Some Issues in the Design of Electropneumatic Impact Machines

P. M. Alabuzhev

Certain problems in the dynamics of machines as applied to design tasks have been posed and developed in the works of Academician I. I. Artobolevsky [14, 15, 16, 17]. In one of these works it is stated: “It is necessary to pose questions connected with the experimental study and investigation of the mechanical characteristics of prime movers, working machines, and machine units. Finally, in order to create engineering methods for calculating machine units, it is necessary to develop approximate methods for integrating the equations of motion” [16]. This statement, to a certain extent as applied to electric drilling hammers, has been reflected in our works [1, 6, 11].

Academician E. A. Chudakov notes that, in order to create a correct picture of the working process, reliable methods of calculation, and, ultimately, a good machine, it is necessary to raise the culture of experimentation and the technique of measurement; it is necessary “to continue improving all automatic measuring equipment, both with respect to dimensional accuracy and to the quality of the metal and the machining of parts” [41, p. 176]. It is necessary to study the working process, the dynamics of machines, and the strength of its elements under conditions close to production conditions. For this purpose, experimental study of pressures and stresses in machine parts is recommended.

Along with the study of various schemes of mechanisms with the aim of identifying the most rational of them in a dynamic respect [14], further improvement is needed in methods for calculating the operating modes of the machine, as well as methods for strength calculation (with the use of statistical methods for analyzing and processing the materials obtained).

The principal questions of the development of the mining industry are covered in works [22, 32, 35, 42, 43]. Specific questions concerning exploratory drilling equipment, increasing the productivity and durability of impact machines, are reflected in works [25, 29, 33, 37, 42, 46]. Certain questions of the theory and problems of designing pneumatic and electropneumatic—

... impact machines are posed in works [4, 7, 10, 18, 19, 20, 23, 26, 36, 40, 44].

On the basis of the literature reviewed, the main requirements imposed on the design of new machines may be briefly formulated as follows: the machine must be productive, economical, and reliable in operation. With sufficient strength of individual assemblies and elements, the machine as a whole must have minimum weight and dimensions. The design of the machine must satisfy the conditions of manufacturability (expediency and simplicity of individual elements, choice of material and method of blank production, choice of manufacturing and machining accuracy, unification and standardization of parts, cost of the machine, etc.).

Additional requirements are imposed on mining machines operating underground under regimes of heavy impact loads: the machine must be especially strong, transportable, sufficiently powerful, and have power reserves; furthermore, it must be explosion-proof and protected against damage (gas, dust, rust) [35, 38, 39].

On the basis of the investigations carried out, we believe that at present the necessary data are available for creating satisfactory designs of electric impact machines, in particular with the use of air as the coupling between the piston and the striker. Here it is possible to implement designs according to various schemes: with a one-sided air coupling, a two-sided air coupling, pneumatic pulsators, and gearless hammers with a rigid coupling on the return stroke [12, 19] (see figure).

As manual impact machines with impact energy \(A_{б} = 1 \div 4\ \text{kgm}\), it is most expedient to use gearless hammers and pneumatic pulsators, as well as designs of the EBMG-10 type hammer with a separately mounted electric motor. To obtain an impact energy of \(A_{б} = 1 \div 10\ \text{kgm}\), it is comparatively simple to implement the design of an electropneumatic hammer with a one-sided air coupling between the striker and the piston. For especially powerful impact machines with impact energy \(A_{б} = 8 \div 15\ \text{kgm}\) (and higher), it is more expedient to use machines based on the scheme of a two-sided air coupling.

The desire to make an electropneumatic drilling hammer necessarily light (hand-held) limited the weight and power of the electric motor. To create an electropneumatic drilling hammer equivalent in the main indices (energy and number of impacts, drilling speed) to modern hand-held pneumatic drilling hammers, it is necessary to equip it with an electric motor of \(3\text{–}4\ \text{kW}\) power. If one takes into account that the weight of such an electric motor will be \(30\text{–}50\ \text{kg}\) and that in the hammer design it constitutes about half of the total weight, the difficulties of creating a hand-held...

of an electropneumatic drilling hammer. Reducing the weight of the hammer by reducing the weight of individual parts, and consequently also lowering their margin of strength, does not lead to the required results.

Diagrams of electropneumatic impact machines:

a—electropneumatic single-cylinder hammer of single action with an air connection; b, v, zh—light hand hammers operating on the same principle; g, d, e—electropneumatic impact machines with double action of the air connection; z—gearless hammer with a rigid connection between the piston and the striker during the return stroke of the striker; i—the cutting part of a coal cutter with electropneumatic hammers; 1—crankshaft; 2—connecting rod; 3—striker; 4—piston; 5—air cushion; 6—tool; 7—compensation opening.

Of substantial importance for hand-type electric impact machines are heating of the machine and recoil. The question of reducing recoil

of a comparatively light hand hammer (weighing 8–10 kg) while achieving an impact energy of \(A_b = 3.5\ \text{kgm}\) presents considerable difficulties. The overall efficiency of the hammer when striking the anvil is, at best, on the order of 40%. Consequently, 60% of the energy is converted into heat. Hence the following conclusion follows: in order to reduce heating of the hammer (more precisely, overheating, which worsens the operating regime of the machine), heat must be removed, i.e., cooling is required—air or water cooling. But the implementation of a cooling system and of means that reduce recoil will, in turn, increase the weight of the hand hammer.

The productivity of an impact machine is related to the impact energy and the number of blows, i.e., to the power of the machine. For a powerful column-type electropneumatic drilling hammer (equivalent in its performance to a pneumatic drilling hammer of the KTsM-4 type), the issues of recoil and heating do not lose their relevance, but they recede into the background, since elastic elements can be introduced into the system without special difficulty, as can water cooling with subsequent use of the water for flushing the borehole and reducing dust formation. In addition, in this case the problem of the weight of the structure is not as acute as in a hand tool.

At present, greater prospects exist for the creation of powerful, high-productivity column-type electropneumatic drilling hammers, especially machines for percussion-rotary drilling of large-diameter boreholes, as well as dynamic cutters [25] with an active edge (using impact). Here there are possibilities for producing a strong and reliable electropneumatic impact machine with a higher efficiency, which in operation can recoup the somewhat increased cost of its manufacture.

Such machines may be used for carrying out capital works, development workings, in drilling deep boreholes for block caving, and also in open-pit operations.

Proceeding from this, we believe that in the near future it is necessary to concentrate attention on the design of a powerful column-type electropneumatic hammer drill with a percussion-rotary tool, intended for drilling boreholes in hard rocks. The hammer drill should have adjustable impact energy within the limits \(A_b = 1 \div 10\ \text{kgm}\) and a blow frequency per minute \(z = 1400\); with a built-in electric motor connected with the impact mechanism by means of a gear transmission and a crank-and-connecting-rod mechanism; with a one-sided (or two-sided) air connection between the piston and the anvil; with transfer of energy from the anvil to the tool through an intermediate link by impact. The machine must provide regulation of the feed rate and must also have

automatic lubrication of the impact unit, water cooling of the motor, and a device for flushing the borehole.

The design calculation of an electropneumatic hammer, like that of any other machine, is a combination of certain design considerations based on experience with theoretical calculations. In theoretical investigations, in revealing the basic regularities and relationships, it is sometimes necessary to simplify and schematize the actual picture of the machine’s working process [1, 6]. Although in a number of cases the calculation does not yield actual quantities, its value for the designer is nevertheless quite great, since it provides the possibility of comparing the newly designed construction with other constructions that have proven themselves in practice; of carrying out machines operating according to an entirely new cycle; of the designer’s influencing the working process of the machine being designed; and of eliminating a number of shortcomings in the operation of the machine that have been noted in advance during the calculation. In heat engineering there is the term “ideal steam engine,” which historically has become attached to a machine operating according to a definite cycle without losses. “By introducing the concept of an ideal machine, it becomes possible, by comparing each working machine with the ideal one (at the same parameters), to evaluate objectively various machines from the standpoint of their economic merits and shortcomings” [34, p. 61].

What has been said above also applies to the design calculation of electropneumatic percussion machines.

Of exceptionally great importance for the designer is the choice of the parameters of a new machine.

As the initial quantities for selecting the parameters of a percussion machine of a definite purpose (for example, for drilling), the physicomechanical characteristics of the material being broken should serve. As shown in works [13, 30, 45], one of the principal characteristics is the magnitude of the specific work required to break one cubic centimeter of rock. In our investigations [7, 8, 9, 11] the influence of the feed force \(P\), the impact energy \(A_b\), and the angle of rotation of the drill between two impacts \(\alpha\) on drilling productivity was established. The value of the optimal angle \(\alpha\) of rotation of the drill between two impacts increases with increasing impact energy \(A_b\) and decreasing rock hardness (and conversely). Drilling with optimal feed forces, i.e., those at which the maximum drilling speed is ensured at the maximum efficiency of the machine, can be ensured in the case where, before the striker and drill collide, between the blade of the drill bit and the face surface there is a clearance \(\delta = 1 \div 4\) mm; the magnitude of this clearance depends on the impact energy and the hardness of the rock [7, 13].

These data must be known to the designer when implementing an elastic and adjustable feed system in percussion-rotary drilling.

Knowing the physicomechanical characteristics of the rock being broken and having the values of the initial parameters \(A_{\delta}, z, a, P\), and proceeding from the limiting impact velocity of the striker \(v_y\), we select the weight of the striker \(Q_{\delta}\). In single-acting machines with an air cushion, the unit weight of the striker is

\[ q=\frac{Q_{\delta}}{\pi D^{2}/4}=25 \div 75\ \text{g}/\text{cm}^{2}. \]

For gearless hammers and machines with a double-acting air cushion, the unit weight is \(q=50 \div 150\ \text{g}/\text{cm}^{2}\) [9, 19].

Knowing \(Q_{\delta}\) and \(q\), we determine the cylinder diameter

\[ D=\sqrt{\frac{4Q_{\delta}}{\pi q}}. \]

If there is a reduction gear, the speed of the crankshaft is equal to the number of blows per minute \(z\); consequently, the mean angular velocity is

\[ \omega_{\mathrm{cp}}=\frac{\pi z}{30}. \]

It remains to determine the values of the crank radius \(R\) and the initial length of the air cushion \(l_{0}\). The choice of the air-cushion length \(l_{0}\) is governed by the required weight quantity of air \(G\) and by the degrees of compression, and is determined, according to the laws of thermodynamics [27, 31], from the condition

\[ A_{\delta}=\frac{Q_{\delta}v_{y}^{2}}{2} =\frac{29.27}{m_{\mathrm{cp}}-1}(T-T_{0})G, \]

where the mean value of the polytropic exponent over the cycle may, in a first approximation, be taken as

\[ m_{\mathrm{cp}}=1.3;\qquad G=\frac{\pi D^{2}}{4}\,l_{0}\gamma 10^{-6}\ \text{kg}; \]

\(T\)—the absolute air temperature at the moment of compression;
\(T_{0}\)—the absolute initial temperature of the air cushion;
\(\gamma\)—the specific weight of air.

The choice of the air-cushion parameters depends on the machine power \(N\) and the number of revolutions per minute \(n\). From the indicator diagram of the air pressure over the cycle we have

\[ N=p'_{\mathrm{cp}}\frac{\pi D^{2}}{4}\,2l\,\frac{n}{60} \quad (\text{where } l_{\min}<l<l_{\max}), \]

whence the volume of the air buffer is

\[ V=\frac{\pi D^{2}}{4}\,l=\frac{30N}{p'_{\mathrm{cp}}\,n}, \]

i.e., for given values of the machine power \(N\) and \(p'_{\mathrm{av}}\), the volume of the cylinder (air link) is the smaller, the more revolutions the crankshaft makes; this regularity holds for all piston machines [3, 34].

Knowing the mean excess air pressure over the cycle,

\[ p'_{\mathrm{av}}=\frac{q v_y(1+e)}{g\,60/z}, \]

where

\[ e=\frac{v_{\mathrm{reb}}}{v_{\mathrm{imp}}} \]

is the rebound coefficient, we determine the work of the piston over the cycle

\[ A_{\mathrm{p}}=p'_{\mathrm{av}}\frac{\pi D^2}{4}\,4R =\frac{Q_b v_y z(1+e)}{15g}\,R. \]

According to the data obtained by us [2, 9], the efficiency of the air link is

\[ \eta_{\mathrm{a.l.}}=0.80\pm0.05. \]

Consequently,

\[ A_b=\eta_{\mathrm{a.l.}}A_{\mathrm{p}}\approx0.8A_{\mathrm{p}}. \]

We determine the crank radius \(R\) from the condition

\[ R=\frac{A_{\mathrm{p}}}{\pi D^2 p'_{\mathrm{av}}} =\frac{15g A_b}{\eta_{\mathrm{a.l.}}(1+e)zQ_b v_y}. \]

Then we find the power of the electric motor

\[ N_{\mathrm{mot}}= \frac{A_{\mathrm{p}}z/60}{\eta_{\mathrm{red}}\eta_{\mathrm{c.r.m.}}} = \frac{A_b z/60}{\eta_{\mathrm{a.l.}}\eta_{\mathrm{red}}\eta_{\mathrm{c.r.m.}}}, \]

where \(\eta_{\mathrm{red}}\) is the efficiency of the reducer (of the order of 0.9);
\(\eta_{\mathrm{c.r.m.}}\) is the efficiency of the crank-and-connecting-rod transmission (of the order of 0.95).

Taking into account the impact-load regime, it is advisable to take the electric-motor power with a reserve of 10–15%. In addition, it is important to know the characteristic of the electric motor, i.e., the dependence \(M_{\mathrm{cr}}=f(n)\).

For optimal operating regimes of an electropneumatic hammer with a one-sided air link, we found experimentally [9] the following relations:

\[ \frac{R}{l_0}=0.61\pm0.04;\qquad \frac{S_b}{2R}=1.22\pm0.05;\qquad \frac{v_y}{R\omega_{\mathrm{av}}}=1.63\pm0.07. \]

These relations can be used with sufficient reliability in selecting the parameters of the impact unit (\(A_b=1\div15\ \mathrm{kgm}\)).

In view of the fact that, when working in different rocks, the operating regime of the machine may change, the design of the impact unit must provide for the possibility of changing, within a certain-

within certain limits of the length of the air cushion \(l_0\) and to compensate for air leakage.

Advantages of a design and economic nature force the designer to increase, within known limits, the number of revolutions of the machine. Increasing the speed of the machine leads to an increase in its productivity and, owing to the greater energy saturation, to a reduction in the weight of the machine. As the speed of the machine increases, the kinetic energy of the rotating parts \(E\) increases. At the same time, for a constant value of the moment of inertia \(I\), the coefficient characterizing the nonuniformity of motion is significantly reduced,

\[ \delta=\frac{\omega_{\max}-\omega_{\min}}{\omega_{\mathrm{cp}}}, \]

as is the value of the characteristic criterion of the machine’s mode of motion

\[ K=2\frac{\varepsilon}{\omega^2}=\frac{1}{E}\cdot\frac{dE}{d\varphi}\ [16]. \]

But an increase in speeds has limitations, since in this case:

1. The magnitude of the inertia forces and the role of internal impacts due to clearances, backlash, etc., increase.

2. There arises the need for more careful balancing of the rotating parts, for a more rigid design of the machine as a whole and greater strength of its individual parts—the striker, drill steel, crankshaft, gear transmission, etc.—as well as for improving the quality of material processing and the accuracy of fitting all parts of the machine.

3. The limiting value of the angular velocity is connected with the parameters of the impact unit (with the complex of quantities \(\Pi=\frac{p_0 gD^2}{Q_{\mathrm{б}}\,l_0\,\omega^2}\)) and with a definite weight quantity of air \(G\) in the volume of the air cushion [3].

When the angular velocity is increased at a given power, it is necessary to reduce the volume of the air cushion, which may in turn cause an increase in pressure and temperature (greater heating, deterioration of lubrication, etc.), i.e., a worsening of the machine’s operating conditions.

1. The choice of speeds is connected with the character and essence of the technological process of the machine’s operation, with the specific features of its work and operation. The literature [15, 18, 24, 28, 30] indicates that, when studying the destruction of materials under the action of dynamic loads, it is necessary to know not only the energy required for destruction, but also the value of the velocity at which destruction occurs, since the effect may be different. Obviously, it is necessary to take into account the velocity of propagation of deformation and the time during which it acts.

1. The transfer of energy from the striker to the tool is limited by the value of the permissible impact velocity. Prof. N. A. Kilchevskii [24] believes that, at relative impact velocities of steel bodies on the order of 7 m/sec (with an elastic limit of 30 kg/cm²), an elastic impact takes place, not accompanied by residual deformations. Developing M. N. Gersivanov’s work [21] on the conditions of elastic impact of piles, V. M. Mostkov [28] shows that the impact between the striker and the drill can be considered completely elastic if the condition is satisfied

\[ E_b \frac{v}{a} < \sigma_{\mathrm{el}} . \]

“With the modulus of elasticity of steel \(E_b = 21 \cdot 10^6\) t/m², the speed of sound in steel \(a = 5.15 \cdot 10^3\) m/sec, and the elastic limit for carbon steel \(\sigma_{\mathrm{el}} = 3.5 \cdot 10^4\) t/m², the value of the piston (striker) velocity for satisfying this condition is \(v_{\mathrm{imp}} < 9\) m/sec” [28, p. 31]. Engineer N. M. Batuev [20] writes that the elastic limit of steel after hardening increases considerably, by 2–3 times; therefore elastic impact of the striker on the tool can be ensured at impact velocities \(v_y = 7 \div 10\) m/sec.

In conclusion, let us show the possibility of selecting the parameters of an impact-action machine on the basis of similarity theory.

In work [3] we established that dynamic similarity occurs in forging hammers. For electropneumatic machines with a one-sided air connection, the similarity criterion is the expression

\[ \Pi = \frac{p_0 g D^2}{Q \omega^2 l_0} = \mathrm{idem}. \]

Let us give an example. It is known from experience [2, 15] that an example of a well-operating machine is a hammer with the parameters: \(R = 39\) mm; \(l_0 = 58\) mm; \(Q = 1.65\) kg; \(z = n = 1400\) rpm; \(N = 4.95\) kW; \(D = 68\) mm and \(A_b = 7\) kgm. Suppose that we are interested in a new machine with number of blows \(z_1 = 1100\) blows/min, with striker weight \(Q = 3\) kg and work on the striker \(A_1 = 10\) kgm. Then, on the basis of the theory of dynamic similarity, we determine the scale coefficients.

The ratio of the times of the processes being performed

\[ t_c = \frac{t}{t_1} = \frac{z_1}{z} = \frac{1100}{1400} = 0.785 . \]

The ratio of masses

\[ m_c = \frac{m}{m_1} = \frac{Q}{Q_1} = \frac{1.65}{3} = 0.55 . \]

The ratio of works

\[ A_c = \frac{A}{A_1} = \frac{7}{10} = 0.7 = m_c v_c^2 . \]

Then the linear scale is

\[ l_c = v_c t_c = 1.13 \cdot 0.785 = 0.885. \]

Since

\[ l_c = \frac{l}{l_1} = \frac{R}{R_1} = \frac{D}{D_1}, \]

the new length of the air cushion will be

\[ l_1 = \frac{l}{l_c} = \frac{58}{0.885} = 65.5\ \text{mm}; \]

the new radius of the crank

\[ R_1 = \frac{R}{R_c} = \frac{39}{0.885} = 44.3\ \text{mm}; \]

the new diameter

\[ D_1 = \frac{D}{l_c} = \frac{68}{0.885} = 77\ \text{mm}. \]

The power ratio

\[ N_c = m_c \frac{l_c^2}{t_c^3} = 0.55 \cdot \frac{0.885^2}{0.785^2} = 0.89 = \frac{N}{N_1}, \]

whence the power of the new machine is

\[ N_1 = \frac{N}{N_c} = \frac{4.95}{0.89} = 5.55\ \text{kW}. \]

It is not difficult to show that, under identical operating conditions for these machines (rebound of the striker, air compensation, clearances and lubrication), the value of the similarity criterion will be the same:

\[ \Pi = \frac{1.980 \cdot 6.8^2}{1.65 \cdot 5.8 \cdot 146.5^2} = \frac{1.980 \cdot 7.7^2}{3 \cdot 6.55 \cdot 116^3} = 0.22, \]

If, for the machine considered, we change only the number of revolutions, assuming the parasitic resistances to remain the same, i.e., take a new time regime while preserving the linear dimensions (\(R = 39\ \text{mm};\ l_0 = 58\ \text{mm};\ Q = 1.65\ \text{kg}\)), then the linear scale is \(l_c = 1\), the mass ratio \(m_c = 1\), and the force ratio:

\[ F_c = m_c \frac{l_c}{t_c^2} = 1 \cdot \frac{1}{t_c^2} = n_c^2; \]

the work ratio

\[ A_c = F_c l_c = n_c^2; \]

ratio of powers

\[ N_c = m_c \frac{l_c^2}{t_c^3} = n_c^3 . \]

Let, for example, \(n_1 = z_1 = 700\) blows/min, then:

\[ n_c = \frac{n}{n_1} = \frac{1400}{700} = 2; \qquad F_c = 2^2 = 4 \]

(including for the air-pressure force):

\[ A_c = 2^2 = 4; \qquad A_1 = \frac{A}{4} = \frac{7}{4} = 1.75\ \text{kg·m}; \qquad N_c = 2^3 = 8 \]

and the power of the new machine

\[ N_1 = \frac{N}{2^3} = \frac{4.95}{8} = 0.64\ \text{kW}. \]

Suppose we wish to change only the cylinder diameter of the hammer while keeping the same ram weight \(Q_b = Q_1 = Q_2\); then from the condition of dynamic similarity it follows that

\[ \Pi = \frac{p_0 g D_1^2}{Q_b l_1 \omega_1^2} = \frac{p_0 g D_2^2}{Q_b l_2 \omega_2^2}, \]

whence

\[ \left(\frac{D_1}{D_2}\right)^2 = \frac{l_1}{l_2} \left(\frac{\omega_1}{\omega_2}\right)^2 . \]

Consequently, in order to preserve dynamic similarity at the same ram weight by changing the cylinder diameter, it is necessary either to change the number of revolutions of the crankshaft, or to change the cushion length (or to change these ratios simultaneously in accordance with the obtained dependence on the magnitude of the change in cylinder diameters).

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