TEST RESULTS OF RECOIL SHOCK ABSORBERS FOR HANDHELD PNEUMATIC HAMMERS

V. I. BABUROV, I. I. TRUSHAKOV, A. V. PIKUL, A. N. SHIPUNOV

(Submitted by the Department of Mining Machines, Mine Transport, and Mining Mechanics)

The vibration of handheld pneumatic hammers is generated by various forces, among which the principal ones are: the action on the housing of compressed-air pressure varying in magnitude and time, and impacts of the housing with the working tool. The intensity with which these vibrational factors manifest themselves depends on a great many causes and, first of all, on the magnitude of the worker’s pressing force. Investigations carried out at our institute have established that, in the range of forces often encountered in chipping and riveting practice, the most dangerous are the limiting basic operating mode of the hammers [1—3] and the operating mode at a pressing force of 30—34 kg (including the weight of the hammer) [3]. For chipping hammers of the MP type manufactured by the Tomsk Electromechanical Plant, the limiting basic mode occurs at a pressing force equal to 22—24 kg, and for riveting hammers of the KE type from the same plant, 24—26 kg [1, 3]. The given values of pressing force correspond to the operating conditions of the machine at an air pressure in the network of 5 atm. In the mode under consideration, the hammer housing runs up against the limiting collar of the working tool and collides with it.

The reactive force acting on the housing during this collision causes an increase in the amplitude of vibration displacement; in this connection the amplitude–force characteristic of the hammer [1] has a maximum in this mode. When the pressing force is equal to the maximum value of the internal disturbing force for hammers of the KE type, i.e., 34—35 kg, the housing does not run up onto the collar. In this case the tool, rebounding from the material being worked, itself strikes the hammer, thereby causing a new increase in the vibration displacement of the housing [1—3]. For hammers of the MP type, the mode mentioned is characteristic at a pressing force of 28—32 kg.

Thus, collisions of the housing with the working tool are dangerous for service personnel [3, 4, 5]. Moreover, the collisions are of an impact nature. Over a short interval of time, 0.0012—0.0015 sec, the collision force reaches 120—180 kg or more.

Damping of collision forces is usually accomplished by installing elastic elements made of rubber, disk springs, etc., between the limiting collar of the tool (chisel, rivet set, or pick) and the hammer housing. The applied arrangements of the elastic elements may be divided into two groups:

a) the elastic element is located outside the hammer (Fig. 1, a);
b) the elastic element is located in the front part of the hammer (Fig. 1, b).

In the first case, elastic element 4 is placed between the barrel 1 of the hammer and the tool 5. As element 4 wears, the worker can easily replace it without spending much time. However, frequent replacement of the working tool creates certain inconveniences in the operation of hammers with the aforementioned shock absorbers. In addition, the latter are not protected from various mechanical effects and therefore quickly become unusable.

In the second case, shock absorber 4 together with the working tool 5 is held by nut 6, which is screwed onto barrel 1.

To reduce the impact action of the tool on the hammer body, a special vibration-damping bushing was developed at our institute (Fig. 1, c). Vibration-damping bushing 3 is inserted into the cavity of the front part of barrel 1. By means of its annular projection, the bushing rests on spring 4 and rubber sleeve 7. The bushing is prevented from falling out by nut 6. Spring 4 serves to absorb collisions between tool 5 and barrel 1. Sleeve 7 is intended to reduce the impact action of the striker on the body, which occurs when the shank of the chisel leaves the bushing. The advantages of the proposed shock-absorption scheme are as follows:

a) the shape of the shank of the working tool can be arbitrary, which is very important in operation;

Fig. 1. Installation schemes of recoil shock absorbers.
1 — barrel; 2 — striker; 3 — guide bushing; 4 — recoil shock absorber; 5 — working tool; 6 — front nut; 7 — rubber sleeve

b) installation of the shock-absorbing elements is possible without increasing the dimensions and weight of the hammer;

c) in addition to reducing collisions between the tool and the body, the hammer barrel is protected from destruction by impacts of the striker when the tool leaves the cavity;

d) simplicity of design and ease of operation;

e) as it wears, the bushing can easily be replaced;

f) the elastic elements are protected from mechanical effects, which helps increase their service life.

It is known that the shock-absorption effect, all other conditions being equal, depends on the type and elastic properties of the shock-absorbing elements [4, 6—9]. To determine the effectiveness of damping hammer vibration from reverse impacts of the tool as a function of the type of elastic elements, the authors conducted the corresponding tests.

The following were installed as shock absorbers:

a) a solid rubber bushing (Fig. 2, a);
b) a rubber bushing with grooves on the outer surface (Fig. 2, b);
c) rubber rings with steel washers (Fig. 2, c);
d) a steel spring (Fig. 2, d).

The dimensions of the elastic elements in the free state are shown in Fig. 2. Before the experiments were carried out, the stiffnesses of all the elements were determined. The deformation was measured by loading the shock absorbers placed in the front cavity of the hammer. Graphs of the variation in deformation as a function of the load applied to the element are given in Fig. 3, and the mean values of their stiffness are given in Table 1. The tests were carried out under the following conditions.

Fig. 2. Types of shock absorbers and their dimensions

The MP-5 chipping hammer (impact energy—1.2 kgm, blow frequency per minute—2200, weight—5 kg) was mounted on a stand in an inclined position [10]. The material being worked was a steel plate (St. 3), and the working tool was a chisel weighing 0.83 kg with a cutting edge in the form of a double-sided wedge with a point angle of 40°. The air pressure in the line was maintained at 5 atm. The pressing force, produced by a pneumatic piston feed device, was determined by the formula

\[ R_{\mathrm{n}} = F_{\mathrm{p}} \cdot P_{\mathrm{p}} + Q_{\mathrm{m}}\sin \alpha,\qquad \text{kg}, \tag{1} \]

where \(R_{\mathrm{n}}\) is the pressing force on the hammer, kg;
\(F_{\mathrm{p}}\) is the cross-sectional area of the feed-device piston, equal to \(20 \text{ cm}^2\);
\(P_{\mathrm{p}}\) is the air pressure in the cylinder of the feed device, set equal to \(1.4 \text{ kg/cm}^2\);
\(Q_{\mathrm{m}}\) is the weight of the hammer, kg;
\(\alpha\) is the angle of inclination of the hammer to the workpiece, equal to \(50^\circ\).

Substituting the indicated values into formula (1), we obtain a pressing force equal to 32 kg.

Table 1

Data from processing diagrams of recoil force and displacement of the body of the MP-5 handheld pneumatic hammer

Using electronic equipment—an MPO-2 oscillograph and a PET-3VM strain-gauge amplifier—the vibrational displacement and recoil force of the MR-5 hammer were recorded. The displacement was recorded by a potentiometric

Fig. 3. Effect of load magnitude on deformation of recoil shock absorbers.
1—steel spring; 2—rubber bushing with cuts; 3—solid rubber bushing; 4—rubber rings with steel washers

sensor [11], and the recoil forces by a cantilever strain-gauge sensor [10]. Thus, we determined the following vibration parameters of the hammer:

a) \(A_k\)—amplitude of vibrational displacement, mm;
b) \(F_{\text{o.y.}}\)—maximum force of the recoil impact of the working tool against the housing, kg (Fig. 4, б);
c) \(t_{\text{o.y.}}\)—duration of the force impulse produced by the recoil impact of the tool against the housing, sec (Fig. 4, а).

From the last two vibration parameters, the magnitude of the impulse caused by the recoil impact was determined by the formula

\[ I_{\text{o.y.}}=\int_{0}^{t} F_{\text{o.y.}}\,dt \approx \frac{1}{2}F_{\text{o.y.}}\,t_{\text{o.y.}}, \quad \text{kg·sec}. \tag{2} \]

The diagrams of recoil force \(R_0\) and housing displacement \(S_{\mathrm{h}}\) for each type of shock absorber tested are shown in Fig. 4. The amplitude of housing displacement under the action of the rebound impact was calculated from the expression

\[ A_{\mathrm{k}} = \frac{(a'_1 + a''_1) + (a'_2 + a''_2) + \ldots + (a'_n + a''_n)}{2n} = \frac{1}{2n}\sum_{i=1}^{i=n}(a'_i + a''_i),\ \mathrm{mm}, \tag{3} \]

where \(a'\) is the displacement of the housing toward the handle in one oscillation caused by the rebound impact of the tool, mm (Fig. 4, c);
\(a''\) is the displacement of the housing toward the tool in the same oscillatory motion, mm.

Fig. 4. Diagrams of recoil force and housing displacement of the MP-5 chipping hammer.
a) with a steel spring; b) with a rubber bushing with grooves; c) with a solid rubber bushing; d) with rubber rings with steel washers.

From Fig. 4 it is evident that, regardless of the type of shock absorber installed in the hammer, the recoil-force diagram \(R_0\) (Fig. 4, c) has the following characteristic features. The diagram \(R_0\) has several clearly expressed peaks (zone 1, Fig. 4, c). This indicates that the tool, rebounding from the material being worked, collides with the body not once but two to four times. Under the action of these impacts, and especially the first, the hammer body overcomes inertia and changes the direction of its motion (the curve \(S_{\text{body}}\) deflects upward). In zone 2 the force acting on the handle decreases somewhat. This is caused by the pressure of the air in the front cavity on the walls of the body. Under the action of the pressing force the body approaches the tool. In zone 3 the recoil force again increases, since compressed air enters the rear cavity of the hammer. Under the action of the internal disturbing force, whose value is greater than the magnitude of the pressing force, the hammer body begins to move toward the handle. Subsequently the air pressure in the rear cavity decreases, the recoil force decreases, and the body, under the action of the pressing force, rushes toward the tool (zone 4).

Thereafter the character of the change in recoil force and body displacement is repeated. Such a periodic change in vibration parameters is characteristic of all types of shock absorbers installed in the hammer. Table 1 gives the data from processing the diagrams of recoil force \(R_0\) and body displacement \(S_{\text{body}}\).

Thus, in the course of the tests it was established that, as the stiffness of the elastic element increases, the amplitude of body displacement and the maximum recoil force caused by reverse impacts of the tool increase. The duration of the collision of the hammer with the chisel decreases in this case. For example, when the stiffness of the shock absorber is increased almost 10 times, the reverse-impact force increases 3 times, while the time of its action becomes 27% shorter.

The installation of a shock-absorbing element makes it possible to reduce substantially the force of shocks caused by rebounds of the working tool from the product being worked and, what is especially important, to increase the duration of action of these shocks. This improves the working conditions of the operators [12].

The test results show that the most effective shock absorber is a steel helical spring. However, when selecting the type of elastic element, it is necessary to keep the following circumstance in mind. If the stiffness of the element is small, then during operation of the hammer the point of collision of the striker with the shank of the working tool will shift relative to the barrel, which will disrupt the working process of the machine. The hammer will operate unstably, which is highly undesirable, since its productivity will decrease. This phenomenon was in fact observed when a steel spring with a stiffness of 20 kg/cm was installed as the shock absorber.

When a rubber bushing with grooves on the outer surface was installed, whose stiffness is equal to 56 kg/cm, the chipping hammer operated already with sufficient stability. It follows from the foregoing that for chipping hammers of the MP type the most acceptable shock-absorber stiffness should be considered to be 50–70 kg/cm.

Conclusions

1. The recoil force caused by collisions of the hammer with the working tool can be reduced by a factor of 2–3 by installing shock-absorbing elements with a stiffness of 50–70 kg/cm in the front part of the machine (handheld pneumatic hammers weighing 4–7 kg).

1. The installation of elastic elements between the working tool and the body of a hand-held pneumatic hammer makes it possible to reduce the displacement amplitude by a factor of 1.5–2.0, which has a favorable effect on the health of workers.

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