DYNAMICS AND WORKING PROCESS OF A GEARLESS ELECTRIC JACKHAMMER WITH A RIGID COUPLING DURING THE RETURN STROKE

P. M. ALABUZHEV

The problem of the electric jackhammer is ultimately a problem of the weight of the structure as a whole, a problem of the strength of parts at reduced weight. Existing special-purpose non-jackhammers, for example, KM-1 and MPT-5, are not limited in weight and can readily be designed in accordance with the customer’s requirements. The shortcomings of hammers of the ZERT type, when used as jackhammers, are the low impact energy at the striker and the variability of the cycle with operating conditions. In KhNSh hammers the impact energy at the striker is insufficient, and breakages of the steel springs occur frequently.

To reduce the weight of the structure it is necessary to use high-speed electric motors. For a given motor power, the productivity of the hammer depends to a greater extent on the impact energy than on the frequency of impacts. The use of the ordinary method of reduction, for example by means of a gear train, to reduce the impact frequency relative to the number of revolutions of the electric-motor rotor leads to complication of the mechanism and to an increase in its dimensions, that is, to an increase in the weight of the hammer. Engineer N. M. Batuev of the Tomsk Electromechanical Plant (TEMZ) proposed the idea of reduction without the use of ordinary gear (or friction) transmissions—the so-called principle of skipping impacts; the structural implementation of the proposed idea was embodied in the gearless hammers MEI, MB-1, MB-2, and MB-3. In the first two machines there was a rigid coupling of piston 1 with striker 2 during the working and return strokes, which in the latter case was effected by means of a special coupling part called a “catch”; the latter is a massive body 3 eccentrically mounted on an axle in the piston. The operation of the catch is based on the action of inertial forces caused by the motion of the piston of the crank-and-connecting-rod mechanism. The prescribed skipping of impacts was ensured by corresponding kinematic adjustment of the coupling. A substantial shortcoming of these designs was the presence of internal impacts; and since springs did not justify themselves in the predecessor designs, air 4 is used as the elastic material to eliminate internal impacts between the piston and the striker. In the MB-2 and MB-3 designs a new, more perfect working process was implemented; moreover, the possibility of using dynamic skipping of impacts was demonstrated.

For the fastest possible calming of the striker, and also in order to pick up the striker from its extreme left-hand position, one may either precompress the air between the striker and the piston or insert between the striker and the housing of the mechanism a comparatively weak calming spring 5.

After the working impact and the calming of the striker, the return stroke of the striker is carried out by means of the inertial catch 3, Fig. 1. Disengagement

the striker with the piston occurs after the appearance of relative motion of the striker with respect to the piston; in this case the rigid connection between them is replaced by an elastic connection (in the form of an air cushion of variable length \(l\)), which exists until the very moment of the working blow.

In light and high-speed electric motors, under the influence of the load during one working stroke of the striker, with the proper productivity determined by the operation of the percussion unit, a considerable drop in angular velocity occurs. In the case of an inelastic blow and in the absence of a skipping mechanism, the entire cycle is repeated, but with an even lower mean angular velocity; after several revolutions the working process would be disrupted because of an excessively large drop in the motor’s angular velocity. If, however, one proceeds only from the normal operation of a light and high-speed electric motor, then we obtain a large number of low-productivity blows on the striker. In reality there is rebound of the striker, and it is necessary to keep in mind the possibility of internal blows or “increased” air pressure between the striker and the piston (with an unsuccessful choice of parameters, sometimes considerably exceeding the permissible values). Consequently, for the stability of the operation of this mechanism, time is required for the striker to come to rest and for the motor and the crank-and-connecting-rod mechanism to replenish the expended energy. The reduction in the number of working strokes (that is, blows of the striker) in comparison with the number of revolutions of the motor is accomplished by skipping the engagement of the piston with the striker. Let us briefly describe the principles of operation of the skipping mechanisms.

Fig. 1

1. In dynamic skipping we proceed mainly from the operation of the motor: the catch must not pick up the striker until a definite value of the angular velocity \(\omega\) has been established in the electric motor; for this purpose the catch, having moved back as far as the stop (to the right), must not itself rotate freely in the opposite direction, which is achieved structurally by cutting in it a small groove, opposite which in the piston there is a hole of the same diameter; into this hole a compressed spring is inserted, one end bearing against the piston and the other against a small weight freely entering the groove of the catch. The weight of the small load and the stiffness of the spring correspond to the angular velocity of the motor at which engagement occurs, denoted by \(\omega_3\), and if 1) \(\omega > \omega_3\), then, under the influence of inertial forces, the small load compresses the spring and “sinks” into the recess, as a result of which the catch becomes free, that is, it can rotate about its axis; 2) when \(\omega < \omega_3\), the developed inertial forces are insufficient (for the selected values of the stiffness of the spring and of the small load, the load will not leave the recess of the catch and will deprive it of the possibility of turning and producing engagement). Thus, skipping is determined by the number of idle strokes of the piston, or by the number of revolutions which the motor crank makes, tending toward steady-state motion.

Theoretical and experimental investigations have shown that the motor equalizes its angular velocity considerably faster in comparison with the time required for the striker to come to rest (especially in blows against steel). Consequently, with dynamic skipping of the blow, the possibility is not excluded of an undesirable rigid connection of the piston with a striker moving in the opposite direction.

1. In order to eliminate the possibility of head-on blows when the rigid connection is suddenly introduced between bodies moving in opposite directions—

...between the striker and the piston, it is necessary to catch the striker after it has come to rest, that is, to introduce a skipping mechanism corresponding to the settling time of the striker. The forced mechanism for skipping blows (kinematic scheme) is structurally implemented in the form of a ratchet wheel (Fig. 2), mounted freely on the common shaft with the gripping device. The teeth of the ratchet wheel are of unequal depth: after one shallower notch there follow several deeper ones, the number of which corresponds to the selected skipping of blows. Inside the connecting rod and the piston there are recesses into which pawls (weights) $M_1$ and $M_2$, with springs, are inserted; the springs press the pawls against the ratchet wheel. These pawls $M_1$ and $M_2$ make it possible for the ratchet wheel to rotate in one direction. When the crank radius makes one revolution clockwise, the ratchet wheel turns through an angle corresponding to one recess (tooth) in the direction opposite to the motion of the clock. The lateral projections of pawl $M_2$ enter from both sides into the recesses of the gripping device. While the inner part of pawl

Fig. 2

$M_2$ slides along the deep notches in the ratchet wheel, the stroke of pawl $M_2$ is small, as a result of which its lateral projections do not leave the recesses of the gripping device; but as soon as the inner part of pawl $M_2$ falls onto a shallow notch in the ratchet wheel, its lateral parts will come out of the recesses of the gripping device and make it free. Engagement and the working stroke will occur; then a skip in engagement will follow, and so on; a quite definite regime is established.

1. Dynamic skipping of a blow can be carried out (proceeding from the operation of the impact unit) by selecting an air cushion of considerable length (a long and massive striker); in that case the kinetic energy developed by the striker is distributed over several piston strokes. The skipping of a blow will be determined by the number of idle revolutions of the electric motor. Undesirable engagements can be eliminated by the presence of a ratchet wheel with the corresponding regime. With long strikers it is possible to implement skipping of blows even without a special device, by selecting the dimensions and shape of the corresponding part of the impact unit, proceeding from the equation of motion of the gripping device, the initial conditions of motion, and also the presence of frictional engagement between the piston and the striker; in the latter case, in the presence of frictional engagement, it is easy to implement a more expedient two-sided form of the gripping device, which reduces stress in the striker and skewing of the piston.

Theory and calculation

It is necessary to note the difficulties in the theoretical study of the working process of an electro-pneumatic hammer: 1) we have a mechanism with two degrees of freedom, in which the elastic coupling—the air between the striker and the piston—does not obey Hooke’s law; 2) for any characteristic of the electric motor, owing to the specific operating conditions of the mechanism, it is necessary to take into account the nonuniformity of the angular velocity of rotation of the crank; 3) it is necessary to keep in mind the phenomenon of rebound of the striker after impact and vib-

of the hammer casing (recoil); 4) at present there are still insufficient values for a number of coefficients and indices, which must be found experimentally.

The first more complete theoretical study of the dynamics and working process of an MB-2 type electropneumatic hammer was carried out by a group of researchers of the Tomsk Polytechnic Institute (TPI), consisting of Prof. Kh. F. Ketov, Docent P. M. Alabuzhev, and assistant N. G. Shlenkina, under the general supervision of Prof. Kh. F. Ketov¹ [1, 2]. Since we had no analytic solution of the problem in its general form, we used an approximate method of numerical integration of the equations of motion of the striker and the electric motor. Further development of the calculation method, with its application to other types of hammer designs, is described in our paper [3].

The method of numerical calculation was used in choosing the values of the principal, more rational parameters determining the prescribed working process of the hammer. The work was carried out in 1945–1946 at the design bureau of TEMZ by order of Glavuglemash Vostok; assistant of the Department of Theoretical Mechanics of TPI V. K. Solov’ev took part in carrying out the calculation work.

A preliminary theoretical study of the working process, for assumed parameter values, reduces to the necessity of solving two simultaneous equations: the equation of motion of the striker

\[ \frac{Q_b}{g}\cdot a = -\frac{\pi D^2}{4}(p-p_0)-c\cdot(x+x_0) \tag{1} \]

and the equation of motion of the motor

\[ \frac{d}{d\varphi}\left(J\cdot\frac{\omega^2}{2}\right) = M_k+M_p+M_c . \tag{2} \]

Here the notation is as follows:

\(Q_b\)—weight of the striker;

\(g\)—acceleration due to gravity \((g=9.8\ \text{m/sec}^2)\);

\(a=\ddot{x}\)—acceleration of the striker;

\(D\)—piston diameter;

\[ p=p_0\cdot\left(\frac{l_0}{l}\right)^{\alpha} \]

—air pressure \((p_0\simeq 1\ \text{kg}/\text{cm}^2)\);

\(\alpha\)—mean polytropic exponent over the cycle (taken by us, on the basis of previous work at the TEMZ design bureau, as equal to 1.32);

\(c\)—stiffness of the calming spring;

\(s_b=x\)—travel of the striker from the beginning of its motion;

\(x_0\)—initial compression of the spring;

\(l=s-x+l_0'\)—length of the air cushion (the location is shown in Fig. 3);

\(l_0'\)—initial distance between the piston and the striker (at the moment of engagement);

\[ s=R\cdot\left[\left(1+\frac{\lambda}{4}\right)-\left(\cos\varphi+\frac{\lambda}{4}\cos2\varphi\right)\right] \]

—the distance traveled by the piston from its initial position (the engagement position);

$R$ — crank radius;

\[ \lambda = \frac{R}{L}, \]

where $L$ is the length of the connecting rod;

$\varphi$ — angle of rotation of the crankshaft;

$\omega = \dot{\varphi}$ — angular velocity of the crankshaft;

$\varepsilon = \ddot{\varphi}$ — angular acceleration of the crankshaft;

Fig. 3

\[ J = J_0 + \frac{Q_{\sigma}+Q_{n}+\frac{1}{3}Q_{\sh}}{g}\cdot R^2 \cdot \left(\sin\varphi+\frac{\lambda}{2}\sin 2\varphi\right)^2 \]

— moment of inertia of the moving masses, reduced to the crankshaft axis;

$J_0$ — moments of inertia of the rotor, crank, and $2/3$ of the length of the connecting rod;

$Q_n$ — weight of the piston;

$Q_{\sh}$ — weight of the connecting rod;

$M_k$ — driving torque determined by the motor characteristic (see Fig. 4);

Fig. 4

\[ M_p = \frac{\pi D_2}{4}\cdot R \cdot (p-p_0)\cdot \left(\sin\varphi+\frac{\lambda}{2}\sin 2\varphi\right) \]

— moment of the air pressure relative to the crankshaft;

$M_c = cR(x_0 + x)\left(\sin\varphi + \frac{\lambda}{2}\sin 2\varphi\right)$ is the moment of the spring-pressure forces relative to the crankshaft.

Equations (1) and (2) are differential, second-order, nonlinear equations with variable coefficients; the solution was carried out approximately, with a degree of accuracy sufficient for practice.

The Role of the Basic Parameters Determining the Working Process

It is, of course, impossible to obtain the desired working process for arbitrary parameter values. Only with a successful combination and selection of them can one “squeeze” everything possible out of the design, in accordance with the requirements imposed on it.

One of the basic parameters is the crank radius $R$. Large values of $R$ provide high velocities of the piston and striker, but, on the other hand, increasing the radius has an adverse effect on the motor, increases vibrations of the casing during the idle stroke, and, finally, makes the dimensions of the hammer head large and consequently increases its weight.

The weight of the striker $Q_b$ and its shape play a substantial role. The weight of the striker affects the total weight of the mechanism, the impact energy, and the recoil. The magnitude of the air-pressure force, as well as the total weight of the barrel, depends on the diameter $D$. The length of the air cushion $l_0$ determines the degree of compression and also, to a considerable extent, the location of the point of impact in terms of the angle $\alpha$.

The value $l_0$, in the presence of a frictional engagement, may be taken as a parameter determining the skipping of an impact (reduction) without a special skipping mechanism. An unfortunate choice of the length $l_0$ creates excessive air pressures at maximum values of

\[ \left(\sin\varphi + \frac{\lambda}{2}\sin 2\varphi\right), \]

since the value of $M_p$ is large, which leads to unproductive energy losses in the electric motor and to a sharp drop in angular velocity.

The spring fixing the striker, holding it after the working impact in its extreme position, is necessary to calm the striker so as to ensure the next engagement at the beginning of a new cycle. In addition, it to some extent “improves” the working process by eliminating pressure peaks (“cutting” them off). An excessively strong spring will increase the recoil of the hammer.

In addition to everything listed above, the motor and its characteristic, as well as the physical properties of the rock on which the impact is made, play a substantial role in the entire working process of the hammer.

The position of the mechanism has an insignificant influence on the working process, since at $n = 3000$ rpm and $R = 2$ cm the inertia forces are 180 times greater than the weight forces.

For a more rational use of the electric motor, a reduction in the recoil of the hammer casing, a reduction in heating, and also an increase in the reliability of individual parts and assemblies, it is important that there be no elevated air pressures $P$ either in the operating regime or after the main impact.

In the calculations we proceed from the condition that the impact energy at the striker $A_b$ should be of the order of $2.5$ kg·m, without elevated air pressures $p$ and without significant oscillations of the angular velocity of the motor $\omega$ over the entire working cycle. With unchanged values of the motor characteristic (Fig. 4), the moment of inertia $J_0 = 0.0080$ kg·cm·sec$^2$, piston weight $Q_p = 0.25$ kg, and connecting-rod weight $Q_{\mathrm{sh}} = 0.11$ kg, the parameters $R$, $l_0$, $l'_0$, $D$, $Q_b$, $c$, $x_0$, $\lambda$ determining the working process of the mechanism were selected.

Summary table of calculated variants MB-3

The final data for the fifteen calculated variants are combined in the appended table (p. 12). Brief remarks on it:

1) in variants 1, 2, 3, 4, 5, the air pressures between the striker and the piston are too high (of the order of \(p \simeq 25\ \text{kg}/\text{cm}^2\)) after rebound of the striker, and there is a sharp drop in the angular velocity of the motor \((\omega_{\min} \simeq 30\ 1/\text{s})\);

2) in variants 6, 7, 8, the striker is too thin-walled;

3) in variants 8, 9, 10, the air pressures in the operating regime are high, as a result of which, before impact, a large drop in the angular velocity of the electric motor occurs \((p_{\max}\) coincides with \(M_p{}_{\max})\);

4) in variants 12 and 15, the electric motor is used insufficiently and the impact energy on the striker is small;

5) in variant 14, the damping spring is absent; the air pressure in the operating regime is high;

a. Impact mechanism with a short operating cycle.

b. Impact mechanism with a long operating cycle.

Fig. 5. Dependences of the path of the striker \((S_b)\), the path of the piston \((S_p)\), and the angular velocity of the electric motor \((\omega)\) on the angle of rotation of the crankshaft \((\varphi)\)

6) variants 11 and 13 are accepted as “optimal” for design, since they basically satisfy the conditions listed above. The working-process graphs for variants 11 and 13, according to the numerical calculations performed, are shown in Fig. 5 (a and b). Here the accepted rebound energy of the striker is equal to 10% of the impact energy. These graphs give the designer a complete idea of the nature of the operating process of the mechanism.

A preliminary theoretical study of the operation of the design makes it possible to eliminate the observed shortcomings in time, before manufacture of the experimental—

of the new model; moreover, it opens up possibilities for designing mechanisms operating according to different cycles. Engineers N. M. Batuev and A. P. Grishin of TEMZ took part in the design implementation of variants 11 and 13.

To ensure the specified operating cycle of the hammer, reliable and precise operation of the inertial engagement and disengagement mechanism is required. From the dynamic point of view, in the problem under consideration it is necessary, for the given initial conditions, to find the law of motion of a physical pendulum whose suspension axis performs a rectilinear oscillatory motion the same as that of the piston of the crank-and-connecting-rod mechanism.

On the basis of d’Alembert’s principle, considering the “grip” in the equilibrium position, we write the moments of all forces with respect to the axis of rotation of the “grip” (see Fig. 6). After reduction of the terms of the equation by \(M_3\) (the mass of the grip), the equation of motion of the “grip” will take the form:

\[ \frac{d^2 Q}{dt^2}=\ddot{\Theta} = R \cdot \omega^2 \cdot \frac{r_c}{\rho^2}\cdot(\cos\varphi+\nu\cos2\varphi)\cdot\cos\Theta - g\cdot\frac{r_c}{\rho^2}\cdot\sin\Theta . \tag{3} \]

Here:

\(r_c\) — distance from the center of gravity of the grip to the axis of rotation;
\(\rho\) — radius of inertia of the grip;
\(\Theta\) — angle formed between \(r_c\) and the \(y\)-axis (vertical) (see Fig. 6),
\(\varepsilon=0\) — (the case \(\varepsilon=\varepsilon_i\ne0\) was also considered by us).

Fig. 6

\(M_3\) — mass of the grip
\(C\) — center of gravity

The differential equation (3) is also a nonlinear second-order equation with variable coefficients and was solved approximately, using the calculation methods described in [3]; moreover, in the calculation there is no special need to choose the origin of the fixed coordinate axes in such a position of the mechanism for which the acceleration of the slider would be equal to zero, as is noted in the book [2, p. 278].

For normal operation of the impact mechanism, the striker must be caught when the velocities of the striker and piston are equal, with no premature rebound (recoil) of the grip at the moment of its contact with the striker. In order to eliminate internal impacts, we considered and calculated various design schemes of grips (in one of the variants an original grip of variable length was considered, ensuring reliable engagement and omission of impacts); questions were considered concerning the choice of the most expedient point of impact in the grip, taking into account reduction of rebound and avoidance of transmission of the impact to the axis of rotation of the grip. At the point of impact there occurs mutual deformation of the elastic bodies (the grip and the wall of the striker) and their indentation into one another; but in the presence of forces

friction, at a certain value of the angle at which engagement occurs, there will be no rebound of the grip from the striker wall.

Along with calculations of the inertial grips for engagement, we carried out verification calculations for disengagement; here the question arose of the possibility of the grip jamming under frictional engagement, which, in turn, could lead to a delay in the onset of disengagement of the striker from the piston and to a disturbance of the normal working process of the hammer. To eliminate jamming, the contact surface of the grip and the striker must have sufficient hardness; and the parts associated with the formation of the angle at which frictional engagement occurs must be made within specified tolerances corresponding to the required accuracy class. As a result of theoretical and experimental studies, the most rational form of the grip was obtained (see Fig. 2), ensuring its reliable operation. It may be noted: 1) the substantial role, in the speed of action of the grip, of the “shape coefficient,” determined by the ratio \(r_{\rho}/S^2\), and of the radius of curvature \(R\); 2) the independence of the increment of the angle \(\Theta\) from changes in the angular velocity of rotation of the crankshaft \(\omega\), until it becomes so small that the second term, due to the weight of the grip, already plays an essential role in equation (3).

From equations (1), (2), (3), the following similarity criteria for gearless electric hammers can be derived [4]:

\[ \Pi_1=\frac{q}{p_0}\cdot\frac{l_0\omega^2}{g}=idem \]

a criterion characterizing the relation between the forces of inertia and the forces of gravity (here \(q=\dfrac{4Q_{\mathrm{b}}}{\pi D^2}\) is the unit weight of the striker).

\[ \Pi_2=\frac{Q_{\mathrm{b}}\omega^2}{g\cdot c}=idem \]

(a criterion characterizing the relation between the forces of inertia and the forces of elasticity);

\[ \Pi_3=\frac{c}{l_0p_0}=idem \]

(the relation between the forces of elasticity and the forces of air pressure);

\[ \Pi_4=\frac{Rr_c\omega^2}{\rho^2\Theta}=idem \]

(the relation between the forces of inertia developed in the crank-and-connecting-rod mechanism and in the grip);

\[ \Pi_5=\frac{g}{\omega^2R}=idem \]

(the relation between the forces of gravity and the inertial forces in the grip).

For dynamically similar machines, the values of the similarity criteria must be one and the same (idem); moreover, when selecting the optimum parameters of the machine it is rational to find, in one form or another, a relationship between the complexes of dimensionless quantities entering into the value of the similarity criteria.

In determining the motion of the striker and of the electric motor, the calculations did not take into account friction forces, heat losses, losses due to vibration of the housing, and it was also assumed that there was no air leakage between the piston and the striker; in addition, the change of air in the working volume was assumed to occur according to a polytropic law with a constant value of the exponent, and therefore the actual working process must differ from the calculated one.

For the purpose of “refining” the design, the Plant made several strikers of different shape and weight, light pistons for changing the length of the air cushion; damping springs of various stiffnesses; a crankshaft with a variable radius was ordered; the motor power was regulated by a rheostat.

The experimental model of the OMB-3 hammer, weighing $\simeq 10\ \text{kg}$, built according to the design variant calculated from (11), during short tests in the factory laboratory showed good and clear operation with an impact energy on the striker of the order of $A_b \simeq 2\ \text{kg}\cdot\text{m}$. The experiment proved that the inertial “grip” provides reliable engagement, disengagement, and the prescribed omission of a blow.

Further tests of experimental models of hammers of this type were carried out in Moscow at VUTI and VNIISTROYDORMASH and are described in works [5; 6; 7; 8].

At the present stage of development of gearless systems, the following advantages may be noted:

1) a considerable reduction in the weight of the design, a reduction in its dimensions, and a reduction in the number of parts;

2) the possibility of omitting blows within wide limits without changing the dimensions and weight of the hammer;

3) the possibility of designing impact mechanisms with strikers of both small and large masses while the overall weight of the machine remains almost unchanged; the possibility of changing the masses of the strikers in order to increase the impact energy by reducing the number of blows is of especially great importance for electric rock drills;

4) definiteness of the operating cycle of the impact mechanism under any production conditions.

The drawbacks must also be noted:

1) the presence of vibrations of the housing during idling, which requires that hand hammers necessarily be provided with rubber handles or that special measures be taken for sufficiently complete balancing of the inertia forces;

Fig. 7

2) a hammer shape, somewhat inconvenient under operating conditions, with an asymmetric transverse arrangement of the electric motor, see Fig. 7, which we borrowed from [8];

3) an insufficiently advantageous shape of the striker—mainly with regard to strength—in the form of a cup, especially in variants with small mass and considerable impact energy;

4) it should be noted that there is significant heating of the electro-pneumatic hammer, chiefly from the air cushion, which, in turn, poses for the designer the problem of removing heat from the machine housing.

The shortcomings noted are not organically inherent in gearless systems and can be substantially weakened or eliminated. Moreover, there are various possibilities here; for example: 1) all attention can be concentrated on the problem of the striker’s strength (in the event of frequent failures and on the technology of its manufacture); 2) but another solution is also possible—the presence of heating partly returns design thinking to the “compromised” spring, and at present there are more grounds for using it as an elastic coupling than there were 10 years ago, since there is experience in calculating and applying springs in aircraft-engine construction; in addition, the technology for manufacturing high-quality springs has been considerably improved. With the introduction of a spring coupling, the question of the technology of manufacturing the striker is greatly simplified, and the problem of its strength is almost removed.

The hammer can be given a more convenient form by replacing the crank-and-connecting-rod mechanism with a mechanism having an oscillating washer, and thus it is possible to make the jackhammer still lighter and more portable.

Questions of strength and reliability under production conditions will require much more work on perfecting and mastering the design before putting it into serial production; nevertheless, it must be noted that this design is promising as a jackhammer and as a light electric drill.^1

LITERATURE

1. Prof. Ketov Kh. F. Investigation of the dynamics and working process of a gearless electropneumatic hammer. In the collection Mashinostroenie. Proceedings of LPI, No. 4, Mashgiz, 1953.

2. Bezhanov B. N. and Bushunov V. T. Production automatic machines. Mashgiz, M.–L., 1953.

3. Alabuzhev P. M. Investigation of the working process of an electric jackhammer with an elastic (air) coupling. Izv. TPI, vol. 61, issue 1, 1947.

4. Alabuzhev P. M. Application of the theory of similarity and dimensions to the study (modeling) of impact-action machines. Izv. TPI, vol. 73, 1952.

5. Batuev N. M. Electric impact drills. Journal Mekhanizatsiya trudoemkikh i tyazhelykh rabot, No. 4, 1947.

6. Batuev N. M. Electric hammers. VNIISTROYDORMASH. Investigation of machines and mechanisms for construction and road works. I, Mashgiz, 1950.

7. Batuev N. M. An electric hammer of a new type. VNIISTROYDORMASH, collection on the exchange of experience of plants of construction and road machine building. Mashgiz, 1950.

8. Batuev N. M. Energy characteristics of electric hammers and ways of increasing their productivity. VNIISTROYDORMASH. Investigation of vibrators and electric hammers VI. Mashgiz, 1953.

^1 Work on the investigation and calculation of various designs of the inertial switch was carried out by us beginning in 1941, at the suggestion of the design bureau of TEMZ. This work was reported at a scientific meeting of the departments of the Mechanical Faculty of TPI in 1946 and at the mining section of the VI Conference of Scientists of the city of Tomsk on 20/III–1947.

Noted Misprints

1. In book [2, p. 271] there is an inaccurate reference to a report on work carried out at LPI, whereas in fact this work, too, was carried out at TPI under a contract with TEMZ in 1942. ↩