JPS5815269B2 - Impact force generator - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION This invention relates to impact force generating devices and, more particularly, to an improved vibratory impact force generator that uses rotating eccentric weights to generate vibratory motion.

Various driving principles for vibrating impact hammers have been developed.

Some use one or more eccentric weights to generate vibrations that are used to break up pavement, compact soil, drill holes, and perform other tasks.

This kind of machine is driven by a rotating eccentric weight.
The most widely used of these is the use of multiple weights that rotate relative to each other in order to convert the centrifugal force of each weight into reciprocating motion along a single line matching the impulse motion of the ram. is now suppressed.

By using a plurality of weights in this manner, vibrations perpendicular to the desired direction of impact movement of the hammer can be prevented.

On the other hand, U.S. Patent Nos. 1,386,329 and 1,410,010
In the system using a single eccentric as in No. 1, vibrations perpendicular to the direction of impact are suppressed or controlled by a rigid guide member that forces the impact member to vibrate along a single path. .

In this case, the forces generated by the single eccentric weight in a direction perpendicular to the direction of impact are transmitted to the support frame, which not only cannot be used for impact movements, but also causes undesirable vibrations.

That is, although the above-mentioned US patent has the advantage of not requiring the use of dual eccentrics rotating in opposite directions, it has the disadvantage of not being able to eliminate lateral vibrations.

Additionally, this type of vibratory hammer, which uses one or more eccentric weights to drive the impact member, generally relies on the impact energy generated with each revolution;
In order to effectively transfer energy to the ram during each drive, the relative dimensions of the unbalanced weight system must be large.

In addition, auxiliary springs are typically used to store and redistribute energy to return the ram during each stroke, but these springs cause further vibration problems and do not direct energy to the ram. It doesn't work as effectively as it does.

Due to the above structure, the conventional vibration shock/damper was loud and heavy to hold.

The present invention places a single eccentric weight on the ram to effectively generate vibratory motion without substantially lateral restraint.

The motion induced in the ram by a rotating eccentric weight may include transverse and longitudinal vibrations at the center of gravity of the ram, or may be limited to vibrations only in the longitudinal direction, i.e. in the direction of impact of the striking plate of the ram. It has become.

The tool, which is impacted by the ram at the striking plate, is not subjected to lateral forces or relative lateral movement of the ram, eliminating wear and energy loss.

The ram, which is merely flexibly supported by an elastic material, generates lateral vibrations due to the lateral component of the rotating eccentric weight.

If the movement is applied in two parts in the lateral and longitudinal directions, the maximum longitudinal impact force applied to the tool at the striking plate of the ram cannot achieve the desired large driving capacity. .

This movement of the ram is stored in the form of restoring energy in conjunction with the flexible elastic fixture and is not transmitted through the tool to the workpiece, but is transmitted to the tool on the next impact and as an input from the eccentric weight. Stored continuously.

Because of this continuous accumulation and distribution of energy to the tool, the drive mechanism does not have to generate all the power of the impact needed for a cycle during that cycle, but rather the input that needs to be applied to the system on each cycle. is simply sufficient to compensate for the work consumed by the tool and the loss due to the mechanical efficiency of the system.

The hammer of the present invention has a vibration range close to the system's natural resonance point to prevent energy loss and store it within the system.

By properly relating the relative non-operating frequencies of the spring and ram systems, the overall efficiency of the hammer can be maximized.

Also, the relative size of the eccentric weight to the size of the ram can be reduced, thereby allowing the use of larger and quieter rams.

This correct combination of operating and design parameters results in less support vehicle vibration and longer hammer life.

The invention further includes restraint means arranged to reduce undesirable resonances and random vibrations.

This restraining means does not interfere with the normal oscillatory motion of the ram and does not detract from the efficiency and advantages of the system.

SUMMARY OF THE INVENTION Accordingly, it is an object of the present invention to provide an improved vibration shock absorber.

A second object of the present invention is to provide a vibratory impactor that can effectively utilize the lateral vibrational motion of the ram and at the same time impart a linear impacting motion to the impact tool.

Another object of the invention is to provide a vibration shock engine using a single eccentric to drive a ram toward a tool in which the ram is free to oscillate at right angles to the direction of vibration.

Still another object of the present invention is to provide a vibratory shock absorber that uses resonant motion of an impact member to effectively store and transmit impact energy.

It is a further object of the present invention to use a relatively large ram, vibrate it over a relatively short stroke, and yet generate a vibratory shock force comparable to that of conventional large, expensive and power-hungry equipment.・The aim is to provide the following.

Another object of the present invention is to provide a directionally constrained free vibration vibration system to reduce undesirable resonances and random vibrations.

That is, the present invention provides a vibratory shock absorber using natural vibratory motion that creates a quiet, convenient, and efficient system.

Other objects and advantages of the invention will become apparent from the description below.

In the illustrated embodiment, a support frame, indicated generally at 10, extends between and connects a pair of parallel, opposed, spaced apart side plates 12,14. It includes support members 16.18 that are rigidly connected to each other so as not to move, and a boom pin 20.22 that extends through the parallel side plates 12.14 and is mounted at right angles thereto.

These side plates 12, 14 have the same shape and are arranged parallel to each other to facilitate manufacturing.

Both side plates 12 to provide rigidity to the entire frame 10
．． A support member 16.degree. 18 extends between 14.

The side plates 12,14 are also arranged parallel to each other by a tool holder which will be described below.

Near the upper end of the mounting area 24 of the frame 10 is a parallel side plate 1.
2, 14 and at right angles thereto.
It's stretching out.

These boom pins 20,22 are adapted to engage a backhoe or other rigid articulated arm used to move or mount the device.

These boom pins 20.22 are long enough to project out from both side plates.

A locking pin 26 is attached to each boom pin to prevent the boom pin 20.22 from coming out from the first direction of the frame 10, and to prevent the boom pin 20.22 from coming out from the second direction of the frame 10. A clip pin 28 is used for this purpose.

This clip pin 28 is used to allow easy removal and insertion of the boom pin 20.22.

Frame 10 on which this hammer is attached to a backhoe
Collars 30 are welded to the side plates 12 and 14 around the boom pin attachment location to increase breaking strength and durability.

Located between both side plates 12,14 is an assembly generally designated 32.

The ram assembly 32 generally includes side plates 12 of the frame 10.
．． a ram 34 elastically attached to 14;
4 and a driving eccentric weight 36 that vibrates.

The shape of the ram 34 is essentially a variation of an "I9" beam, but because of its complexity, it is conveniently cast in iron or steel, such as ducttail iron.

The ram 34 is completely fitted between the steel plates 12,14 and has its generally "1" beam side or "open" portion facing each side plate.

The shim 34 is an impact member 34 that is elastically attached to the support frame 10 so as to be able to vibrate freely relative to the support frame 10 .

That is, in the illustrated embodiment, the ram or impact member 34 has an eccentric weight housing 38 defining therein a cylindrical space 40 for receiving the eccentric weight 36, and a ram or impact member 34 arranged at right angles to the axis of said cylindrical space 40. a mounting bracket 86 extending upward from the eccentric weight housing 38; a second mounting structure 92 extending downward from the eccentric weight housing 38; and a lower bracket extending downward from the second mounting structure 92. 106.

Eccentric weight 36 is a vibratory drive means 36 operatively attached to said impact member 34 to produce a linear motion in said impact member 34 at a location on said impact member 34 .

That is, the eccentric weight or vibration drive means 36 is housed within the cylindrical space 40, and its axis is aligned with the side plate 12.1.
The housing 38 is pivoted at right angles to the housing 38.

Reference numerals 109, 110 and 112, 114 are elastic members which connect said ram or impact member 34 to said side plates 12.14.
One end is attached to the side plates f2, 14 of the support frame 10, and the other end is attached to the mounting bracket 86 and the second mounting structure 92 of the impact member section 34 in order to elastically hold the shock member section 34 between the side plates f2 and 14 of the support frame 10. are installed respectively.

Reference numeral 107 is a striking plate, which strikes the upper surface of the lower tool 146, that is, the anvil surface 148, and the lower surface of the striking plate 101 is connected to the impact member 34 that can obtain the linear motion.
forming a shock transmission surface located above.

Reference numeral 146 designates a tool, which is slidably attached to the support frame 10 at a position such that it makes impact contact with this impact transmission surface.

Reference numeral 72 indicates a hydraulic motor mounted on the outer surface of the housing 38, and the hydraulic motor 72 has a motor shaft 78 for engaging with an axial hole of the eccentric weight 36 to rotate the eccentric weight 36. has.

82 is a balance weight 82 mounted on the outer surface of the housing 36 on the opposite side of the hydraulic motor γ2.

As will be described further below, the total weight W of the ram assembly 32 that moves together with the ram 34, including the impact member, namely the ram 34, the eccentric weight 36, the hydraulic motor γ2, and the balance weight 82, and the center of gravity of the entire ram assembly 32. The moment of inertia (2) of the ram assembly 32 around an axis that passes through The distance ds is selected such that the relationship between the distance ds and the axis of None, so that at the point of the node, only vertical motion occurs, and no horizontal motion occurs.

In an exemplary embodiment, shown for illustrative purposes only and not to limit the scope of the invention, the weight of the ram assembly 32 is a movement restraint moving with the ram 34, such as an eccentric weight 36 or a hydraulic motor. Approximately 1 including all spring parts
It is 65 kg (3601b).

In the embodiment of the present invention, the moment of inertia about the axis parallel to the axis of the eccentric weight 36 and on the center of gravity of the ram assembly is approximately 4.4 kg·m (160001b, 1n2), and the eccentricity of the eccentric weight 36 is approximately 0.76kg・Cr
(64in-1b).

That is, it can be understood that the weight of the eccentric body is extremely small compared to the weight of the entire ram that generates vibrations.

An eccentric weight housing 38 is formed within the ram 34 above the center of gravity of the ram 34.

For this eccentric weight, for Ujifugu 38, eccentric weight 3
A cylindrical space 40 containing 6 is included.

Circular flanges 42.44 protrude from the cylindrical inner wall of housing 38 for accommodating bearing housings 46,48.

The bearing housing 46.degree. 48 is cylindrical in shape, and its dimensions are such that it fits circularly into the flange 42.degree. 440.

In order to fix the flange 42゜44 of the eccentric weight housing 38 in the bearing/Ujifugu 46.48) 5
0.52 is used.

The cylindrical bearing housing 46 is adapted to engage the circular flange 42, 44 to which this bolt 50,52 is attached.
An annular ring 54,56 extends outwardly from 48°.

This arrangement of the bearing housing 46,48 allows the ram 34
The axis of the eccentric weight 36 relative to the axis is determined.

The distance between the eccentric axis and the center of gravity of the present invention is approximately 1o, 8Cr
(31/4 inches).

Eccentric weight 36 is circular in shape as shown in FIG.
, 60.

Axles 62, 64 projecting inwardly from the bearings 58.60 engage the eccentric weights 36.

A hole 66 is formed on one side of the eccentric weight, parallel to its center line, in which an axle 62, 64 is accommodated.

The axles 62, 64 are circular in cross-section, with each end having a bearing 58.
．． It is housed within the inner race of 60.

The inner end of the axle 62, 64 is connected to the hole 66 in the eccentric weight 36.
and are coupled together such that torque can be transmitted through the axle 64 to drive the eccentric weight 36.

A collar is formed at the center of each axle 62,64,
Each collar 10 is larger than the inner race of the bearings 58,60 and the hole 66 in the eccentric weight 36, and serves to center the eccentric weight in both bearings 58,60.

The hydraulic motor 72 is attached by fasteners 52 to the bearing housing 48 and thus to the eccentric weight housing 38.

The hydraulic motor 72 has a fluid inlet lower 4 and a fluid outlet lower, and fluid is supplied to the motor 72 or the attached impact member 3.
It is sent through a flexible tube that can withstand 4 degrees of vibration.

The shaft 78 of the motor is formed with splines that engage splines 79 in a hole 80 formed in the axle 64 for torque transmission.

A balance weight 82 for balancing the motor 72 is provided at the other end of the eccentric weight 38 on the opposite side of the hydraulic motor 72.

This balance weight 82 is fixed to the eccentric weight housing 38 by a fastener 50 via a bearing/magnet 46.

When the ram 34 vibrates due to the rotation of the eccentric weight 360, the unbalanced load on the motor γ2 is canceled by the balanced weight 82.

This is accomplished by aligning the motor 72 and counterweight 82 with their combined center of gravity at the centerline of the ram 34.

A semicircular cutout 84 is provided in the side plate 12.14 to accommodate the motor 72 and the counterweight 82.

These notches 84 are caused by vibrations generated by the 2mm 34 to cause vibrations from the motor γ2 or the balance weight 84 to the side plate 12.
It is larger than the motor T2 and the balance weight 82 so that no impact is applied to the motor T2 and the balance weight 82.

The ram 34 extends upwardly from the eccentric weight bracket 38 to form a mounting bracket 86, which is formed with two flat mounting surfaces 88 and 90.

A second mounting structure 92 extends downward from the eccentric weight housing 38, and this mounting structure 92 also has two mounting surfaces 94 and 96 formed therein.

This installation is carried out on both sides of the structure 920 by four webs 98100.
102 and 104 are formed.

These webs 98, 100, 102, 104 are connected to the ram 3.
4 to a lower bracket 106 at the lower end.

The axis of the eccentric weight 36 of the ram 34 is such that the node for horizontal movement of the ram 34 is located at the lower bracket member 10.
It is positioned so that it is on the bottom of 6.

A striking plate 107 made of hardened steel is attached to the underside of the lower bracket 106 to form an impact surface for transferring energy from the vibrating shim 34 to the impact target.

This striking plate 107 is bolted to the lower bracket member 106 with four bolts 108.

Since the horizontal nodal point of the ram assembly 32, or the point of the nodal point for horizontal motion, is on this striking plate 107, only vertical motion occurs at this point.

Since there is no horizontal movement between the target and the center of the surface of the striking plate 107 during impact, wear on the striking plate 107 and the target can be prevented, lateral impact loads on the ram assembly 32 and the target are eliminated, and vertical movement towards the target is eliminated. The overall impact energy increases.

In order to correctly position the striking plate 107 relative to the center of gravity, the following method is used.

First, the center of gravity of the entire ram assembly 32 is determined.

Next, the moment of inertia about an axis passing through the center of gravity and parallel to the axis of the eccentric weight 36 is calculated.

Finally, the distance ds between the center of gravity and the striking plate 107 is the distance d between the center of gravity and the eccentric weight 36.

It is expressed as a coefficient of That is, here, ■ is the moment of inertia of the ram assembly 32 about the axis passing through the center of gravity and parallel to the axis of the eccentric weight 36.
b-1n2, W is the weight of the ram assembly 32 expressed as Ib, de is the distance from the center of gravity of the ram assembly 32 to the axis of the eccentric body 36 expressed as jn, and from this formula, the striking plate is The correct position of 107 can be determined.

Although the above method of determining the horizontal nodal point does not take into account the total amount of elastic force that attaches the ram assembly 32 to the frame 10, the motion of the ram is much greater than the natural frequency of the elastic fixture and the horizontal nodal point Since attachments further away from the horizontal node are more resilient than attachments closer to the horizontal node, and these attachment members are more flexible, elastic attachments allow the sum of forces to have less influence on the position of the horizontal node.

The ram 34 is resiliently attached to the frame 10 by means of resilient mountings designed to allow free relative movement of the impact member 34 with respect to the frame 10.

These resilient means include resilient members 109, 110 in combination with the mounting bracket 86 and resilient members 112, 114 in combination with the mounting structure 92.

The elastic members 109 and 110 are the same as the ram assembly 32.
Prevent unbalanced loads.

The members 109, 110 are constituted by cylindrical rubber blocks, one end of which is engaged in each plate 116, 118.

These plates are mounted on flat mounting surfaces 88.90 of brackets 86 and bolted thereto with fasteners 120.

The other ends of members 109, 110 are connected to plates 12L, 123, which are bolted to side plates 12.14 via spacers 122, 124.

The elastic members 112 and 114 have the same shape and are cylindrical.

One end of each elastic member 112, 114 is attached to a mounting plate 126.1.
28, and each mounting plate 126, 128 is attached by a fastener 180 to the mounting surface 94, 96 of the structure 92.
is bolted to.

The other ends of the blocks 112, 114 are bolted to the side plates 12, 14 by bolts 136 through spacers 132, 134.

These resilient members 109, 110, 112, 114 are all attached to the ram assembly 32 and the side plate 12 so as to flex during shearing motion due to the vibrations produced by the rotating eccentric weight 36.
, 14.

In the above typical example, the static spring constant of the upper elastic member 109°110 in the shear direction is 41701 g/m (2251 b/i
n), the static spring constant of the lower elastic member 112° 114 in this shear direction is 12,700 kg/m (700 ib/in
).

Each elastic member 109, 110, 112, 114 is made of a rubber block.

When used in this way, the dynamic motion spring constant of the rubber exhibits a value different from the actual static spring constant.

Empirically, the dynamic spring constant of the entire mounting means under vibration conditions when the system of the present invention is driven is approximately 49,800 kg/m (27
50 ib/in).

The state of motion of the ram assembly 32 relative to the frame 10 is
0 conceptually.

This movement shown conceptually is due to the elastic members 109, 110, 1
It is constituted by the unrestrained vibration of the ram assembly 32 held at 12 degrees 114 and driven by an eccentric weight 36.

No guide means are required to obtain the free movement shown.

Due to the inertia issues and elastic support system previously discussed, the center of gravity of the ram assembly 32 follows a circular path 137.

At striking plate 107, a completely longitudinal motion, as shown by straight line 138, is imparted to the ram by horizontal nodes.

This straight line 138 is the locus of the center point of the striking plate during vibration of the ram assembly 32.

The diameter of the circular path 137 of the center of gravity of the ram assembly 32 is equal to the length of the straight path 138 at the lower elastic members 112, 114, and the ram follows an elliptical path 140 with its principal axis on the centerline of the ram 34. Exercise.

The length of the main axis of the elliptical path 140 is the same as the major axis of the circular path 137 and the straight line 138.

At the upper elastic members 109, 110, the ram 34 traces an elliptical path 142 whose principal axis is perpendicular to the centerline of the ram 34.

The main axis of this elliptical path 142' has the same length as the straight line 138.

The movement of the ram 34 described above is based solely on the inertial characteristics of the ram 32.

That is, when correctly positioning the striking plate 10γ so that the striking plate 107 coincides with the horizontal node, only the inertial characteristics are used, or the relative spring between the upper elastic member 109° 110 and the lower elastic member 112i14 is used. Since the constant is small, these elastic members 109°110.112,114
The effect on movement is negligible.

The upper elastic members 109, 110 are more flexible than the lower elastic members, so lateral movement of the ram 34 is resisted more by the lower elastic members 112, 114.

That is, a restoring force system is formed due to the difference in flexibility and the difference in movement path of the two sets of elastic members, and this restoring force system causes the vibration movement to occur near the striking plate 101 where the horizontal vibration becomes a wave node (
node).

Therefore, when creating a horizontal node at the striking plate 101, it is widely practiced to consider the inertial system as described above.

The vibration of the ram 34, which moves as described above due to the inertial and spring characteristics of the hammer, is generated by the rotation of the eccentric weight 36.

This eccentric weight 36 is symmetrical with respect to a plane parallel to the side plates 12 and 14, and is designed to generate a driving force within this parallel plane.

Due to this symmetrical property as well as the symmetry of this overall hammer shape, the movement of the ram 34 takes place parallel to the side plates 12.14.

Furthermore, since the frequency of vibrations generated by the rotating eccentric weight 36 is the same as the rotating frequency of the eccentric weight,
The symmetry and speed of the eccentric weight 36 determines the characteristics and speed of the ram motion.

The stroke or distance traveled by the ram 34 during free vibration of the ram assembly 32 within the frame 10 is the weight of the ram 34;
The force generated by the rotating eccentric weight 36, the elastic members 109, 110, 112° 114, and the eccentric weight 3
It is determined by the frequency of 6.

The size and spring characteristics of this device are determined at the time of manufacture, but the frequency can be changed during use.

The natural frequency of the free vibrating ram in the embodiment of the present invention is 1
Approximately 490 cycles per minute.

Care must be taken as this frequency may have little effect on the impact capability of the hammer, or even at this natural frequency the stroke may be long enough to create a failure condition.

The non-mer must be able to operate above the natural frequency of 2@, and must quickly pass through this range.

In an embodiment of the invention, the free vibration of the ram is 160 vibrations per minute.
A fluid divider 144 is provided to control the frequency of free vibration of the 1-ram 34, which has achieved a stroke that will not damage the ram or the elastic member even when the ram reaches 0 cycle.

When the vibration frequency is 1600 epm, there is no need to apply a large amount of power due to the free vibration of the ram 34, and the rotating eccentric weight 36 can be driven by the motor γ2 without sending a high input from the hydraulic input. Although the motor 72 does not require high pressure, it must have high flow rates and high pressure capabilities in order to operate the hammer in unloaded as well as loaded conditions, so the hydraulic input is free to use ram 3.
A higher input is given than when vibrating 4.

The fluid divider 144 is provided to allow hydraulic fluid to be distributed to the inlet cuff 4 of the pump 72 regardless of the pressure in the system.

In the system of the present invention, the flow rate is reduced to 1 by this fluid divider 144.
Limited to 2gpm.

In addition to this fluid divider 144, the hydraulic fluid supply line going to the motor 72 is provided with a pressure control device in the form of a pressure release valve 145.

In this particular embodiment, this pressure relief valve is 119 kg/7
! A setting of 2 (1700 psi) has been found to be desirable.

Excess fluid above this pressure is diverted through this pressure relief valve 145 to the return line from the motor to maintain the above pressure.

That is, when the use of this device is interrupted, motor γ2 or 12
Using full gpm, the pressure downstream of fluid divider 144 is automatically reduced to a value much lower than the value set by valve 145 and merely that required to drive eccentric weight 36.

In this way, controlling the flow velocity rather than pressure control of the vibration frequency during non-load conditions has the advantage that the inertia of the hydraulic system can be maintained at the operating level during load conditions.

That is, there is no need to initially accelerate the system when a load is applied to the ram 34.

Additionally, it is possible to move the device to a new location, which would be undesirable or impossible with other devices, or to allow the device to continue operating in a freely suspended state while the backhoe vehicle is being moved.

- In order to effectively utilize the impact energy from the ram, the target must be placed within the free vibration path of the ram 34.

By doing so, the vibrational energy of ram 34 can be transferred to the target during each impact vibration.

Of course, the vibration characteristics of the hammer change significantly when the target is placed within the free path of the ram 34.

A target is shown as a tool 146 in this embodiment of the invention.

The concept and method of operation of the present invention is not limited to a configuration in which the ram 34 is applied directly to the material on which the impact force of the hammer acts.

Other variations include placing a tamping foot at the horizontal node of the ram, which tamping foot is used for operations such as directly compacting the soil.

The tool 146 in this embodiment merely serves as a conductor for transmitting impact energy from the ram 34 to the workpiece.

The tool 146 is provided with an anvil surface that receives the impact force from the ram 34, and a generally cylindrical body 150 extends downwardly from the anvil surface 148.

A machining position 152 at the lower end of this tool body 150 is shown in FIG.

This tool 146 is restrained by a tool holder 154 so as to move vertically relative to the frame 10.

This tool holder 154 is attached to the side plate 1 by a fastener 156.
2, 14 and the frame 10.

Eight fasteners 156 are shown in the drawing.

The tool holder 154 has a through hole 15 that accommodates the tool 146.
It is constructed by a cast block having a diameter of 8.

The diameter of a hole 158 passing through the tool holder 154 is sufficiently larger than the diameter of the tool so that the walls of the hole 158 do not interfere with the movement of the tool 156.

The diameter at both ends of the hole 1580 is increased to accommodate the upper and lower bushings 160, 162.

Upper bushing 160 and lower bushing 162 are in sliding engagement with tool body 150 extending therethrough.

The upper bushing 160 is placed in the hole 158 of the tool holder 154 by pushing it into this hole 158 until it reaches the bottom of the enlarged portion of the hole, and then the set screw 16
4 fixes the bushing 160 to the tool holder 154.

Upper bushing 160 is thus fixed in position.

On the other hand, the lower bushing 162 is inserted into the hole 158 from below the tool holder 154 and inserted upwards to the point where the diameter of the hole 158 becomes smaller, and the tool stop plate 166 is bolted to the lower surface of the tool holder 154. This prevents the lower bushing 162 from moving downward or being pulled out of the tool holder 154.

The cylindrical tool 146 is attached to the tool holder 154 by a tool holding pin 168 that matches the notch 170 in the tool body 150.
held within.

As shown in FIG. 5, this tool holding pin 168 crosses the tool holder 154 while partially passing through the hole 158 of the tool holder 154, and has extensions 172 at both ends,
114 are two holding plates 1 extending from the tool holding pin 168.
76.178.

These retaining plates are attached to both sides of the tool holder 154 and tightened with bolts 180.

Notch 110 is formed to allow tool 146 to move freely within hole 158.

- When the tool 146 is raised away from the workpiece, the tool 146 falls down where it engages the upper end of the tool holding pin 168 or notch 170.

The length of this notch 170 is such that the tool 168 can fall completely out of the path of the rotating ram 34.

Tool retaining pin 168 is designed so as not to restrain tool 146 under load conditions.

That is, notch 170 extends far enough down tool body 150 to prevent tool retaining pin 168 from contacting the lower end of notch 170 when the tool is raised toward ram 34.

To prevent the tool 146 from moving into the path of the ram 34, the cylindrical body 150 is provided with a collar 182.

This collar 182 is attached to the tool stop plate 166 at a predetermined load position.
It is provided in a position where it engages with.

In normal operation of the hammer, when the tool 168 is raised, the collar 182 and the tool stop plate 166 are engaged.

Limiting the movement of tool 146 provides two benefits.

First, the tool automatically disengages from impact engagement with the shim 34 when the hammer is lifted off the workpiece.

This automatic separation of the tool prevents the destruction of the non-mer due to the collision of the ram with the frame.

Other shock engines require sophisticated mechanisms to prevent damage to the ram or frame when operating under unloaded conditions.

A second advantage of limiting tool 146 movement is that a preselected pre-load condition of the ram can be automatically achieved regardless of the downward force applied to the non-mer by the backhoe.

The backhoe is operated by pressing the frame 10 against the workpiece until the collar 182 does not come into contact with the tool stop plate 166, but when further downward force is applied by the backhoe, the collar 182 prevents the tool 146 from moving upward. , the force is directly transmitted to the workpiece.

That is, the position of the tool 146 before loading is the tool body 150.
is determined by the position of collar 182 of.

The position of tool 146 before loading is shown in FIG.

The preload position of tool 146 determines the degree to which the tool intersects the path of vibratory ram 34.

That is, if the preselected pre-load condition is excessive, the eccentric weight 36 will not provide sufficient force to lift the ram off the tool for the next impact.

This state is caused by the elastic member 1 being pushed up against the frame 10 by pushing the tool 146 and the ram 34 upward.
This occurs when the rotating weight 36 cannot overcome the downward force of 09, 110, 112° 114.

Conversely, if the preload condition is insufficient, the velocity in the impact direction will be lost before the striking plate 101 of the hanger 34 contacts the anvil surface 148 of the tool 146.

In this state, the maximum effective impact force cannot be obtained.

Therefore, the range of pre-load conditions varies depending on the shape of each individual beam.

This only requires changing the distance from collar 182 to anvil 148.

By positioning the tool 146 in a position when the resilient member is in a deformed state at the impact position, the inertia of the ram 34 allows the entire impact force to be transmitted.

If the tool is placed in a position where the shim 340 elastic member is further pulled downward upon impact, that downward force will be added to the inertial force created in the shim 34 upon impact.

The collar 182 of the tool is attached to the elastic member 109, 110° 112
．． By placing the impact location in a relaxed position of 114, a substantially linear vibration stroke response can be achieved.

This response state is shown in FIG.

Leaving the elastic member in a substantially unloaded state at the impact location results in a non-linear vibratory stroke response.

This nonlinear vibration stroke response state is shown in FIG.

In practice, one must determine the vibration-stroke response curve that will result in the most desirable conditions for which the hammer will be used.

This vibration-stroke response relationship is controlled in part by selecting the amount of upward movement of the tool relative to the ram 34.

Since stroke length has the greatest effect on linear vibration, hammers that must be used in a variety of conditions should not be preloaded to linear vibration-stroke response conditions.

Since it is difficult to maintain a constant vibration when the elasticity of the workpiece changes, the stroke should be changed gently with respect to the nonlinear vibration-stroke response system.

Furthermore, by not making the stroke length excessive, the life of the member can be extended.

The vibratory motion of the ram 34 varies greatly depending on the position of the tool 146 in the path of the vibratory ram.

When the motion is resisted by a tool placed on the workpiece, the time-displacement curve of the vertical vibration of the shim 34 changes from a sine curve to an intermittent waveform motion, as shown in FIG.

If the impact of the ram 34 on the tool 46 is completely elastic and the tool is midway through the vertical travel of the ram, the natural frequency of this system operated in a straight line will be the free natural frequency of the shim 34. However, since the system of the present invention is expected to be in a nonlinear state, its natural frequency is the same as that of the freely moving ram 3.
It does not become 2@, which is the natural frequency of 4.

That is, the natural frequency of the ram in this embodiment is not 980 cpm, which is the free natural frequency of the shim 34, which is 490 epm, but
The response state of the system of the invention is approximately equal to the curve of FIG.

Therefore, the frequency during operation is wide, far exceeding the nonlinear natural frequency.

It has been found that the system of this example operates best at 1150-1350 epm under load.

This indicates that the operating region of the system of this embodiment is in the nonlinear curve 184 of FIG.

In the free vibration system of the system of this embodiment, the system is symmetrical, so the movement of the ram 34 is performed in a plane parallel to the two side plates 12 and 14.

The operating vibration range of the hammer under load depends on the input pressure to the motor 72 and the elasticity and hardness of the workpiece.

The ram 34 is driven at the same frequency as the driving eccentric weight 36.

A pressure relief valve control valve 145 is provided downstream of the fluid divider 144 to act as a controller of the operating frequency of the system.

As previously discussed, when the ram 34 is freely oscillating, the input power is controlled by the maximum flow rate through the fluid divider, and the pressure downstream of the fluid divider 144 is as high as the motor 12 demands in this unloaded condition. It decreases until it becomes equal to the power.

When the tool 146 is in shock engagement with the ram 34, more power is provided to the eccentric weight 36 since some of the power is required to maintain the vibration of the system.

As eccentric weight 36 applies more resistance to motor 36, the pressure at input T4 increases.

In the control system of this embodiment, the valve 14
5 and the flow to motor 72 is reduced due to the loss of fluid through this discharge pressure system.

Equilibrium is achieved when the flow of hydraulic fluid into the inlet cuff 4 is less than the amount provided by the fluid divider 144.

The remaining flow through the fluid divider 144 is vented through the pressure relief system to the output passage to maintain a predetermined pressure at the input 14.

In the equilibrium state, vibration is achieved such that the eccentric weight 36 is rotated under a predetermined hydraulic pressure state.

By changing this device and its controller, such as by changing the pressure setting of the valve 145, the amount of pre-compression by the tool 146, and the flow rate to perform operation close to a different resonance mode, the frequency of the load system can be adjusted. can be changed.

Changing the pressure setting of the discharge valve 145, with other factors remaining the same, moves the system closer to or further away from its natural frequency.

Thus, the frequency and response stroke can be adjusted by adjusting the pressure of the hydraulic fluid supplied to the motor 12.

The pressure in this example is 119 kg/Cr2 (1700 psi
) is set.

As a result, even when the work status changes, it will change from 1150 to 1.
A working range of 350 cpm is obtained.

In this state, the stroke is approximately 1 nm (0.4 inches).

As can be seen from Figure 12, when the stroke response of the system reaches a low frequency point where it jumps from the upper curve to the lower curve, the stroke becomes extremely small and the efficiency of the system decreases. It is advantageous to maintain the stroke response of the system below a discontinuity in the nonlinear vibration-stroke response curve.

An unconstrained motion of the ram 34 would have the advantage of applying all the force of the backhoe directly along the centerline of the tool, but not only is this condition not always achievable, but the hammer may be placed at an angle to the vertical. In some cases, the vibrating hammer is used to attach and drive the vibrating hammer, and since the load changes temporarily when the resistance received from the tool 146 changes, this unbalanced state causes undesirable resonance and random vibration in the vibratory hammer.

Such vibrations are undesirable because they generate large vibrations that disturb the stable vibration state of the ram 34 and damage the elastic member.

To reduce such resonances and irregular vibrations, the frame 10 is provided with restraint means that cooperate with the vibration shim 34.

This restraining means is combined with the ram 34 near the horizontal nodes of the ram 34 in order to eliminate the aforementioned undesirable vibrations without compromising the operating efficiency of the ram 34 under normal vibration conditions.

The restraint means are generally 186° 188° mounted to the frame 10 at positions on either side of the vibrating gym 34 at horizontal nodal points.
It is composed of two guides shown in .

Since these two guides 186 and 188 are the same, only one guide 186 will be described.

Guide 186 includes a base plate 190 secured to side plate 12 by fasteners 192 .

A key 194 extends within a groove 196 formed in the tool holder 154 to prevent the guide 186 from rotating from its position.

Substrate 190 has one vertical groove 198 along its entire length.
is formed.

Both sides of the substrate 190 are provided with flanges 200 along its inner surface.
The flange 200 extends inwardly to form two lips that cover the groove 198.

The bottom of the groove 198 is open and an elastic damper 20 is provided as a means of elastically sealing the upper end of the groove 198.
2 is provided.

The elastic damper 202 is coupled to a thin plate 204 extending outward from the damper 202 so as to cover the top of the substrate 190.

An angle member 206 for fixing the damper 202 is arranged on the thin plate 204.

The angle member includes a leg 198 that partially covers the groove 198 and forms a lateral support for the resilient damper 202.

Two fasteners 210 secure the angle member 206 to the base plate 190.

The lower bracket member 106 has two sliding pins 2f2,
214 is included.

These sliding pins are cylindrical and extend outwardly from the lower bracket 106 of the ram 34 and are centered on either side of the striking surface of the striking plate 107.

This position relative to the striking plate 107 allows the pins 212,
214 is positioned at the horizontal node of the vibrating ram 34.

Sliding pin 212 is located within guide 186 and sliding pin 21
4 extend into the guide 188, respectively.

A sliding block 216, 218 is arranged around each sliding pin.

Both sliding blocks 216, 218 are identical and are generally rectangular and housed within the groove 198.

Since the sliding block 216 is housed within the groove 198, the sliding pin 212 is prevented from moving laterally.

A flange 200, which is associated with the base plate 190 to lock the sliding block 216 within the groove 198, projects inwardly.

Further, the sliding pin 212 is restricted in its upward movement because the sliding block 216 comes into contact with the elastic damper 202.

Therefore, not only are the sliding pins 212, 214 and the bottom of the ram 34 incapable of lateral vibration, but the damper 202 prevents irregular excessive vertical vibration of the ram 34.

The placement of the guides 186, 188 at the horizontal nodes of the ram 34 allows normal vibration of the ram 34 to occur unhindered by the pids 186, 188.

Therefore, the efficiency and advantageous exercise of gym 34 is not hindered by these guides.

Furthermore, the ram 34 is guided by the vibration guides f86 and 18 of the hammer.
There is no constant force applied to 8.

Since the present invention has the above-described configuration, the horizontal node or wave node of the horizontal movement of the ram assembly 32 is located on the striking plate 101, and at this point only vertical movement occurs and no horizontal movement occurs.

Therefore, since there is no relative horizontal movement between the center of the lower surface of the striking plate 101 and the anvil surface 148 of the tool during impact, the wear of the striking plate 107 and the anvil surface 148 is minimized.
The effect is that the lateral impact load on the anvil surface of the ram assembly 32 is eliminated, the vibrations returned to the support frame are reduced, and the longitudinal impact energy directed toward the anvil surface is increased overall.

The system is designed for long component life, easy operation, and high efficiency.

Although the invention has been illustrated and described by way of example embodiments, it should be understood that various modifications can be made by those skilled in the art without departing from the spirit of the invention as described herein.

That is, the present invention is limited only by the scope of the claims.

[Brief explanation of drawings]

FIG. 1 is a side view of the vibration shock absorber of the present invention. FIG. 2 is a partial front cross-section taken along line 2--2 of FIG. 1 showing the ram with the eccentric weight assembly exposed. FIG. 3 is a sectional view taken along line 3--3 in FIG. 2, showing the ram with one side plate removed. FIG. 4 is a bottom view of the device taken along line 4--4 in FIG. 2, showing the tool in cross section. FIG. 5 is a bottom sectional view taken along line 5--5 in FIG. 2 showing the tool holder and tool holding pin. FIG. 6 is a bottom sectional view taken along line 6--6 of FIG. 2 showing the striking plate and the lower bracket of the ram. FIG. 7 is a side sectional view taken along line 7--7 of FIG. 2 showing the guide of the ram. FIG. 8 is a partially removed front view of the apparatus showing the ram during its downward stroke engaged with a tool; FIG. 9 is a front view of the device showing the ram during its upward stroke engaged with a tool; FIG. 10 is a conceptual side view of the ram with one side plate removed to show the free movement of the ram. FIG. 11 is an explanatory diagram of a vibration-stroke response curve of a linear spring-weight system with damping. FIG. 12 is an explanatory diagram of a vibration-stroke response curve of a nonlinear spring-weight system with damping. FIG. 13 is an explanatory diagram of the time-displacement curve of the system of the present invention during nonlinear preload. Reference code 10... Support frame, 34... Impact member, ram, 109, 110, 112, 114... Elastic member, 36... Vibration drive means, eccentric weight, 107
...Blow plate, 146...Tool.

Claims (1)

Claims: 1. a support frame 10; an impact member 34 elastically attached to the support frame 10 so as to be able to vibrate relatively freely with respect to the support frame 10; a plurality of elastic members 10 attached to the impact member 34 and having the other end attached to the impact member 34;
9, 110 and 112, 114; an eccentric weight 36 driven to impart vibrational motion to the impact member 34, the eccentric weight 36 being rotatably attached to the impact member 34 and located at one position of the impact member 34; a single eccentric weight 36 positioned so as to generate a linear motion in the impact member by a certain amount; and a lower surface of the impact transmission surface, i.e., the striking plate 101, provided on the impact member 34 to be moved linearly; impact force generator. 2. The method of positioning the eccentric weight 36 so as to generate a linear motion in the impact member 34 in one position of the impact member 34 is performed by selecting the distance ds so that the following relationship holds; Where, W=total weight of the ram assembly 32 including the impact member, i.e., the ram 34, the eccentric weight 36, the hydraulic motor 72, and the balance weight 82 and moving together with the ram 34; ■=through the center of gravity of the entire ram assembly 32; and the moment of inertia of the ram assembly 32 about an axis parallel to the axis of the eccentric weight 36, ds=distance between the center of gravity and the lower surface of the striking plate 107, ie, the impact transmission surface, de-the center of gravity and the the distance between the axis and the horizontal axis, such that the nodes of the horizontal movement of the ram are located on the shock transmission surface, so that at the point of the node of the ram only vertical movement occurs and the horizontal said impact member 3 in one position of the impact member 34, characterized in that no movement occurs.
4, the eccentric weight 36
How to position.