A new operating principle of percussive electrical mechanisms.

Hadar, Anton ; Marin, Dumitru ; Szabo, Adam 等

1. INTRODUCTION

At the beginning of the XXth century it has been created the first scaling hammer called Pickhammer, actuated with compressed air (Johnson, 2002). Due to the mechanical performances (2000 strokes/minute and impact energy of 30-50 Nm/strike, with masses between 20-100 kg., the device has been used as well in mining field as in building field. The gears whose mass exceeds 25 kg, were mounted on auto-propeller carriages. The only aspect linked to the actuating mechanism (by compressive air) didn't be updated the last 100 years, diminishing a lot the technical potential of these equipments, due to the following causes: the high costs of energy, the low energetic efficiency, due to the losses on the network and the mass of the equipments. If these equipments are used in constructions where there are no networks of compressed air, moto-compressor groups must be brought, increasing the costs, the noise and pollution. Novelty which brings the proposed mechanism consists in that converts the rotational motion into a linear one, in a manner different from the crank and connecting rod assembly or pneumatic system, namely through a screw-nut mechanism with very high and not controlled stroke. The percussive mechanism presented in this paper could be a part of three different types of gears electrically actuated:

--electrically disintegrator hammer (used in mining activities, in civil and industrial buildings, at the disintegration of coal structures, raw metal, rocs, concrete, in transportable weight and size pieces);

--percussive mechanism for placing pillars of stabilization and consolidation of land became unstable;

--roto-percussive electric drill (allows the execution of all current maintenance, repair, renovation, conversions works);

2. DESCRIPTION OF THE PERCUSSIVE AND ROTO-PERCUSSIVE MECHANISM

In figure 1 is presented a percussive mechanism applied to an electric disintegrated hammer. It consists of cylindrical body 3. inside which rotates the axle 4, trained by electric motor 1.

On the splined portion of the axle 4 is mounted the coupling-decoupling bushed bearing 5. It turns simultaneously with the axle 4, having the possibility to slip over the splined portion in the limit of 5 mm.

[FIGURE 1 OMITTED]

At the inferior part of the axle 4 is mounted the bushedbearing 6, free on the axle, fitted with a coarse pitch exterior acme thread (90 mm). As well the coupling-decoupling bushed bearing 5 as the screw one 6 have indented contact surfaces, which interpenetrate, being maintained in contact by a helicoidal spring 8. The bushed bearing 6 is screwed on the piston 7 fitted with an interior acme thread compressing the exterior spring 9 with about 45 mm (Fu et al., 1997).

The piston 7 raise the bushed bearing 5 and uncouple it from the mechanical sleeve 6, which became free on the axle 4, permitting the violent detent of the spring 9; the piston 7 is projected on the head pickax 10 remising all the kinetic energy. During the spring detent, the bushed bearing 6 turns free, the spring 8 rebounds, recoupling the bushing sleeve 5 and 6 the cycle being recurred. Because the piston must not be carrying in a rotational motion, it is fitted with two ordinary keys which glide in channels carried out in the cylinder wall 3 (Harrison, 1997).

3. CALCULUS AND CONSIDERATIONS ON THE OPTIMIZATION OF THE EQUIPMENT CHARACTERISTICS

The initially electrical engine taken into consideration for the equipment driving has the main characteristics:

--the engine power: P=3kW;

--the engine speed (the percussion frequency): n=1000-3000 rot./min.

Other designing data: the piston mass m=0,6 kg.; the piston stroke [DELTA]x=45 mm, the pitch of the driving thread p=90 mm. Starting with these data, and considering an engine speed of 3000 rot/min, it results the equipment service condition:

[t.sub.1] + [t.sub.2] [less than or equal to] 0,02 s (1)

where [t.sub.1] and [t.sub.2] represent the climbing and descending time of the piston 7 (period of an operating cycle).

The elastic constant of the spring is:

k = G x [d.sup.4] / 8 x [D.sup.3] x I = 8.4 x [10.sup.4] x [5.8.sup.4] / 64 x [56.sup.3] x 3 = 22,55N/mm (2)

where: G--the shear modulus (8,4 x [10.sub.4] MPa);

d--the diameter of the spring wire;

D--diameter of spring involution;

i--number of spires.

The oscillation pulsation [[omega].sub.1] of the spring is:

[[omega].sub.1] = [square root of k/m] = 193,86 rad./s (3)

In these conditions, the forces which act upon the spring in the two positions (pretension with [x.sub.1] = 20 mm and final compression with [x.sub.2] = 65 mm) will be:

[F.sub.1] = k x [x.sub.1] = 22,55 x 20 = 451N (4)

[F.sub.2] = k x [x.sub.2] = 22,55 x 65 = 1465,75N. (5)

The mean force will be: or the considered spring, being in an oscillatory motion, the frequency of this movement will be:

[F.sub.med] = 451 + 1465,75 / 2 = 960N. (6)

[n.sub.1] = [[omega].sub.1] / 2[pi] = 30,853osc./s x 60 = 1851,22osc./min (7)

Using the theorem of mechanical energy conservation for a mechanical oscillator, one could obtain the maximum speed oscillation [[upsilon].sub.max ]of the spring (Mobley, 1999):

[v.sub.max] = [square root of k[A.sup.2] / m] = 8,72m/s (8)

where A is the oscillation amplitude equal to the stroke value (45 mm). The maximum acceleration [a.sub.max] of the oscillatory motion will be:

[a.sub.max] = [[omega].sup.2] A = [193,86.sup.2] x 0,045 = 1691m/[s.sup.2] (9)

A value of the same quantity very close to the value obtained above will be obtained using the mean force:

[a.sub.max] = [F.sub.med] / m = 1600m/[s.sup.2] (10)

The climbing time of the bushed bearing (of the spring compression) will be (Routh, 2003):

[t.sub.1] = [DELTA]x / p ([t.sub.1] + [t.sub.2]) = 45/90 x 0,02 = 0,01s (11)

The descending time of the bushed bearing, calculated through three different procedures (uniform motion, uniformly variable motion and harmonic motion) is situated between 5,16 x [10.sup.-3]s and 8,1 x [10.sup.-3]s:

[t'.sub.2] = [DELTA]x / [v.sub.max] = 0,045 / 8,72 = 5,16 x [10.sup.-3] s; (12)

[t".sub.2] = [square root of 2 x [DELTA]x / [a.sub.max]] = [square root of 2 x 0,045 / 1600] = 7,5 x [10.sup.-3] s (13)

[t'".sub.2] = [pi] / 2 x [[omega].sub.1] = [pi] / 2 x 193,86 = 8,1 x [10.sup.-3] s. (14)

Consequently, the total time (climbing and descending) will be:

[t.sub.min] = 0,01 + 5,16 x [10.sup.-3] = 0,01516 s. (15)

[t.sub.max] = 0,01 + 8,1 x [10.sup.-3] = 0,0182 s. (16)

The moment M, the power P and the impact energy E of the oscillatory system will be:

M = [F.sub.med] x tg[theta] x [D.sub.med] / 2 = 960 x tg38 [degrees]30'x0,017 = 12,98 = 13Nm (17)

P = M x [omega] = 13 x 193,86 = 2520,18W (18)

[E.sub.1] = 1 /2 k[A.sup.2] [congruent to] 23 Nm; (19)

where [D.sub.med] represent the mean diameter of the acme thread 38 mm and [theta], the lead angle.

The power imposed by the engine characteristics is:

[P.sub.3000] = P x [omega]' = 13 x [pi] x 3000 / 30 [congruent to] 4000W (20) [P.sub.2000] = P x [omega]" = 13 x [pi] x 2000 / 30 = 2722W

One could remark that for an engine with 3000 rot/min the corresponding power will be 4 kW, leading to a too heavy mass of the engine. We conclude that an adequately engine should have a power of 3 kW, with a rotative speed of 2000 rot/min. In this situation, [t.sub.1] and [t.sub.2] are covered, since [t.sub.1] + [t.sub.2] [less than or equal to] 0,03.

4. CONCLUSIONS

The use of the percussive mechanism presented in the paper leads to a performance equipment having a mass substantially reduced (8-9 kg) with technical and functional characteristics superior face to those existent. Impact energy can be increased several fold or decreased by simply changing the main spring without changing the diameter of the cylinder and piston. The blows frequency increases greatly, because for this variant of producing impulses, there is a possibility that at a single axis of rotation of the turning arbor could have two or even three shots out of the drilling or disintegrated tool.

5. REFERENCES

Fu, K.S., Gonzalez, R.C. & Lee, C.S.G. (1997). Robotics, McGraw-Hill

Harrison, H.R. (1997). Advanced Engineering Dynamics, John Wiley & Sons Inc., New York

Johnson, W. (2002). Impact strength of materials, Edward Arnold

Mobley, R.K. (1999). Vibrations Fundamentals, Newnws, Boston

Routh, E.J. (2003). Dynamics of a system of rigid bodies, Part 1 and Part 2, Macmillan