Some Results of an Experimental Study of the Rack-and-Screw Mechanism of an Automatic Feed Device for Rock Drills

O. D. Alimov, D. N. Malikov, I. V. Rodionov

The results of a theoretical study of the mechanism of an automatic feed device for drilling machines were presented by us in [1]. This article gives some results of the experimental study.

For the purpose of experimental investigation, the automatic-feed mechanism was manufactured as an independent unit. In the investigation the aim was to establish the deviations of the experimental data from the theoretical data and to identify the causes of these deviations, as well as to assess the operability of the automatic feed device in drilling rocks.

Test Method

To measure the torque on the drill bit \(M_b\), the braking torque on the lead screw \(M_t\), and the feed force \(F\), spring sensors were used. The spring sensors were cylindrical helical tension–compression springs. The stiffness and length of these springs were selected by calculation so that they could provide the necessary sensitivity over a specified range of loads.

The basic diagram of the experimental setup is shown in Fig. 1. The feed force was determined by the degree of compression of spring 4 under the action of the reactive axial forces arising on lead screw 3 during operation of feed mechanism 1. The feed-force record was made on paper tape 7 by pencil 6, fixed in lever 5, which was in contact with the collar of the lead screw. When sensor spring 4 was compressed, the collar of the lead screw moved to the left and deflected lever 5 with pencil 6 relative to the fixed guides 10. When spring 4 was unloaded, the collar of the lead screw moved to the right. In this case, return spring 8, whose stiffness was chosen to be less than the stiffness of spring sensor 4, served to maintain contact between lever 5 and the screw collar. The load on the drill bit and on the lead screw was created by means of brakes 2 and 9.

The arrangement for calibrating and recording the torques on the drill steel and the lead screw is shown in Fig. 2. To one of the shoes 2 of brake 1 a lever 3 was welded, to the end of which a sensing spring 11 was attached. The other end of spring 11 was fixed immovably.

Fig. 1. Schematic diagram of the experimental setup.

By means of string 5, the end of lever 3 was connected to lead 7, carrying pencil 8.

In the absence of braking torque, the neutral position of levers 3 and 7 was fixed by stop 10. When the braking torque on the drill steel (or the lead screw) increased, lever 3 and the end of lead 7 connected with it deflected to the right. When the braking torque decreased, lever 3 moved in the opposite direction under the action of sensing spring 11, and lead 7 under the action of return spring 6.

The sensing springs of the brakes were calibrated after installation of the setup by suspending loads 4 of various weights; in this case the calibration line was drawn automatically by pencil 8 on the moving tape 9. After calibration, the loads 4 were removed.

The recordings of the torques and the feed force were made on one common paper tape. In this case all pencils were arranged on one line, perpendicular to the direction of motion of the tape, which made it possible to align in time the processes of variation of the recorded parameters. The paper tape was drawn at a constant speed \(V_l = 500\ \text{mm/min}\).

Fig. 2. Arrangement for calibration and recording of torques on the drill steel and the lead screw.

Figure 3 gives a record obtained during calibration of the torques on the drill steel \(M_b\). The numbers along the calibration li-

They express torques (in kg·cm). From these records, calibration curves were constructed for determining the torques on the drill bit \(M_b\). In the same way, calibration curves were obtained for determining the feed force \(F\) and the braking torque on the screw \(M_t\). The power consumed by the feeder motor was recorded by a self-recording wattmeter.

Fig. 3. Record obtained during calibration of torques on the drill bit.

Tests of the feeder without a rock drill

By clamping the brake shoes 9 (Fig. 1), a specified braking torque was created on the screw \(M_t\). The feeder motor and the mechanism for pulling the paper tape were switched on. With a given constant braking torque on the screw, by clamping the brake shoes of the drill-bit brake 2, the torque on the drill bit \(M_b\) was periodically increased from 0 to the value at which the feed force \(F\) decreased to 0. Then the torque on the drill bit was periodically reduced to its initial value. The tests were repeated at different values of \(M_t\). A typical diagram recording \(M_t\), \(M_b\), and \(F\), as well as the corresponding diagram of the power \(P_1\) consumed by the feeder motor, is shown in Fig. 4. The numerical values of the parameters \(M_b\), \(M_t\), and \(F\) on the diagrams were determined from the calibration curves, and the power \(P_2\) on the motor shaft from the typical load characteristic of the feeder electric motor.

The results of the experimental study of the feeder without a rock drill are presented in Fig. 5, from which it follows:

1. An increase in the torque on the drill bit \(M_b\) causes a decrease in the feed force \(F\), and conversely.

2. As the braking torque on the screw \(M_t\) increases, the maximum values of the torque on the drill bit \(M_{b \max}\) and of the feed force \(F_{\max}\) increase.

3. At a constant braking torque on the screw \((M_t=\mathrm{const})\):

a) changing the torque on the drill bit from 0 to \(M_{b \max}\) causes a nonproportional change in the feed force \(F\), and the function \(F=f(M_b)\) is represented by a curved line, for example curve 1;

b) changing the torque on the drill bit from \(M_{b \max}\) to 0 causes a proportional change in the feed force \(F\), and in this case the function \(F=f(M_b)\) is a straight line (line 2).

Thus, the function \(F=f(M_b)\) is not single-valued and is expressed graphically by the curve \((1—2)\), which has the form of a loop. The shape and size of the loop characterize the degree of insensitivity of the feeder mechanism to self-regulation of the parameters \(M_b\) and \(F\). The formation of the loop is explained by a redistribution

Fig. 4. Typical recording diagram.

of forces and moments in the automatic-feed mechanism, depending on the nature of the change in the torque on the drill bit.

Fig. 5. Results of an experimental study of the feeder without a rock drill.

In the course of the experiments it was established that, with a slow change in the torque on the drill bit from 0 to \(M_{b\max}\), the loop narrows and curve 1, straightening, approaches straight line 2, whereas with a rapid

of variation of the torque, curve 1 is a broken line; the loop takes the form of a triangle.

Analysis of this circumstance makes it possible to conclude that the sensitivity of the mechanism under study to self-regulation of the parameters \(M_b\) and \(F\), at a constant braking torque on the screw \((M_t = \text{const})\), depends primarily on the rate of change of the torque on the drill bit \(M_b\) and on the ratio of the values \(M_b\) and \(F\). Thus, the mechanism has the greatest capacity for self-regulation when the load on the drill bit changes slowly and when operating under such regimes in which large torques on the drill bit are produced by small feed forces during drilling.

Fig. 6. Distribution of the power consumed by the feeder.

The power \(P_2\) consumed by the feeder is expended on rotation of the drill bit \(P_b\), on braking the screw \(P_t\), and on overcoming friction in the links of the mechanism \(P_m\). The pattern of distribution of the power \(P_2\) among the indicated elements at \(M_t = 310\ \text{kg·cm}\) is presented in Fig. 6. The values of \(P_t\) and \(P_b\) were calculated by the formula \(P = 0.0010267 M_n\). The numbers of revolutions per minute of the screw \(n_v\) and of the drill bit \(n_b\) were determined at a feeder motor load \(P_2 = 1.5\ \text{kW}\), with subsequent correction for rotor slip at the corresponding load. The limits of fluctuation of the quantities \(n_v\) and \(n_b\) were insignificant.

Analysis of the pattern of distribution of the power \(P_2\) makes it possible to draw the following conclusions:

1. The sum of the losses due to friction in the links of the mechanism and due to braking of the screw, with \(M_t = \text{const}\) and with a variable load on the drill steel, remains approximately constant.

2. The useful power \(P_b\) expended on rotating the drill increases in proportion to the magnitude of the torque on the drill. But in the drilling process, at a constant braking torque on the screw, the useful power can increase only up to a certain limit. This limit is bounded by the zero value of the feed force, since a further increase in the load on the drill causes the feeder to move away from the face and, consequently, the drill to be unloaded. Therefore the highest efficiency of the feeder, corresponding to the maximum value of the torque on the drill and the zero value of the feed force, does not exceed 55%. At average ratios of the torque on the drill and the feed force, the efficiency is of the order of 30–40%. The experimentally obtained values of the feeder efficiency agree well with the results of the theoretical investigation [1].

Tests of the feeder during drilling

When testing the feeder together with the operation of the rock drill, the brake of the drill was removed. A pneumatic rock drill of type OM-506 was attached to the feeder. Rotation and feeding of the drill against the face were carried out by the feeder, while blows on the drill were delivered by the rock drill. Drilling was performed in rock of strength \(f = 12 \div 14\) according to the scale of Prof. M. M. Protodyakonov. To record the feed rate during drilling, the carriage with the pencil, which had previously been connected to the drill brake, was connected to the feeder body by means of a rod.

The tests were conducted in the following order. By clamping the brake shoes, a small torque \(M_t\) was created on the screw. Compressed air at a pressure of 4 atm was supplied to the rock drill, and drilling was carried out for 30 sec. After this, the braking torque was periodically increased and the process was repeated until the load on the feeder motor reached the limiting value. Drilling at air pressures of 5 and 6 atm was carried out in the same order. Approximate diagrams recording the braking torque on the screw \(M_t\), the feed force \(F\), and the feed rate \(V\) are presented in Fig. 7.

The determination of the torques \(M_b\) arising on the drill during drilling was carried out indirectly, which is explained by the absence of suitable recording apparatus.

Let us denote by \(P_c\) the power used to overcome harmful resistances (it represents the sum of the power expenditures for braking the screw and for friction in the links of the automatic feeder mechanism). The power \(P_c\) was determined as follows. From the data of experiments in which there were no useful power expenditures on drilling (\(M_b = 0\)), a graph of the dependence was constructed

\(P_{\mathrm{c}} = f(M_{\mathrm{t}})\), where \(P_{\mathrm{c}} = P_2\) is the power consumed by the feed device. From the value of the specified braking torque on the screw \(M_{\mathrm{t}}\), in drilling experiments, using the previously obtained relationship, the power \(P_{\mathrm{c}}\) going to overcome harmful resistances during drilling was determined. In this, the additional power losses due to friction in the mechanism, caused by loading of the drill, were not taken into account. The power expended on rotating the drill under load was determined as the difference between two quantities

\[ P_{\mathrm{b}} = P_{\mathrm{a}} - P_{\mathrm{c}} \ \text{kW}. \]

Then

\[ M_{\mathrm{b}} = 97\,400 \frac{P_{\mathrm{b}}}{n_{\mathrm{b}}} \ \text{kg·cm}, \]

where \(n_{\mathrm{b}}\) is the number of revolutions of the drill per minute.

Fig. 7. Approximate recording diagrams of the braking torque on the screw \(M_{\mathrm{t}}\), the feed force \(F\), and the feed speed \(V\).

The results of the experimental study of the automatic feed device during drilling are summarized in the graph of Fig. 8. The solid lines show the interdependence of the torques on the drill \(M_{\mathrm{b}}\) and the feed forces \(F\) at different braking torques on the screw and different operating modes of the drilling hammer. By the operating mode of the drilling hammer in this case is meant its operation at a definite compressed-air pressure. Obviously, any point in the area bounded by the coordinate axes characterizes the relationship between \(M_{\mathrm{b}}\) and \(F\) and is determined by the following factors: a) the magnitude of the braking torque on the screw; b) the operating mode of the drilling hammer; c) the physico-mechanical properties of the rock; and d) the design of the drill bit.

If the feed device operates with a constant braking torque \((M_{\mathrm{t}} = \mathrm{const})\), then the dependence \(F = f(M_{\mathrm{b}})\) can be expressed

by a straight line. Similar dependences in Fig. 8 are denoted by the index \(M_{\mathrm{t}}\).

Let us consider the process of self-regulation of the parameters \(M_b\) and \(F\) in the following example. Suppose that, with a constant braking torque on the screw \(M_{\mathrm{t}} = 300\ \mathrm{kg\,cm}\), drilling is carried out at an air pressure in the hammer of from 4 to 6 ati. In this case the dependence \(F = f(M_b)\) is represented in Fig. 8 by the straight-line segment \(AB\). At an air pressure of 4 ati (point \(A\)), \(M_{b1} = 220\ \mathrm{kg\,cm}\) and \(F_1 = 235\ \mathrm{kg}\), while at a pressure of 6 ati (point \(B\)), \(M_{b2} = 310\ \mathrm{kg\,cm}\) and \(F_2 = 203\ \mathrm{kg}\).

Thus, the change in air pressure in the hammer caused an increase in the torque on the drill. In turn, the increase

Fig. 8. Results of an experimental study of the automatic feed device during drilling.

in torque on the drill led to a decrease in the feed force. If the feed force had not changed, as occurs in ordinary feed devices, then the torque on the drill would have increased to \(M_b^{1} = 375\ \mathrm{kg\,cm}\) (point \(C\)). It follows from this that the process of self-regulation of the parameters \(M_b\) and \(F\) leads to a certain equalization of the torques arising on the drill under different drilling conditions. It should be emphasized that the more varied the operating conditions of the drill are, the more clearly this property of the automatic feed device will be manifested and, consequently, its advantage over feed devices used in practice. Thus, for example, in the event that the drill becomes jammed, the torque increases to \(M_{b\max}\), which will cause the drill to be withdrawn from the face. With an ordinary feed device, however, jamming of the drill leads either to stoppage of the drilling hammer or to breakage of the drill.

Fig. 9. Nomogram characterizing the operation of the feed device at
\(z'_7 = 30,\ z'_9 = 67\), and \(n'_4 = 126\) rpm.

Fig. 10. Nomogram characterizing the operation of the feed device at
\(z'_5 = 55,\ z'_6 = 42\), and \(n'_4 = 194\) rpm.

CONCLUSION

As a result of theoretical [1] and experimental studies of the rack-and-screw mechanism of the automatic feed device, the following conclusions may be drawn:

1. In the drilling process, the automatic feed mechanism ensures (without the presence of special regulating equipment) self-regulation of the interdependent parameters: the torque \(M_b\) arising at the drill and the feed force \(F\), up to a negative value of \(F\) when the drill jams.

2. Self-regulation of the parameters \(M_b\) and \(F\) leads to equalization of the torques arising at the drill under different drilling regimes, which improves the operating conditions of the drilling machine and creates a more stable load on the feed-device motor. This property of the automatic feed mechanism can be used with great effect in rotary and rotary-percussive drilling machines.

3. The principal shortcoming of the feed device under investigation is the considerable expenditure of power on braking the screw. As a result, the efficiency of the feed device is on average 30–40%. Modern drilling hammers operate most effectively in the regime \(M_b = 150 \div 200\ \text{kg·cm}\) and \(F = 50 \div 200\ \text{kg}\).

The indicated parameters are provided by the automatic feed device at a reactive (braking) torque on the screw \(M_t = 100 \div 200\ \text{kg·cm}\). In this case, a power of \(0.33\)—\(0.65\ \text{kW}\) is expended on braking the screw; being converted into thermal energy, this causes strong heating of the braking device.

What has been said above raises the question of the need for further research with the aim of reducing the power expended on braking the screw or replacing friction braking with some other useful type of loading device, for example a pump.

Analysis of the calculated relationships obtained as a result of the theoretical investigation of the operation of the automatic feed mechanism [1] makes it possible to conclude that a considerable reduction in unproductive power losses can be achieved:

a) by changing the gear ratio \(\dfrac{z_7}{z_9}\), or b) by carrying out mutual braking of gear wheels 4 and 5. In the first case, for \(z'_7 = 30\), \(z'_9 = 67\), and \(n'_4 = 126\ \text{rpm}\), the nomogram characterizing the operation of the feed device takes the form shown in Fig. 9; in the second case \((z'_5 = 55,\ z'_6 = 42\) and \(n'_4 = 194\ \text{rpm})\)—that shown in Fig. 10.

LITERATURE

1. P. M. Alabuzhev, O. D. Alimov, I. V. Rodionov, D. N. Malikov. Investigation of the rack-and-screw mechanism of the automatic feed device of an electro-pneumatic drilling hammer. Article in the present collection.