud screw 14 is placed for
means of coupling. Successive plate springs are pressed
together via stud screws, front and rear support members 2
and 3, and press bars 51. Intervening washers or shims may
be disposed in between one plate spring and the next. In
the embodiments described above, the mounting planes at the
ends of each plate spring are coupled to opposing side
walls of the frame 4, and a central mounting plane is
coupled to a corresponding front or rear support member 2
or 3.
The sectionalized bellows ~waveform) plate spring sets
are non-linear elastic components. The non-linear charac-
teristics may be conveniently selected and designed to suit
different application environments. Figures 8A, 8B, and 8C
illustrate a half portion of several alternative designs of
such plate springs. As can be seen, the spring may have an
edge which is straight or curved to be concave, or convex
in shape. Each spring set may include a combination of
such plate sprir.gs successively arranged as presented
above. Figures 9A, 9B, 9C, 9D, and 9E illustrate respec-
tive edge diagrams of such plate springs. These edges may
be made to satisfy various sine, exponent, or other func-
- 13 - ~ ~57~
tional curves which reflect a corresponding elastic behav-
ior of the spring. In the overall vibration assembly, each
waveform plate spring of one configuration may be used in
combination with waveform plate springs of the same or
other kinds to suit the particular reguirements of differ-
ent application environments. These sectionalized plate
springs have smooth non-linear characteristics and a wide
range of potential behavior in combination to support a
correspondingly wide range of desirable flexibility and
rigidity. There is very minimal internal damping, and a
very long fatigue life. The plate springs provide high
precision in directional motion, and high adjustability for
self-adapting control. In all these aspects the subject
invention manifests optimal performance in comparison with
common linear springs, sectionally linear composite
springs, and other existing non-linear springs.
The number of waveform plate springs in each spring
set 9 may be varied and selectively adjusted to vary the
vibration frequency responsive to an exciting frequency
from a source. If the number of plate springs is raised to
equal the square of an integral times the number required
for a harmonic response, the vibration frequency may be
multiplied a corresponding integral times. Similarly, if
the number of plate springs is made to equal the square of
an inverse integral times of the number required for a
harmonic response, then the vibration frequency can be made
to equal an inverse integral times of the exciting fre-
quency. In this manner, frequency conversion and multipli-
cation under harmonic resonance can be achieved via mechan-
ical adjustments of the number of plate springs. The needfor a complex electronic frequency converter is completely
eliminated. With the subject vibration assembly, the reso-
nance state can be persistently maintained under no load,
or varying rated load conditions. The concept is based on
the unique characteristic of strong non-linear systems in
being able to generate harmonic resonance. The spring-set
- 14 - 12573~
stiffness may be varied by varying the number of plate
springs, and thus can be made to e~ual the square of an
integral or of an inverse integral times the degree of
stiffness under harmonic response. The inherent vibration
frequency may be raised an integral times or an inverse
integral times by combined adjustments in the mass of vi-
brating parts and the vibration stiffness. ~ccordingly,
the subject vibration assembly is able to achieve harmonic
resonance characteristics in mechanical vibration as in
non-linear electromagnetic wave vibration. The instant
device enables frequency multiplication and frequency con-
version of harmonic response to be achieved for resonance
type mechanical vibration without requiring an independent
electronic frequency converter device.
When the vibration device is activated, and electrical
exciting current flows through coil 8 (note again FIGURE
3), pulsed attraction is generated by electromagnetic
action. The spring-sets are correspondingly elastically
flexed in unison towards a common direction along a linear
axis parallel to the longitudinal axis of the upper con-
necting plate 56. Because the springs are sectionalized,
and because opposing ends of each section of each plate
spring are couple to different vibrating parts, main and
relative, as discussed above, two self-reciprocating par-
allelograms are formed which share a major side along anaxis parallel to the central longitudinal axis of the upper
connecting plate 56. Each side wall of frame 4 constitutes
the other major side of each parallelogram, and a corre-
sponding section of each plate spring collectively forms
every minor side of each of said two parallelograms. As
shown in Figure lOA, parallelogram A is reciprocating in
direction of arrows A' while parallelogram B is reciprocat-
ing in unison in the direction of arrows B', and both par-
allelograms share a common edge along an axis parallel to
the central longitudinal axis of the top plate 56. This
arrangement is more particularly illustrated in Figure lOB,
~257321
- 15 -
wherein the parallelograms A and B are schematically shown
apart from the underlying structure, and cooperatingly
reciprocate in unison. The dashed lines illustrate succes-
sive positions of the parallelograms during oscillation.
In this manner, a highly stabilized cooperative reciprocat-
ing motion is achieved together with precise linear guiding
action between the corresponding parallelogram structures.
The cooperatingly reciprocating dual parallelogram
structure of the instant invention provides highly accurate
and precise directional reciprocating linear vibration
without use of external guiding structures. In the absence
of external guiding structure, guiding friction and wear
are eliminated. The dual reciprocating parallelogram
structure manifests a guiding influence on each other to
ensure stable cooperative oscillations. Without requiring
external guiding structures for stable and precise vibra-
tion, the instant vibration assembly has a greatly expanded
range of application and may even be enclosed within a
sealed container as no maintenance is required.
When resonance motion takes place in non-linear type
electromagnetic vibration facility of this invention due to
the in-flow of resonance exciting current, the iron core of
the electromagnet and the mass of parts rigidly connected
with it as well as the armature of electromagnet and mass
of parts rigidly connected with it exhibit a relative
motion with each other. As noted above, the amplitudes are
distributed inversely proportional to the ratio of relative
mass of the two parts.
In one embodiment of this invention, frame 4, upper
cover 50 and other members are connected with the iron core
6 of electromagnet and the whole structure is located out-
side the structure of armature of electromagnet front sup-
port seat 2, rear support seat 3 and other members. So, in
the whole vibration machinery, if the total mass of mem-
~\
- 16 - 1257~1
bers, and parts rigidly connected to the iron core of elec-
tromagnet compared with the total mass of members and parts
rigidly connected to the armature of electromagnet is
smaller, then the amplitude produced in the electromagnet
iron core and members rigidly connected with it will be
larger than that produced in electromagnet armature and
members rigidly connected to it. Thus, the former struc-
ture will become the main resonance part of the vibration
system. Such electromagnetic vibration facility may be
referred to as an external relative motion type electromag-
netic vibration facility.
When the working frequency current through half-wave
rectifier flows in the facility according to the electric
circuit shown in the example of Figure llA; the waveform of
current is shown in Figure llB. During the positive half
wave cycle, current passes through electromagnetic coil
8. Electromagnetic attraction force is produced between
electromagnet iron core 6 and magnet armature 7, thus ena-
bling the iron core 6 together with the iron core base seat
7, frame 4, upper cover and other members rigidly connected
to it to drive both ends of the front and rear sets of
bellows plate-spring groups 9 and to produce motion rela-
tive to the armature 5 which is rigidly connected on the
base-plate 1 and front and rear support seats 2, 3, thus,
reducing deformation in the bellows plate-spring group and
storing a certain amount of elastic potential energy.
During the negative half cycle, no current flows through
electromagnetic coil 8, and electromagnetic attraction
forces diminish (dissipate). The iron core system, under
the action of a recovering spring force is released from
the armature 5 and moves in an opposite direction. With
periodic motions, the main resonance part of the vibration
system produces fixed direction reciprocating vibration at
a frequency of the working frequency current. By adjusting
the air-gap adjustment bolt 13 and making the distance of
air gap between iron core 6 and armature 5 a little larger
.
- 17 - 1257321
than a half amplitude of fixed direction reciprocating
vibration, it is possible to prevent the possibility of a
collision of iron core 6 with armature 5 in motion, and
assure the production of maximum electromagnetic attraction
force. Because of the non-linear elasticity characteris-
tics of a bellows type plate-spring group 9 which may pro-
duce pulse type non-linear mo~ion characteristics, the
number of spring pieces in the bellows type plate-spring
group can be adjusted to vary the total rigidity of the
vibration system and to realize one and more than one har-
monic resonance. Utilizing characteristics and control and
adjustment operations substantially as described above it
has been determined that when the frequency of vibration
reaches 3000 times /min., the total amplitude reaches 4 to
6 mm.
Advantageously it is also possible, with the instant
invention, to vary the amplitude of the vibration by vary-
ing the input voltage.
Figure 12 illustrates a dual vibrator push-pull reso-
nance state. Two small-scale electromagnetic vibrators
each with power of 100 watts are used. Their respective
coils are fed through the same power wire to enable
synchronous vibration movement and to form a push-pull co-
operating relationship. The vibrators work in unison and
are able to drive a large working body 18 of large mass and
inertia. In this manner, the subject invention employs
small-scale vibrators in cooperation for driving a massive
working body. In so doing, the subject invention reduces
the required size of the vibrating assembly and also the
otherwise required stiffness of the working body. The dual
lO0 watt vibrators acting in unison are able to generate
the practical effects of a large S00 watt scale vibrator,
thus achieving a significant increased efficiency with
reduced power consumption. Push-pull resonance can not be
achieved with conventional revolving link exciters and
- 18 -
12573~:1
vibrators or inertia type exciters and vibrators, because
these type of devices can not be strictly synchronized.
The subject invention is able to employ a single power
source for driving a plurality of electromagnetic vibrators
so as to ensure that they function in a resonance state
with the same frequency and at the same phase.
A modified version of the external relative motion
type electromagnetic vibration facility mentioned above
exists in an internal relative motion type electromagnetic
vibration facility. If the total mass of the electromagnet
iron core and the members and parts rigidly connected to it
is compared with the total mass of electromagnet armature
and the members and parts rigidly connected to it, and the
former one is greater, then the amplitude obtained by the
iron core system will be relatively smaller than that ob-
tained by the armature system and the iron core system
becomes the relative vibration part, while the armature
system becomes the main vibrating part.
In a preferred embodiment of the instant invention a
solenoid type electromagnet is utilized. In comparison
with a common n or E type electromagnet, the solenoid type
has a larger electromagnetic attractive force and travel,
and is much easier to connect its two ends to double elas-
tic components, and to carry out position fixing and air
gap trimming. In addition, the structure of the entire
electromagnetic vibrator is made power-head-like so that it
is easy to connect to and maintain. The core, upper cover
and adjoining frame to the core, the armature, and front
and rear support blocks of the armature all can be used as
the unit for producing the principal vibration. In this
manner, the structure can be employed flexibly and mechani-
cal vibration can be sent out conveniently through attach-
ments with simple connecting parts at a number of alterna-
tive positions.
1257321
-- 19 --
A significant breakthrough in the technical field of
non-linear vibrating machines made by the present invention
lies in the fact that subject springs have been selected to
exhibit exponential/sine waveforms. A curve for shifting
its standard rigidity is determinable and the parameters
for the springs optimal working characteristics are con-
veniently selectable. This type of non-linear vibration
not only depends upon the average rigidity of the vibratory
elastic component but has also a close relationship with
its rigidity curve and structural damping. The exponen-
tial/sine waveform spring adopted in the present invention
has smaller structural damping than other springs and the
relationship between stress and strain is approximately a
logarithmic function.
Let "f" be the stress and n ~1~ the strain. Then
,R = aln~l+pf),
where ~ and ~ are parameters which can be determined by
the method of least square from experimental data. Con-
versely, ~f" can be written as an exponential function of0 the strain:
e = l+pf
.:f = ~(e~ -1).
23 Its rigidity function is found to be
o
df = le~.
dQ ~
Calculations by microcomputer show that the vibration type
non-linear spring with such rigidity function exhibits a
stable periodic vibration (under rather extensive external
; force conditions and for certain amplitudes). The computer
has also drawn out its stable limiting rings of the phase
space under various parameter conditions.
- 20 - 1257321
Even with an elastic vibration component of excellent
performance, normal operation of a vibrating machine still
cannot be fully ensured with conventional systems. This is
because external conditions may vary in different ways when
the machine is in operation. Thus, for example, when it is
used for cutting, the damping in cutting would vary with
the change of the hardness and dimension of the work piece
being cut; when it is used for conveying and sifting of
material, the quantity and variety of the material would
entail change in the vibrating mass; and when it is used in
compressing gases (vibratory compressor), which is equiva-
lent to attaching a gaseous spring, the change of rigidity
curve would arise. The change in these conditions would
affect the type of respective vibrations and may cause them
to transfer from the near resonance area to the far reso-
nance area and to come out of stable limiting rings and get
into a non-stable state, with the result that the vibrating
machine either could not operate normally or would exhibit
reduced performance during operation.
The present invention has effected three kinds of
vibration control, viz:
Damping control (stability control for varying damp-
ing);
Rigidity control (stability control for varying rigid-
ity); and
Mass control (stability control for varying mass).
The fundamental principle of control is the method of
feedback function, i.e., a preset feedback (self-compensat-
ing) function is to be formed by designing electromagnetic,
elastic, and mass damping systems. Once the working condi-
tions of a vibrating machine undergoes a change within a
- 21 - 12573~1
certain range, the feedback function would automatically
play a role to control and offset the change in external
conditions so that the machines may continue to work nor-
mally.
The feedback function is implemented by the following
procedures:
(a) Procedure of Electromagnetic Feedback
As the amplitude of vibration decreases with the in-
crease of external damping, the velocity of vibration, and,
therefore, the resistance of the coil, is reduced, leading
to the increase of the current. The result is the increase
of the amplitude. This control effect is called electro-
magnetic feedback, which is particularly manifested as the
increase of power with increase of external damping (work-
ing damping). To exploit the feedback effect to the ut-
most, the present invention has made the optimum selection
of the gap between the electromagnet and yoke and has used
a plug-in yoke so that the maximum feedback effect can be
obtained without any collision. The hard spring which the
present invention has adopted causes the vibrating body to
reduce its velocity abruptly when it reaches the limiting
position, which also contributes to the maximum exploita-
tion of this control effect. Moreover, in the design of
the coil-rectifier circuit, various forms such as full-
wave, full-wave rectification, half-wave rectification, and
"full-wave (electromagnet) x full-wave rectification
(yoke) n may be employed to meet different working require-
ments.
(b) Procedure of Preset Stress
In the course of the design, working area, i.e., near
resonance, is chosen not only according to the curve of the
elastic component, but by matching the selection of para-
- 22 - 12573~1
meters the component is made to work outside the working
area. Once either the external damping or the mass in-
creases, the elastic component would return from the out-
side to the inside of the working area. The resonance
force inside the working area would assist to compensate
for the deterioration of external conditions so that the
machines can operate normally.
(c) Procedure of Auxiliary Springs
In the design of vibrating machines, beside the elas-
tic component producing the principal vibration, severalsets of auxiliary springs may be installed to meet any
possible variation of external conditions. When external
changes interfere with the mass and rigidity, these auxil-
iary springs, would to varying degrees, play a part auto-
matically, resulting in compensatory changes in the rigid-
ity of the vibration system so that the machines can
operate normally.
The above control procedures may be used either sepa-
rately or jointly with one or more of them together. In a
specific design, differential equations may be formulated
on the basis of test data to be solved by simulation and,
if necessary, a computer may be used for optimizing operat-
ing conditions.
Bellows ~Waveform) Type Plate sPrings
Non-linear rectangular bellows type plate spring group
is a generally significant operating element of the subject
invention by which the main characteristics of the elec-
tromagnetic vibration facility and hence, those of the
vibration machinery using such facility as its power source
are determined.
1257321
-- 23 --
The main characteristic of this element is the appli-
cation of thin bellows type plate spring pieces with a
generally rectangular outer dimension, and separate segment
sine configuration or exponential function and the 6paced
lamination of such plate springs into a group. Each
bellows type plate spring piece has an assembly hole or
separate segment assembly plane, so as to use a bolt for
fixing and restraining it. Aluminum washer shims may be
used between these assembly planes, so that friction may be
avoided, when deformation between plate springs is pro-
duced.
The main characteristics of the subject bellows type
plate springs may be described in some detail as follows:
(a) Constructed by Thin, Rectangular Plate
Spring Piece
A conventional plate spring piece used in electromag-
netic vibration machinery is constructed by a thick (2mm)
rectangular plate spring piece. There is neither a bellows
configuration nor assembly hole on it. The pieces are used
in a non-fixed-restraint mode, with the plate spring lami-
nated in a group, and comprising elastic elements with
linear elastic characteristic. There were also previously
known single-piece, thick, circular plate spring pieces
with concentric circle series washer elements, forming a
non-linear characteristic or deformation with separate
segment lineacity to a definite stage and thereafter pre-
senting a non-linear characteristic.
The bellows type plate spring of the subject invention
is composed of thin, generally rectangular plate spring
pieces.
Each bellows type plate spring piece is generally 0.3-
1.5 mm thick. To compare the thin type plate-spring piece
.. \
- 24 - 1257~z~
with thick type plate-spring piece, we find that the inter-
nal damping of the former decreases exponentially, when
they are laminated in a group and used. Total damping is
still less than that of a comparatively thick type plate-
spring piece with the same rigidity, but the strength ofthe transverse beam constructed by the former increases
with the increase in total area of cross section. In com-
paring rectangular bellows type springs with disc shape
bellows type springs or with rectangular plate springs, it
is found that the internal damping determined by the con-
struction factor is also much smaller. When a generally
rectangular bellow type plate spring is in the state of
working deformation, deformation in a rectangular bellows
type plate spring through the bending or twisting action on
a vertical line of the bellows may be much higher than that
of various conventional plate springs (such as rectangular,
non-bellows or disk shape, bellows plate springs) and that
of rubber springs under the same action force; yet, in the
direction of non-working deformation, there is very large
mono-directional rigidity. Therefore, when the subject
invention is used in a vibration facility, and vibration
motion is produced, both large vibration amplitude and high
orientation is obtained.
Each generally rectangular plate spring may have a
shaped outer dimensions, such as straight, convex or con-
cave, as illustrated in Figures 8A, 8B, and 8C.
(b) Vertical Bellows With Sinusoidal Configuration
Or Exponential Function
As mentioned above, with generally rectangular bellows
type plate spring pieces, the vertical bellows has a unigue
influence upon reducing the structural internal damping and
assuring correctly oriented motion. Furthermore, there are
various kinds of cross-sections of the bellows, whatever
the overall outer dimension of the plate spring, such as
,
- 25 - ~257~2~
sinusoidal configuration type, exponential function type
and other different forms of bellows type cross-section.
These kinds of plate spring pieces have excellent longevity
in the fatigue type working state, because the deflection
5 curve is very smooth and the rate of variation may approach
an infinitesimal value.
(c) Continuous Non-Linearity of the Elastic
Characteristic Curve and the Versatile
Selectivity of the Curve
The elastic working characteristic of the vertical,
bellows type spring piece of this invention presents the
variation of continuous-non-linearity within the entire
scope of working deformation from start to end. Many kinds
of conventional non-linear springs, such as a non-cylindri-
15 cal spiral spring and so on, usually enter a region of non-
linear characteristics only after they have been deformed
to a specified extent, and some other non-linear springs,
such as linear plate springs, non-linear spiral springs and
some combinations of such springs constitute a non-linear
20 characteristic curve by exhibiting separate segment linear
characteristics. The generally rectangular, bellows plate
spring of the subject invention not only has a pronounced,
continuous, complete curve of non-linear elastic character-
istic, but also may selectively obtain many different kinds
25 of characteristic curves of non-linear elasticity through
changing the design parameters of plate-~pring, such as its
outer shape, thickness, different bellows shape cross-sec-
tion elastic waveform, depth of each bellow trough and
different techniques of heat treatment. A transistor tri-
30 ode can be used as an example to explain the above men-
tioned situation. Every kind of transistor triode has its
own amplification characteristic curve which may be selec-
ted according to a technical specification, and therefore,
may satisfy different technical requirements of usage and
- 26 - ~:573~
control. Particularly, in the state of different external
load fluctuations, if the proper curve of non-linear elas-
ticity character can be designed and selected, then proper
non-linear feed-back characteristics can be achieved to
5 maintain the stability o vibration of the entire system.
(d) Fixed Restraint
The generally rectangular, bellows plate-spring group
of the subject invention is different from conventional
plate-spring groups. It adopts an assembly mode of fixed
restraint with assembly holes which differ from that of
conventional, single, non-linear springs. In addition, it
utilizes a method of spaced lamination in a group for use,
so, springs of this invention has both the characteristic
of non-linear elasticity as well as versatile and unique
modulation characteristics. The total rigidity of the
vibration system can be modulated by increasing or reducing
the number of pieces of the bellows plate-spring sets, to
realize one order harmonic resonance or multiple order
harmonic resonances. Under the conditions of resonance,
owing to the above-mentioned characteristics of low inter-
nal damping and non-linear motion of the spring of the
instant invention, it is possible to obtain much larger
resonance amplitudes and spontaneous acceleration char-
acteristics than those of previously known linear and non-
linear electromagnetic vibration facilities. Therefore,the energy consumption per unit of vibration strength de-
creases considerably, and a novel harmonic resonance type
double frequency variation method is thus formed.
To summarize, non-linear rectangular bellows plate
spring group of this invention has main characteristic of
fixed restraint in application, such as, separate segment
sine configuration type or exponential function type
bellows plate spring with thin and rectangular outer shape
and laminating in group, therefore, its elasticity charac-
- 27 - 12S7321
teristic curve has continuous non-linearity and versatile
selectivity of characteristic curve. When it is used in
electromagnetic vibration facility or various kinds of
vibration machinery, if compared with any known linear or
non-linear spring, it may obtain the best and the most
stable performance characteristics of vibration in many
respects, such as: convenience in design, selection and
test of characteristics curve of non-linearity, minimum
internal damping, large amplitude, large spontaneous accel-
eration and high precision orientation, one or more thanone harmonic resonance, realization of double frequency
variation of excited resonance, as well as high strength
and long fatigue-resistant duration time.
Solenoid TYpe Electromagnet
A conventional electromagnetic vibration unit typical-
ly uses a shape "E" shape or ~H" sbape electromagnet. An
electromagnetic vibration facility of this invention uses a
solenoid type electromagnet wherein the armature is gener-
ally T-shaped and an electromagnetic core is generally U-
shaped which may better concentrate its electromagnetic
attraction in the magnetic field. Furthermore, the elec-
tromagnet of the instant invention is made to include
unique electromagnetic elements. Silicon steel sheet of
high silicon content is used to replace that of low silicon
content as a basic material of the electromagnet. In this
manner, larger electromagnetic attraction forces can be ob-
tained than that of a conventional unit. In addition, dual
reciprocation structures of parallelogram shaped cross-sec-
tion can be easily constructed to include both front and
rear sectionalized waveform plate spring sets on opposing
sides of the electromagnet so as to assure a high precision
fixed direction motion of the vibration system during sys-
tem operation.
~57~
- 28 -
Structural Configuration
(a) Fixed Direction
The instant invention has solved a problem of how to
assure a high precision of fixed direction motion. The
subject invention uses generally rectangular, laminated
bellows plate spring groups which exhibit an excellent pro-
perty of fixed direction deformation. Moreover the instant
structure exhibits a fixed restraint mode in the whole
structure of the electromagnetic vibration facility.
Mounting the plate-spring groups 9 in a fixed-restraint
mode on transverse beams 4 and the front-rear double group
support seats 2 and 3 which has a high degree of rigidity
forms a stable parallelogram structure which can maintain
high precision fixed direction motion, without any need for
mechanical guiding structures, such as a guideway. Since
there is no friction between a guideway structure and the
vibrating unit, the working parts in vibration motion can
achieve a comparatively high precision in movement without
friction losses, and the scope of the application of elec-
tromagnetic vibration facility may be enlarged.
(b) Fixed Restraint Mode of Non-Linear Bellows
Plate Springs
Plate spring pieces of a conventional plate spring
type electromagnetic vibration device are assembled in non-
fixed restraint mode. Several sets of clamping bolts areused to clamp the plate spring group together with washer
pieces between plate spring pieces on the electromagnetic
vibration facility. Accordingly, the performance charac-
teristics of the whole machine may be easily influenced or
disrupted by the clamping operation, deformation of the
washer pieces, temperature variation and so on. In the
instant invention there are assembly holes punched on the
assembly plane in several separate segments of the bellows
- 29 _ ~ 2 5 ~321
plate-spring element and clamping bolts are used to fit
through these holes and clamp the bellows plate-spring
group tightly to the frame and support seats of the elec-
tromagnetic vibration facility in a fixed-restraint mode.
By doing so, not only is stable vibration performance char-
acteristics assured, but motion characteristic of non-
linearity can also be assured.
(c) Micro-Adjustment Mechanism For Air-Gap in
Electromagnet
In previously known electromagnetic vibration devices,
it was necessary to adjust and fix the air gap between an
electromagnet iron core and an armature by several sets of
bolts which connect the electromagnet iron core with a
machine body. By this method, it was difficult to make
precise synchronous micro adjustments. It was inconvenient
to operate and it was easy to disruptively influence the
uniformity of the electromagnetic attraction force over the
whole air gap.
In the structural concept of the electromagnetic vi-
bration facility of this invention, a base seat 7 isriveted onto an electromagnet iron core. The base seat 7
forms a slideway with an internal wall of machine frame 4
in the direction of its width. Then, a differential motion
bolt 13, which has two segments of screw thread with the
same lead direction but with different pitch, is connected
with iron core base seat 7 and machine frame 4 respective-
ly, to permit micro-adjustment of the electromagnet air
gap. Only a differential motion bolt is needed to complete
the micro-adjustment motion which is the difference of two
different pitches for one turn of screw thread. In addi-
tion, proper fit of the slideway along its width direction
between the base seat and machine frame is provided to
assure high precision synchronous micro-adjustment, and at
.
- 30 ~ ~ 2 57;32 1
the same time, to assure the uniformity of electromagnetic
attraction force over the whole air gap.
(d) General Structure of the Dynamic Head
The non-linear electromagnetic vibration device
5 according to preferred embodiments of this invention are
provided with a versatile assembly base plate. A dynamic
head is installed in conformity with an outer dimension and
general design of an incorporating facility. Very minimal
expense and little machine processing is needed to make the
vibration facility fit for many different purposes and such
structure makes design and manufacture convenient and ena-
bles the facility to be compact in construction, small in
volume, low in energy consumption and use of materials.
(e) Simple Electrical Layout
When conventional linear type electromagnetic vibra-
tion facilities are used in various kinds of vibration
machinery, under the conditions of variable external loads,
it is typically necessary to add and use a complicated
electric control mechanism to maintain the stability of its
performance characteristics.
In the instant invention when the non-linear, electro-
magnetic vibration device is used in various kinds of vi-
bration machinery under conditions of variable external
loads, it is possible to depend on self-feedback of non-
linearity to obtain a stable performance characteristic.Therefore, from the viewpoint of structure, the electrical
layout of the instant facility is very simple in construc-
tion, and reliable in operation.
- 31 - 12~7;~1
Control and Adjustment
(a) Control and Adjustment of Performance
Characteristics of One or More Than one
Harmonic Resonance
In conventional linear and non-linear electromagnetic
vibration facilities, only the form of a base wave reso-
nance can be used for performing resonance control and
adjustment.
According to the instant invention, harmonic resonance
10 can be produced by adjusting the total rigidity of the non-
linear spring sets in the vibration facility. Moreover,
through an adjustment of the number, size, bellow shape,
and heat treatment of the springs, the total rigidity can
become the square of an integral fold or square of a
fraction of an integral fold of the predeterminable
rigidity in harmonic response. A simultaneous adjustment
of the vibrating mass and the total rigidity of the vibra-
tion system can yield a vibration frequency higher than
that in base wave resonance by an integral fold or a frac-
tion of an integral fold. Therefore, harmonic resonancecharacteristics traditionally achieved via supplemental
electrical controllers can be obtained mechanically. In
this manner, frequency variation for excited resonance
frequency can be realized without adding any frequency
variation electrical devices.
By adjusting the number of plate springs in one or
more sectionalized waveform plate spring set to egual the
square of an integral (n) times a number predeterminably
required therein for the vibration device to achieve har-
monic response, one can yield an integral (n) times thevibration frequency of the source as the operating vibra-
tion frequency of the device. Similarly, by adjusting the
number of plate springs in one or more sectionalized wave-
-` 1257321
-- 32 --
form plate spring sets to equal the square of an inverse
integral (l/n) times a number predeterminably required
therein for the vibration device to achieve harmonic re-
sponse, one can yield an inverse integral (l/n) times the
5 vibration frequency of the source as the operating vibra-
tion frequency of the device.
An example of a determination of the number of spring
plates is illustrated below:
The number of plates may be obtained by using
the following equation
(1) W (,~ ) =~)
Where W(~) is the operating frequency of the system
is the amplitude.
m is the inertia mass in operation.
K(~) is the spring characteristic.
Once X(~) has been determined, the number of plates
can be determined according to this equation.
The wave shape and the plate form are important
to the spring characteristic design. The spring
design can be solved theoretically as well as exper- -
imentally. The spring characteristic depends on the
wave shape, the number of waves, the form of the
plate and the applied constraint conditions. The
spring design must meet the following requirements:
~1) The spring characteristic must comply with
the vibration characteristic.
(2) The spring fatigue life must satisfy long-
term operating conditions.
~ 33 ~ 1257321
As the applications of springs are extensive,
springs used on various vibrating machines may be
different from each other. Various springs may be
designed qualitatively according to the following
comments to determine their parameters, then modi-
fied by routine experiments.
I. Spring Characteristic
Using the following terminology:
K( ~) is the spring characteristic
r is the radius of wave shape
3 is the plate thickness
b is the plate width
is the plate length
n is the number of waves
The spring characteristic R(~) will be
proportional to S3,
proportional to b,
inversely proportional to ~2,
inversely proportional to r2 ~using an
equivalent arc for other curves),
proportional to 1 , if the spring material has
been determined (consider every wave as a
spring, then spring with n waves is equivalent
to n shorter springs connected in series.
,25 Therefore
2 R3 Kn )
'
``~
\~
~ 34 ~ 1 2 573
II. Selection of Spring Shape
A. Wave Shape: In determining the wave shape
of a spring, a matter of main concern is
the fabricating method and its fatigue
strength. The sinusoidal and exponential
reduction curves are superior to others.
Research results of large deflection theory
indicates that all nth-order differentiable
curves and combined curves will not cause
stress concentrations, thus behave better
for shock resistance and fatigue resis-
tance. Other curve shapes can not be used,
however, that forementioned curves are
better. In cases of large changes of ac-
celeration, the forementioned curves are
greatly preferred.
B. The plate form may be tailored to comply
with the stress distribution and adjustment
of the spring characteristic, provided the
directional vibration requirement has been
met.
Referring again to the above equation (1), resonance
will take place so long as W~z~)=n W~ (where n=1,2,3...k
or n=l,l/2,1/3....1/k) with the operating frequency of
25 power supply being ~ , it is so called harmonic resonance,
which is a non-linear vibration characteristic.
In application, W(Z~) is selected as required, then
K( ~) can be determined based on the inertia mass"m", where
K(~) is the general characteristic of the springs required
by the system. It is possible to design the spring accord-
ing to K(~ ), suppose the characteristic of each spring is
R(Z~), then R( ~ )=K( ~)!n, where"n"is the number of the
plate springs.
~257~2~
- 35 -
In a non-linear electromagnetic vibration sieve for
example (to be described in detail below), assume that the
corrugated plate spring characteristic of a vibrator is
KA(~ ) and the connecting spring characteristic of a sieve
is KB( ~), its determination is quite simple. WA(~) and
WB( ~ ) should be decided first according to the require-
ments of the sieve. When WA~Z~) and WB(z~) are n times or
l/n times as much as the applied electric frequency (n is
an integer) resonance will take place. KA(~) and KB(Z~)
will be determined so long as the mass of electromagnetic
vibrator MA and the mass of sieve MB are given, and then
the characteristic and number of the springs can be
designed according to the method described above.
(b) Control and Adjustment of the Stability of
Performance Characteristics
Conventional linear electromagnetic vibration devices
and vibration machinery are very limited in the region of
the stability of performance characteristics under the
conditions of fluctuating external loads. Accordingly,
complicated electrical control mechanisms are frequently
required to improve the stability of performance charac-
teristics. Some existing non-linear type electromagnetic
vibration devices, such as rubber spring electromagnetic
vibration units and the vibration machinery constructed
with it, have been slightly improved in performance stabil-
ity owing to the feedback of non-linearity characteristic
curve, but an electrical control mechanism is still needed
and it is very difficult to control and adjust the self-
feedback.
No matter how external loads are fluctuated in the
form of mass variation, rigidity variation or damping vari-
ation and so on, an electromagnetic vibration device
according to the instant invention using non-linear bellows
- 36 - 1257~21
plate-spring groups, by taking such spring group as power
source and adding a non-linear type elasticity control
element according to the requirements, may utilize the
self-feedback of the non-linear type elastic element to
control and adjust the performance characteristics in a
comparatively large range by selecting and fitting the
proper control spring matching the non-linear elasticity
curve according to the scope of external load fluctua-
tions. Therefore, the vibration system can maintain a
stable state of resonance within the scope of load fluctua-
tlon from start to the end. In this manner, the stable
performance characteristics can be obtained without adding
any electrical control mechanism.
(c) Control and Adjustment of Parallel Connection
Combination Push-Pull Resonance of Multi-Sets
of Vibration Facility
Conventional electromagnetic vibration facilities are
usually designed on the basis of a complete series. The
types in the series of large electromagnetic vibration
facility used as the power source of large vibration machi-
nery are relatively numerous. Accordingly, the large elec-
tromagnetic vibration facilities used in large vibration
machinery is usually low in efficiency.
Another method of control and adjustment of this in-
vention is illustrated in Figure 12. In a large vibration
~ machinery having sufficient total rigidity it is possible
; to use parallel multi-sets of small non-linear type elec-
tromagnetic vibration devices. The same power supply can
be utilized to realize the parallel connection combination
~30 type push-pull resonance of multi-sets of smaller facili-
; ties in the same phase. Conseguently, small devices can be
combined to match large machinery and control and adjust-
ment can be made practical.
`:
;
. ~, , .
-- 37 --
1257;:121
Particular Applications
A sieving machine is a kind of general machinery for
selecting solid material according to grain grading.
Most conventional sieving machines employ a connecting
rod or inertia vibration mechanism constructed with a
rotary type motor. The power consumption of such units is
high while vibration frequency can not be easily in-
creased. Moreover, it is difficult to control and adjust
the stability of vibration of such machines with the change
in the mass of body which is being sieved.
As shown in Figure 13, one support type sieving
machine with a bi-layer sieving is constructed including a
non-linear electromagnetic vibration facility 50 as the
power source. The non-linear electromagnetic vibration
dynamic head 50 is fixedly connected with a holder 51 to a
sieving frame 52. At its front upper part, there is a
hopper 53 for feeding material, and in its middle part and
at its bottom, is a layer of sieviny net 54. At both sides
of the nets 54, a tension holder is fixed and the net 54
can be tensioned by a tension device fixed on the sieving
frame 52 and tension bolts 57. In the material discharge
direction of sieving frame 52 and at the bottom of frame
52, are hoppers 58 and 59 for discharging material are
respectively connected and by which the material sieved by
using the sieving net with bi-layer and different net can
be collected based on coarse, medium and fine grades. The
whole sieving frame is placed on the support 60 of sieving
machine through three groups of non-linear type control
elastic elements installed respectively at its two sides.
Non-linear type control elastic elements consist of (a) a
combination spring group which is composed of rectangular
bellows type plate spring group 61, compressed spiral
spring 62 and tension spiral springs 63, and (b) the spiral
coil spring 64 or (c) other kinds of proper non-linear type
- 38 - 1257321
elastic elements. At the bottom of support 60 of sieving
machine, a vibration isolator piece 65 is installed.
When exciting vibration current flows in the dynamic
head S0 of non-linear electromagnetic vibrator, the elec-
tromagnetic vibrator is excited, by which the sieving frameis directly actuated to cause working vibration. The
number of bellows type plate spring pieces of the electro-
magnetic vibrator can be adjusted to cause the machine to
reach a resonant state. In reliance on the self feedback
characteristics of the non-linear spring group, the reso-
nant state of the sieving machine can be kept stable con-
sistently within the scope of mass variation from no load
to full load.
Compared with a connecting rod type vibration sieving
machine having a conventional rotary motor and under the
conditions of same area and productivity of sieving net, it
has been found that the power consumption of this sieving
machine can be 5 to 10 times less than that of the latter
one.
When concrete surfaces are deposited by layers, the
surface of each layer must be roughened after removing an
outer surface skin to ensure binding strength between suc-
cessive layers. It is also often necessary in building
maintenance and renewal jobs to roughen a generally planar
concrete, asphalt or other similar surface.
Conventional manual roughening is low in efficiency
and heavy in labor intensity with a high rate of loss in
concrete. A method of roughening by water under high pres-
sure not only employs complicated and costly equipment with
high energy consumption but is also subject to rather rigid
seasonal restrictions. A method of roughening by sand
blasting, besides requiring large energy consumption, also
- 39 - i25~
wastes large amounts of sand and requires auxiliary work
such as cleaning, material transportation, etc.
As shown in Figure 14, a concrete chisel roughening
machine including a non-linear electromagnetic vibration
5 unit is shown. At the front lower part of the dynamic head
base plate of the electromagnetic vibrator is a roughening
chisel head. The chisel head includes a plurality of
chisel tips 31 arranged on a chisel tip plate 32. Plate 32
is fixedly connected with a base plate 33 of the roughening
10 chisel head and the base plate 1 of the electromagnetic
vibrator. The coarse or dense distribution and regular
arrangement of chisel tips can be adjusted after disassem-
bling the chisel tip plate 32. The heads of chisel tips 31
are welded with carbide alloy, wear-resistant steel and may
15 be sharpened by multiple wheel grinding. The base plate 1
installed with the dynamic head of the electromagnetic
vibrator and the part of the roughening chisel head is
connected with a base plate 34 of the carriage by a control
spring group which includes three groups of combination
20 springs being composed of spiral coil spring 35 and pull
spring 36, and two groups of rectangular bellow type plate
spring groups 37. Other kinds of non-linear elastic ele-
ments, such as "U" type plate spring etc. can also be used
as control spring groups. Pretension adjustment of the
25 combination spring group can be made by bolt 38 and nut 39.
A controller is mounted at a front lower part of the
base plate 34 of the carriage for controlling the depth of
chiseling and cutting. A spiral coil spring 41 and a con-
trol head 42, welded with a carbide alloy tip, is enplaced
30 within a seat 40 connected to the base plate 34. On the
control head 42, there is a guide key slot. Along the
guide is a limit bolt 43, the control head can move for a
certain degree of controlling the chiseling depth, thereby
adapting itself to change in a certain working condition.
35 Self-lubricating sleeve 46 is installed between the shaft
~ 40 - 125~3~1
44, which is connected on the base plate 34 of carriage,
and two wheels 45, thus, making the operation easy and
maintenance simplified. At a rear end of the base plate 34
of the carriage, handle seat 47 and handle 48 are in-
stalled, and between seat 47 and base plate 34, a shockabsorbing seat 49 is installed for damping shock during
operation.
The concrete chisel roughening machine, by electromag-
netic vibration, is a kind of small scale, flexible and
carriage type construction machine tool which can be
operated by one man. When exciting vibration current flows
in the dynamic head of electromagnetic vibrator, the
dynamic head is vibrated, and the vibration can be con-
verted into workpiece chiseling and cutting action by the
roughening chisel head. An adjustment of the number of
bellow type plate spring pieces of electromagnetic vibrator
can cause the machine to reach a resonant state; then, the
roughening operation can start. During operation, the
worker should press the chisel head firmly on the ground,
walk backwards, slightly moving the operation handle to the
left and right, and in the respect of construction tech-
nology, preflatten the concrete which is still during the
initial solidified period by concrete flat plate vibrator
for vibrating the concrete on surface layer as technical
preparation, thus, ensuring the removal of an upper layer
on the concrete surface with high efficiency. In the case
of concrete with different strength owing to different
pouring periods, i.e. under the conditions of varying scope
of rigidity and damping of external load, corresponding
feedback variation in rigidity and damping of whole vibra-
tion system can be employed via the non-linear control
plate springs having self feedback regulating characteris-
tics. Thus, machine performance can be kept stable.
As shown in Figure 15, a flat plate concrete vibrator
including the non-linear electromagnetic vibrator drive is
- 41 - ~ 2 5 7~3~ 1
shown. A group of non-linear cantilever type plate springs
19 or other kinds of non-linear elastic elements are in-
stalled between the dynamic head of electromagnetic vibra-
tor and a working flat plate 20. In this embodiment the
5 entire dynamic head of electromagnetic vibrator is in-
stalled in a cantilever form. Onto the working flat plate
20, one arcuate impact block is welded, the position of
which is corresponding to that of another arcuate impact
block welded under the base plate 1 of the dynamic head of
10 electromagnetic vibrator. A layer of polyurethane rubber
is adhered on the surfaces of the impact blocks so as to
reduce the impact noise. The distance between impact
blocks can be adjusted by a washer piece for controlling
the impact level. A connecting ring is welded at the four
15 corners of working flat plate 20. Then, lifting ropes on
the edge may be secured to connector bores on the flat
plate 20 can be employed when the concrete flat plate
vibrator is used for vibrating a surface of thin concrete.
When exciting vibration current flows in the dynamic
head of electromagnetic vibrator, the head is vibrated, and
vibrating force, vibrating moment and the impact force
produced by the impact blocks are generated on the working
base plate 20. The direction of resultant vibration is
mainly perpendicular to the base plate 20, and the resul-
tant vibration is transmitted downward, at the same time,there are also horizontal components, thus, making the
concrete flat plate vibrator for vibrating the surface of
concrete haul easily and more conveniently. Adjustment of
the spring piece number of bellows type plate spring of the
electromagnetic vibrator can cause the machine to reach the
resonant state. When vibration is met with the different
concrete ~dilute or dense level of different concrete), the
rigidity of the external load will be changed. At this
time, self feedback adjustment will be made responsive to
changing load by the non-linear plate springs, so as to
!
- 42 - lZ5~3Z~
make total rigidity of the system nearly unchange and main-
taining the stable vibrating working characteristics.
When the vibration strength produced by this machine
is similar to that generated by flat plate vibrator with
5 1.1 KW inertia type vibration motor, the power consumed is
only 0.1 - 0.2 KW.
Summar~ Of The Major Advantages
The non-linear elastic component of the vibrator is
provided via waveform plate springs fixedly bounded in
relation to the electromagnetic core and armature. Smooth
non-linear characteristics are generated with enhanced
efficiency in many functional respects. The degree of non-
linearity may be conveniently selected and altered by
design and various combinations of flexible rigidity can be
obtained with minimal internal damping and longer fatigue
life. The range of self-stability and adjustability, and
precision of directional control are greatly expanded. The
vibration amplitude can be multiplied by a factor of 2 or 3
and above. Additionally, harmonic resonance frequency
multiplication, reduction and conversion can be achieved
mechanically and simplistically without requiring a supple-
mental electronic controlling device.
The invention enables a plurality of vibrators to
reach push-pull resonance, resulting in increased effi-
ciency relative to a larger scale single vibrator of equi-
valent yield. The increased efficiency may be reflected by
a reduction in energy consumption approaching 50~ and
more. With a simple structure, not requiring any auxiliary
electronic control mechanism, the instant vibrator has
greatly enhanced overall effectiveness, performance effi-
ciency, and range of self-adaptability to varying applica-
tion environments, as well as a reduced energy consumption
- 43- ~z57~21
requirement from 1 to ten fold less than a corresponding
assembly according to the prior art.
With the instant non-linear vibration device, a much
greater amplitude and vibration acceleration can be ob-
5 tained for the same vibration frequency; e.g., for a vibra-
tion frequency of 50 Hz, double amplitude of 6 to 7 mm or
more is obtainable. Under complicated and randomly varying
loads, the vibrator of the instant invention can always
maintain a stable resonance state by its self-adapting
10 feedback behavior without any electrical control
mechanism. The instant vibrator has a simple structure,
produces precise vibrating motion without the use of
external guides, or requires no lubrication, generates
minimal noise, can be readily sealed, has long service
15 life, and employs a power-head like construction for easy
mounting to a base component.
In describing the invention, reference has been made
to preferred embodiments. Those skilled in the art, how-
ever, and familiar with the disclosure of the subject in-
vention, may recognize additions, deletions, substitutions,modifications, and/or other changes which will fall within
the purview of the invention as defined in the claims
below.
For instance, an alternative embodiment may employ
25 only one set of sectionalized waveform plate springs,
arranged only on one side of the electromagnetic source.
Also, plate springs having more than two sections thereon
may also be utilized.