STUDY OF THE INFLUENCE OF CERTAIN TECHNOLOGICAL FACTORS AND LOAD CONDITIONS ON THE WEAR OF GEAR WHEELS IN GEARBOXES OF MINING ELECTRIC DRILLS

V. T. Gorbenko, Yu. S. Semenov, S. I. Shubovich

(Presented by the scientific seminar of the Department of Applied Mechanics)

In calculations of enclosed gear transmissions, their load-carrying capacity is, as a rule, limited by contact strength, while the bending strength of the teeth in this case proves to be greatly overestimated. For example, calculations of the gear transmissions of gearboxes of hand-held mining electric drills carried out by us according to the methods of [1] and [2] show that, even with a two- to threefold overload in contact strength, the transmissions satisfy the requirements of calculations for bending and seizure. Meanwhile, since the gearboxes of mining electric drills, by the conditions of operation, have small dimensions, such a load is usual for them. Naturally, operation with a considerable overload, in addition to spalling of the working surfaces of the teeth, leads to their wear. The investigations carried out, for example, in [3], [4], showed that even progressive spalling does not cause any noticeable increase in wear and does not substantially reduce the service properties of the transmission.

The tooth profiles change with wear, become involute [4] and, consequently, non-interchangeable. Therefore one should strive to ensure that gear wheels, when worn, are of equal value. As a first approximation to this it is desirable that the magnitudes of wear of the teeth of the pinion and the wheel be close to one another and, as far as possible, small. Such wear and the factors promoting it (for example, an appropriate ratio of hardnesses, cleanliness of the tooth surfaces, etc.) will be called optimal.

The gear wheels of gearboxes of mining electric drills are made with high hardness of the tooth surfaces, and after heat treatment no finishing operations are applied. In view of their small dimensions, the gear wheels of gearboxes of mining electric drills operate at low peripheral speeds (up to 3 m/sec); therefore such a method of manufacturing them is quite expedient [5]. During heat treatment of cemented pinions, a peculiar surface is sometimes formed (Fig. 1, a), which we shall call “warty.” It is not excluded that such pinions go into assembly. The shape of the “warty” protrusions is almost hemispherical; their height and diameter are 0.2–0.3 mm. To study the microstructure of the metal of the “warts,” polished sections were prepared on oblique cuts of the pinion teeth. It was established that the metal of the “warts” is not etched by a 6% solution of nitric acid in alcohol. Such a feature of the metal of the “warts,” together with a high microhardness (900–1000 \(H_v\)), is possible only with an austenitic-martensitic structure (Fig. 1, b). The latter is consistent with the high wear resistance of the metal of the pinion teeth. Light areas that etch poorly and have similar properties—

... and structure may form during grinding at elevated regimes of carbon and alloy steels [6], and also with significant wear of working surfaces [7]. Apparently, the mechanism of formation of such a structure during carburizing is quite different than during grinding or wear. Additional studies are being carried out to clarify questions concerning the origin and process of formation of “beards.”

When gear wheels with a relatively small number of teeth (pinions) are cut, the tooth profiles very often do not constitute a smooth curve, and, as a result of faceting, a break in the profile occurs (Fig. 2). Such a break (edge) may be on one side or on both sides of the tooth (profile); usually it is located on the tooth flank, sometimes in the dedendum zone, and very rarely on the tooth tip.

To assess the degree of influence on gear-wheel wear of the factors indicated, as well as of the magnitude and character of the load, the hardness and the quality class of the working surfaces of the gear pair, tests were carried out at the Department of Applied Mechanics of the Tomsk Polytechnic Institute on the gear wheels of two-stage gearboxes of mining electric drills ER-5 (tests conducted by Yu. S. Semenov) and ER-16 (tests conducted by V. T. Gorbenko). The low-speed stages of these gearboxes are spur-toothed, \(z_1 = 17\), \(z_2 = 28\), \(m = 2\ \text{mm}\); the high-speed stage of the ER-5 gearbox is spur-toothed, \(z_1 = 14\), \(z_2 = 46\), \(m = 1.5\ \text{mm}\); of the ER-16 gearbox it is helical-toothed, \(z_1 = 17\), \(z_2 = 58\), \(m_n = 1.25\ \text{mm}\), \(\beta_n = 18^\circ\). All spur gear wheels were uncorrected, and the helical ones corrected (\(\xi_1 = +0.3\), \(\xi_2 = -0.55\)). The material of the wheels was steel 40Kh, hardened; the pinions were carburized: the spur ones made of steel 20Kh, the helical ones of steel 18KhGT.

Fig. 1. View of a “bearded” surface (a) and of the metal microstructure in the “beard” zone (b).

The gearboxes were tested on a closed test rig; the rotational speed of the wheel of the low-speed stage in all tests remained constant—508 rpm. The ER-5 gearboxes were tested under different loads: a variable load, changing according to a linear law from \(N_{\min} = 0.7\ \text{kW}\) to \(N_{\max} = 1.8\ \text{kW}\) with a frequency of 15 cycles per minute [3],

and with loads constant in magnitude: minimum—$N_{\min}=0.7$ kW, nominal—$N_{\mathrm{n}}=1.2$ kW, variable equivalent in contact strength—$N_{\mathrm{e}}=1.3$ kW, and maximum—$N_{\max}=1.8$ kW. The gearboxes of the ER-16 electric drill were tested only under the constant maximum load, $N_{\max}=1.8$ kW.

The hardness of the gear wheels during the tests was varied over wide limits. The gear pairs of the high-speed stages of both types of gearboxes for the most part had the following hardnesses: for pinions, $\mathrm{HRC}_1=55$–60; for wheels, $\mathrm{HRC}_2=40$–45. This hardness ratio proved to be optimal for the high-speed stage of the ER-5 gearbox. In the latter, at all loads except the equivalent load, gear pairs of equal hardness were also tested, $\mathrm{HRC}_1=\mathrm{HRC}_2=40$–50; i.e., equality of hardnesses here was achieved by reducing the hardness of the pinion to the hardness of the wheel. At the equivalent load a pair with pinion and wheel hardness $\mathrm{HRC}_1=\mathrm{HRC}_2=48$ was tested; at the variable and maximum loads—a pair with wheel hardness increased to $\mathrm{HRC}_2=48$, with the same pinion hardness as in the main experiments ($\mathrm{HRC}_1=55$–60). In the high-speed stage of the ER-16 gearbox, experiments were also carried out in which the hardness of one or both elements of the gear pair was changed to $\mathrm{HRC}_1=52$ and $\mathrm{HRC}_2=46$.

Fig. 2. Fracture edge of the profile on a pinion tooth.

For the low-speed stages of the ER-5 gearboxes, a large difference in hardness was mainly used: for the pinion $\mathrm{HRC}_1=55$–60, for the wheels $\mathrm{HRC}_2=30$–35. In addition, at maximum and variable loads, gear pairs with wheel hardness $\mathrm{HRC}_2=38$–40 were tested, while the hardness of the pinions either remained $\mathrm{HRC}_1=55$–60 or was reduced to $\mathrm{HRC}_1=49$. For the low-speed stages of the ER-16 gearboxes, the hardness of the pinions in all experiments was $\mathrm{HRC}_1=55$–60; the hardness of the wheels in different experiments varied: $\mathrm{HRC}_2=37$–38, $\mathrm{HRC}_2=44$–45, $\mathrm{HRC}_2=48$. Thus, in addition to experiments with a large difference in hardness between meshing gear wheels in the low-speed stages of both gearboxes, in order to determine the optimal combination of pinion and wheel hardnesses at a different gear ratio, gear pairs were tested with the same hardness as that optimal for the high-speed stage of the ER-5 gearbox, and also with closer hardness values, obtained in one case by reducing the pinion hardness and in the other by increasing the wheel hardness.

Most of the gear wheels in our experiments had a surface finish of the working tooth surfaces of $\nabla 6$; at the same time, some gear pairs of the low-speed stage of the ER-16 gearbox had a surface finish of $\nabla 7$ [8].

In the tests, some variation was allowed in the presence and location of the break edge of the tooth profile of the pinions, and, in the gears of the low-speed stages, in the presence of a “warty” surface. Since, when cutting gears, in the overwhelming majority of cases there is a break edge of the profile on the tooth flank, and since a “warty” surface is formed comparatively rarely during heat treatment, gear pairs with a tooth working-surface cleanliness of $\nabla 6$, having on the pinion a profile break edge on the tooth flank, were mainly tested; therefore, below we shall specifically note only deviations from these conditions. This does not apply to the helical pairs of the high-speed stage of the ER-16 gearboxes, whose pinions did not have a clearly expressed break edge on the tooth profiles; apparently, the smoothness of their profile is explained by the fact that they are helical and therefore have a large difference between $z_1$ and $z_{\min}$.

The tests were conducted both with grease (solidol M) and with liquid (avtol AK-6, nigrol) lubricants. However, since no substantial difference in tooth wear was observed in this case [9], the type of lubricant for one or another gear pair is not specified below.

The test results are presented in the form of graphs in Figs. 3, 4, 5, and 6. From Figs. 3 and 4 it is seen that the magnitude of the load is the principal factor; as the load increases, tooth wear, as a rule, increases. Under variable load, the amount of wear of gear wheels is on average greater than under equivalent, or close to it, nominal constant loads. However, with a small (optimal) difference in hardnesses of the mating gear pairs (Fig. 3), the difference in wear values under variable and equivalent loads is comparatively small, and, without great error, a constant load equivalent to a variable load in terms of tooth contact strength may be considered equivalent to the variable load also in terms of wear. With a large difference in hardnesses of the mating gear wheels, the amount of tooth wear of the wheel under variable load is considerably greater than under equivalent and close-to-it nominal loads, and approaches the amount of wear under the maximum load. The pinion, as under maximum load, wears little, and the difference in wear values of the teeth of the pinion and the wheel, especially during the running-in period, is great. It should be noted that driving gear wheels, other conditions being equal, as a rule wear more than driven ones.

The tests showed that, with the optimal ratio of tooth hardnesses of the pinion and the wheel, and also with any other combination of hardnesses of the gear pair obtained without lowering the hardness of the wheel and increasing the hardness of the pinion relative to the optimal value, tooth wear (Fig. 3) during the running-in period is small, and its magnitude under all loads is close for both mating gear wheels. The total wear of the gear wheels over the entire testing time is also small and is not catastrophic in character. With the optimal ratio of hardnesses of the gear pair, the wear of the wheel as a whole is greater than the wear of the pinion, but the amounts of wear per tooth are close, since the wheel has a larger number of teeth. With equal tooth hardness of the gear pair, on the contrary, the mating gear wheels under all loads have close wear values; however, the wear per tooth on the pinion is greater than on the wheel, and, as the tooth hardness of the gear pair increases (curves $4_1$ in Fig. 3), the amount of wear of the gear wheels decreases. In gear pairs with increased hardness of the wheel, the amount of wear ...

(curves \(1'''_4\) and \(5'''_3\), Fig. 3) is less than with the optimum hardness ratio; however, here too, as with equal hardness of the teeth of the gear pair, the total amount of wear of the gear wheels is close, while the wear per tooth is greater for the pinion than for the wheel.

Fig. 3. Wear graphs of the gear wheels of the high-speed stage of the ER-5 reducer under: \(1\)—variable, \(2\)—minimum, \(3\)—nominal, \(4\)—equivalent, \(5\)—maximum loads.

Solid lines denote driving gear wheels; dashed lines denote driven gear wheels. Wear graphs of gear wheels at \(\mathrm{HRC}_1 = \mathrm{HRC}_2\) are marked by one prime; at increased hardness of the wheel—by two primes.

The presence on the pinion of a fracture edge on the tooth profiles in the near-pitch zone increases the wear of the wheel teeth (curves \(1_2\) and \(5_1\), Fig. 3), especially during the running-in period; the running-in wear of the pinion teeth in this case decreases. However, this depends on the ratio of the hardness values of the gear pair. Thus, gear pair \(5'''_3\), with increased

with the hardness of the wheel, for comparison with gear pair \(5_1\), which operated under the same load, also had a fracture edge on the pinion tooth profiles in the near-pitch zone. Comparison of the wear curves \(5''_3\) and \(5_1\) (Fig. 3) shows that, in the case of increased wheel hardness, the influence of the fracture edge of the tooth profile, even in the near-pitch zone, is no longer significant. The pinions of gear pairs \(1'_3\) and \(4_2\) had a smooth profile

Fig. 4. Wear curves of the gear wheels of the low-speed stage of the ER-5 reducer under: \(1\)—variable, \(2\)—minimum, \(3\)—nominal, \(4\)—equivalent, \(5\)—maximum loads. Solid lines are driving gear wheels; dashed lines are driven gear wheels. Wear curves of gear wheels at \(HRC_2 = 38\text{–}40\) are marked with a single stroke; in the case of \(HRC_1 = 49\) with a stroke; in the case of \(HRC_1 = 55\text{–}60\), with two strokes.

without a clearly expressed fracture edge on the working surface of the teeth and, as can be seen from Fig. 3, in this case the wear of the teeth of both the pinion and the wheel decreases sharply.

It is evident from Fig. 5 that the wear of the teeth of the helical gear pairs of the ER-16 gearboxes is not optimal, despite the fact that, at practically the same gear ratio, in the principal tests they had the same hardness ratio as the optimum for the high-speed stage of the ER-5 gearboxes, whose gear wheels are spur gears. The wear of the teeth of the helical wheels proved to be considerably greater than the wear of the pinions meshing with them. Evidently, this is a consequence of the fact that at different points of the contact lines of the teeth of helical pairs the total stiffness is not the same, and it has its greatest value near the pole. The influence—

Fig. 5. Wear curves of the gear wheels of the high-speed stage of the ER-16 gearbox.
Solid lines—driving gear wheels; dashed lines—driven gear wheels. Wear curves of gear wheels at $\mathrm{HRC}_1=52$ or $\mathrm{HRC}_2=46$ are marked with one dash; at $\mathrm{HRC}_1=52$ and $\mathrm{HRC}_2=46$—with two dashes.

—ence of this factor proved to be very similar to the influence of the presence of the profile-break edge in the near-pitch zone of the pinions of the spur gear, which, as shown above, led to increased wear of the less hard surface of the gear pair (the wheel). When the hardness of the wheel is increased, its wear (curve $2'$, Fig. 5), as in the spur gear, decreases sharply. Wear of the pinion increases as its hardness is reduced (curve $13'$, Fig. 5). Moreover, as curves $2'$ and $13'$ show,

Fig. 5, increasing the hardness of the wheel or decreasing the hardness of the pinion brings the wear values of the teeth of the gear pair closer together. Increasing the hardness of the wheel should be considered more optimal, since in this case the wear of the teeth of both the pinion and the wheel is lower. Here, too, bringing the hardnesses closer together leads to a reduction in wheel wear and an increase in pinion wear (curves 14″, Fig. 5).

With a large difference in the hardnesses of the mating gear wheels, the ratios of the wear values of the teeth of the pinion and the wheel (Fig. 4) during the running-in period depend to a greater degree on the magnitude of the load. At

Fig. 6. Wear curves of the gear wheels of the low-speed stage of the ER-16 reducer.
Solid lines—driving gear wheels; dashed lines—driven gear wheels. Wear curves of gear wheels at \( \mathrm{HRC}_2 = 37—38 \) are marked with one prime, and at \( \mathrm{HRC}_2 = 48 \)—with two primes.

maximum load, the wear of a wheel tooth is tens of times greater than the wear of a pinion tooth. As the load decreases, this difference diminishes, and at the minimum load (approximately equal to the limiting load in terms of contact strength) the wear values of the pinion and wheel teeth become close. When the hardness of the wheel is increased to \( \mathrm{HRC}_2 = 37—40 \) (the lower limit of the optimum wheel hardness for the high-speed stage of the ER-5 reducer), the wear of the wheel teeth (Figs. 4 and 6) decreases significantly; however, the difference in the wear values of the pinion and the wheel is still very large. This occurs even when, with an increase in wheel hardness to \( \mathrm{HRC}_2 = 38—40 \), the hardness of the pinion was reduced to \( \mathrm{HRC}_1 = 49 \). As

As can be seen from Fig. 4, in this case as well the wear of the gear pair is far from optimal. When the hardness of the wheel is increased to $\mathrm{HRC}_2 = 44\text{--}45$ (the upper limit of the optimal hardness of the wheel of the high-speed stage of the ER-5 gearbox), the amounts of wear of the pinion and wheel teeth (Fig. 6) become somewhat closer, but the difference between them still remains large. Only when the wheel hardness is increased to $\mathrm{HRC}_2 = 48$ do the amounts of wear of the pinion and wheel teeth become sufficiently close, i.e., the hardness ratio of the gear pair approaches the optimal one.

Thus, the optimal hardness ratio depends on the gear ratio. The smaller the gear ratio, the closer the hardness values of the gear pair should be taken. Since the difference in hardnesses both in the case when $\mathrm{HRC}_1 = 49$ and $\mathrm{HRC}_2 = 38\text{--}40$, and when $\mathrm{HRC}_1 = 55\text{--}60$ and $\mathrm{HRC}_2 = 48$, was approximately the same (on the order of 10 HRC), while only the latter is the optimal hardness ratio, it may be concluded that, when the gear ratio is reduced, bringing the hardnesses closer together should be achieved by increasing the hardness of the wheel, and not by lowering the hardness of the pinion.

Increasing the surface finish of the teeth by one class to $\nabla 7$ sharply reduces the wear of the wheel teeth and brings the ratio of the wear values of the gear pair (curves 1—6, Fig. 6) into a region close to optimal.

The presence of a profile-break edge in the near-pitch zone of the pinion teeth also increases wheel wear here, especially with a large difference in hardnesses and a large load (curves $5_1$, Figs. 4, 6). Moreover, wheel $5_1$, which had a significantly greater difference in hardness with the mating pinion than wheel 8 and operated as the driving member, became completely worn down, up to sharpening of the tooth tips, only during the running-in period. Under minimum load, increased wheel hardness, or improved surface finish of the teeth, the presence of a profile-break edge in the near-pitch zone of the pinion teeth, although it does somewhat increase the wear of the wheel teeth, does not do so substantially (curves 22, Fig. 4; $13''$, $1'$, Fig. 6). The absence of a profile-break edge on the pinion teeth (a smooth profile), as in the high-speed stage, sharply reduces tooth wear, especially for the wheel (curves $1_2$, $4_2$, $5_4$, Fig. 4). However, because of the large difference in hardness of the teeth of the gear pair, its wear even in this case is not optimal: the wheel teeth wear more strongly than the pinion teeth, especially under heavy load.

The pinions of gear pairs $4_1$, $4_2$, $5'_5$ (Fig. 4) and $10'$ (Fig. 6), after heat treatment, had a “warty” surface (Fig. 1, a). As can be seen from the wear graphs, the presence of such a defect on the working surface of the pinion teeth, all other conditions being equal, greatly increases the wear of the wheel teeth. Moreover, from a comparison of curves $5'_5$ and $5_3$, $4_2$ and $3_2$ (Fig. 4), it is evident that a “warty” surface nullifies the beneficial effects of such factors as a close hardness of the gear pair and a smooth profile of the pinion teeth.

Thus, analyzing the results of the tests carried out, it may be concluded that, when a gear pair operates with considerable overload, optimal wear of the pinion and wheel teeth can be achieved in various ways, for example, by using an optimal hardness ratio of the pinion and wheel teeth, by improving the finish of the working surfaces and the accuracy—smoothness—of the tooth profile (especially of the pinion, as the harder element). It should be noted that, since for any of the hardness ratios used in our experiments the difference in the amounts of wear of the pinion and wheel teeth increases with increasing load, and at minimum load this

the difference is relatively small even when there is an edge at the break in the pinion tooth profile in the near-pitch zone, preliminary running-in of the gear pair under a reduced load (up to the limiting load in terms of contact strength) is very useful. The use of one or another of the measures named, or all of them together, depends on the capabilities of the manufacturing plant, and on the size and purpose of the gear wheels. In particular, for the gear wheels of mining electric-drill gearboxes, which have small dimensions for both the pinion and the wheel, it is apparently most expedient to use the optimal hardness ratio of the gear pair. To avoid increased wear, the pinion should not be allowed to have a “warty” surface or an edge at the break in the profile in the near-pitch zone.

REFERENCES

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3. Yu. S. Semenov. Results of an experimental study of wear of gear wheels in gearboxes of mining electric drills. Izvestiya TPI, vol. 108. Study of Drilling Machines. Metallurgizdat, Sverdlovsk, 1959.

4. Yu. S. Semenov. Change in the parameters of gear meshing under wear. Izvestiya TPI, vol. 96, part 2, Tomsk University Press, Tomsk, 1961.

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6. Yu. A. Geller. Tool steels. Metallurgizdat, Moscow, 1964.

7. I. M. Lyubarskii, O. F. Podgornaya. Running-in of heavily loaded gears. Increasing the Wear Resistance and Service Life of Machines. Mashgiz, Kiev—Moscow, 1953.

8. Yu. S. Semenov, V. T. Gorbenko. Wear and cleanliness of the working surfaces of gear teeth. Izvestiya TPI, vol. 114, Tomsk University Press, Tomsk, 1964.

9. V. T. Gorbenko, S. I. Shubovich. On the wear of gear wheels in gearboxes of mining electric drills when operating with liquid lubrication. Izvestiya TPI, vol. 133, Tomsk University Press, Tomsk, 1965.