Reducing the vibration of bearing units of electric vehicle asynchronous traction motors

2.1. Analytical methods for studying vibration characteristics of bearings

The available methods used for carrying out vibration analysis of electrical machines at the design stage do not make it possible to take into account many constructive and technological factors. The discrepancy between the design and actual levels in EMs can reach 40 dB (Migal, 2002). When carrying out vibration analysis, they make a number of unjustified assumptions, which do not make it possible to obtain the reliable results and recommendations for improving the design of electric drives with very stringent vibration requirements (category S and stricter) IEC 60034-14: 2007, ISO 29.160.

The main challenges are due to the account of real values of operational clearances, the rigidity distribution, masses and natural frequencies of mating parts, units and systems, the availability of many detachable joints, temperature effects, friction, the quality of bearings, etc. The technical problem of reducing the vibration of EMs should be managed by increasing the accuracy of manufacturing the rotor and bearing units, reducing the radial electromagnetic excitatory forces and the vibration they generate at the frequencies of gear harmonics, the eccentricity of the rotor, and in the «rotor–stator» gap.

However, the vibration levels of manufactured EMs both for electric and hybrid vehicles continue to remain high and do not meet the requirements of the developed vibration classes, Figure 1.

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Figure 1.

Ranges of electric motor vibration classes.

In addition, the achieved increase in equipment reliability by reducing the vibration is not sufficient to implement the typing of design and technological solutions as applied to vehicle traction motors. To take account of the specified factors, it is necessary to conduct full-scale experimental studies for reducing the vibration levels of asynchronous EM bearing units to vibration rates D, E, and F, Figure 1.

The straight line AB was adopted as permissible vibration levels of asynchronous motors, which is limited by the excess vibration level to 40 dB at a frequency of 5 Hz and 80 dB at a frequency of 10,000 Hz (Migal, 2002; Migal et al., 2019). In electric and hybrid vehicles, it is not recommended to use EMs, the vibration levels of which exceed the AB line.

It should be emphasized that this article is the result of a number of experimental studies. Theoretical analysis of the vibrational characteristics of bearings is the subject of different studies. Theoretical calculations of the level of vibration usually have a form of a qualitative analysis of vibration sources. It is not possible to quantify with high accuracy the effect of all sources of vibration of bearing assemblies on the vibration levels. This is because the impact of all vibration sources, connected with each other, occurs at certain frequencies (Arhun et al., 2020a, 2020b).

2.2. Methods and means of experimental research

From the point of view of improving the reliability and lifetime, the most hazardous are vibrations in the frequency range from the operating speed to 1000–1250 Hz. The main sources of vibration in this frequency range are: imbalance, the quality of rolling bearings, the manufacturing quality of bearing units (geometry of fitting surfaces, their misalignment, circularity deviation and nonperpendicularity of mating surfaces, and surface cleanliness), the operating clearance in the «bearing–housing» system, and static and dynamic eccentricity of the «rotor–stator» air gap. Practice has shown that computational methods used for determining the levels of vibration excited by the listed sources have little convergence with real values in EMs (Migal, 2002; Migal et al., 2019).

To determine the rational clearance of the bearing fit in the housing (primer), one proceeds from the following requirements:

Reliable guaranteed axial movement of the bearing in one support (primer), while compensating for temperature changes in the dimensions of parts;

In this regard, it was experimentally determined under which bearing fit clearances in the primer and the shaft preload the damping properties of «clearance-preload» and «floating» of the bearing in the housing, which causes a decrease in the forced oscillations of the bearing races and the shaft with the bearing in the housing, the damping properties are retained.

The clearances in the interferences and the rigidity of parts fastening significantly affect the change in the natural frequencies of EM technical systems. With a rigid reliable fixing of parts, the conditions for the total mass of joined parts are created, and with a decrease in the matching rigidity, the conditions for occurrence of gaps appear; each mated part has its own oscillation frequency.

The results of experimental studies were processed using the methods of mathematical statistics, in particular, the method of the series criterion, based on the median of the sample and the criteria of the squares of sequential relations, and the method of nonparametric analysis of variance using the Friedman criterion (Arhun et al., 2020a, 2020b).

2.3. Means for studying the vibration characteristics of bearings and bearing units

To trace the impact of each element and factor on the vibration levels excited by bearing units, the effect on vibration levels of dimensions and quality of bearings, then the effect of fit on the shaft mandrel, the clearance between the outer bearing race and the housing, and the method of mounting the bearing into the housing were studied.

The results, discussed in the article, were obtained during the field experimental studies. The research was conducted in the vibroacoustic laboratory of the Research Electrotechnical Institute of the Research and Production Association of Kharkiv Electrotechnical Plant. Testing of individual bearings and bearing units was carried out on a standard stand VNIPP-508K, as shown in Figure 2. The vibration values of the bearings and bearing units were measured by vibration analyzers 2033 and 2034, with recording of the vibration spectra by the plotter 2313. Acceleration a = 3 × 10⁻⁴ ms⁻² was taken as the zero vibration level.

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Figure 2.

Stand VNIPP-508K: (a) with the tested bearing and (b) with bearing unit: 1—axial loading unit of the tested bearing and bearing unit, 2—spring damper with the wave spring and textolite case, 3—rubber vibration absorber, 4—bearing unit, 5—removable shaft mandrel, 6—elastic shell coupling, 7—vibration absorbers, 8—two-speed asynchronous motor, 9—spindle head, 10—place for mounting vibration converter, and 11—device of axial loading of the bearing and bearing unit.

The radial ball bearings of the normalized series according to ISO 492: 2014 were investigated. The bearings are manufactured according to TU 4477-E-82, the accuracy class 75–170, and the base frame size №№ 306, 310, 313, 317, and 322 with noise indexes QE1 ... QE6.

The vibration measurement of each bearing was repeated four times. A sample of each dimension type and accuracy class of bearings constituted 10 pieces.

Rolling bearings were tested with an axial preload force according to the MVM VNIPP.002-04 measurement method and the rotating frequency of the inner race of 750 and 3000 r/min, which meet the conditions for the main operation of bearings in the EM bearing units. The axial force was applied at three points to the bearing end, Figure 2(a).

Bearing vibration measurement was performed using the root mean square values of vibration accelerations in decibels (dB) in the radial direction when mounting a vibration converter of type 4376, and the vibration measurement of bearing units (Figure 2(b)) was performed with a rigid mounting of the vibration converter by a cap nut (Migal et al., 2019). The measurement of vibration in the housing was repeated 3 times. The actual dimensions of the fitting diameter of the shaft mandrel (Figure 2(b)), the openings in the bearing housing, and the dimensions of the bearings proper made it possible to install the outer race in the housing with a clearance of 5 μm–60 μm and the mandrel shaft with a preload from 5 μm–25 μm. To eliminate the impact of the fitting effort from the shaft journal on the bearing quality, the initial vibration characteristics of the latter were each time monitored at the VNIPP-508K test bench.

3.1. Effect of rolling bearing quality on vibration levels of EMs

The effect of the rolling bearing quality on the vibration levels of asynchronous motors with a power of 30 kW and a rotational speed of 3000 r/min is shown in Figure 3, where 1 is the vibration levels of the EM on the bearings of normalized series and 2 on the low-noise bearings of class QE3. Motor vibration levels on QE1 and QE2 bearings are intermediate values. Figure 3 shows the average values of vibration measured for three EMs.

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Figure 3.

Average values of vibration levels of electric motors in one-third octave bands.

From the above spectra, it can be seen that the use of low-noise bearings significantly affects the vibration levels of the EM. However, none of the tested low-noise bearings (QE1, QE2, and QE3) meets the requirements of the straight line AB of the permissible vibration of EMs both for electric and hybrid vehicles (Figure 1). It remained to determine by what means the quality of bearings directly contributes to the vibration levels and the effect caused by the working conditions of the latter in bearing units, namely, the bearing shaft preload, the bearing-primer clearance, bearing inner race misalignment on the shaft with perpendicularity of the shaft end, and the shaft journal ovality.

3.2. Impact of the size, accuracy class, and the operation mode of bearings on vibration levels of bearing units

The results of vibration measurements at a rotation frequency of 1500 r/min are reduced to the typical vibration spectrum shown in Figure 4.

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Figure 4.

Typical spectra of the bearing and bearing unit vibrations with rigid fixation of the bearing outer race with the cap: (a) the J, K, and C levels of bearing vibrations for noise classes QE1–QE6, (b) a typical spectrum of one bearing vibration and the straight line AB of the equal resource limitation of maximum allowable vibration, and (c) testing of bearing assemblies on bearings of accuracy class P0.

Figure 4(b) shows a typical spectrum of one bearing vibration and the straight line AB of the equal resource limitation of maximum allowable vibration values of straight line AB at frequencies 50, 2000, and 10,000 Hz for each tested bearing and bearing unit.

Figure 4(a) shows the J, K, and C levels of bearing vibrations for noise classes QE1–QE6.

Vibration measurements were carried out on the heads of the vibration characteristics of all tested bearings of accuracy classes P0 and P5 (QE1–QE6) of standard sizes 313 and 322 exceeded the straight line AB by 2–6 dB in the frequency range of 2500 Hz. Bearing vibration levels of standard size 306, the accuracy class, and noise level QE1–QE6 were lower than the straight line AB by 1–4 dB. The maximum vibration levels of bearings were in the region of the resonant frequency of their outer races.

Testing of bearing assemblies showed that their vibration levels on bearings of accuracy class P0 of all tested standard sizes (from 306 to 322) exceed the straight line AB by 3–10 dB (Figure 4(c)) at rotation frequency of already 1500 r/min.

It is possible to control the vibration levels of the bearing units in the high-frequency region (2000–10,000 Hz) by changing the rigidity of installation of the bearing outer race into the housing, the axial preloading force, and the rotational speed. Changing the rigid method of fixing the bearing by caps in the housing to an elastic one, when a wave spring is installed between the bearing and the cap (G and H), reduces the vibration levels in the high-frequency region by 3–5 dB. The vibration levels of bearing units when the bearings are fixed by a cap with an axial clearance are increased by 2–3 dB compared with the rigid fixation. The resonant frequency of the rolling bearing outer race, depending on the size, type, and rigidity of the bearing in the housing, was in the range of 2000–3150 Hz (see. Figure 4).

Analysis of the vibration characteristics of bearings of different sizes shows that with an increase in the nominal diameter of the bearing to the diameter of the bearing of the next standard size, the vibration increases linearly by 1–3 dB. Therefore, increasing the size to enhance the rigidity and durability of the bearing unit is not always advisable. Minimum vibration levels of the bearing unit are achieved by installing radial ball bearings one at a time in the support.

The use of rolling bearings of class P0 does not make it possible to obtain the class D vibration of EM bearing units (see Figure 4).

By using bearings of the fifth class of accuracy and noise, QE1–QE6, it is possible to reduce the vibration by 4–10 dB and to ensure the technical level of EM bearing units with vibration classes D and E, Figure 1. The guaranteed reduction of bearing units vibration up to class D can be obtained by using bearings 306 and 313 of already noise class QE1.

Reducing the vibration of bearing units by 16 dB to class E is possible using bearings 306 and 310, which makes it possible to double the life of induction motor bearing units compared with using bearing units of accuracy class P0.

However, it is possible to implement the vibroacoustic properties of the above bearings, while ensuring the appropriate quality of the fitting surfaces of the shaft and the housing of the bearing units, the manufacturing accuracy and uniqueness of the bearing unit assembly, and the motor as a whole. The effect of rotating frequency and load. Tests of radial ball bearings were carried out at rotating frequencies from 750 to 5000 r/min when powering the drive of the VNIPP-508K stand with an adjustable frequency of 25–75 Hz from a synchronous generator. Bearings of accuracy class P5 and noise from QE1 to QE5 and sizes 306 and 310 with a radial clearance of 10–20 μm were tested.

The effect of rotating frequency and load on vibration levels of rolling bearings is shown in Figure 5. It shows that an increase in the rotating frequency of bearings has a much greater effect on the growth in vibration compared with an increase in the radial load Fr.

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Figure 5.

Dependence of vibration levels of bearings: (a) on rotating frequency and (b) loading: 1—type of bearings 310Е4Ш1, 2—type of bearings 306Е4Ш5, and 3—is the spread of vibration levels (averaged over four measurements) for sampling of six bearings of each type.

3.3. Determination of the rational parameter of bearing press fit on the shaft journal

The bearing preload installation on the shaft excludes the cranking of the bearing inner race on the shaft, increases the rigidity of the bearing unit, and affects the spectrum of eigenfrequencies of both the bearing and the rotor. The damping properties of the coupling «bearing–shaft» are determined by the value of the shaft bearing preload, the radial bearing gap, the magnitude and nature of the acting loads, and vibration. The shaft bearing preload affects the level of bearing unit vibration within a wide frequency range excited, mainly, by periodic changes in the bearing’s rigidity at ball pass, bearing race misalignment, wavelengths and the granularity of bearing elements and rolling tracks. The low radial rigidity of bearing races causes them to copy all the defects of the shaft journal fittings.

The shaft bearing preload 5, 11, 20, and 26 μm was determined by selective fitting of the bearing and mandrel dimensions. The rational shaft journal rolling bearing preload of noise class QE3–QE6 is the preload of 10–15 μm, Figure 6.

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Figure 6.

Dependence of vibration levels of bearing units on the preload pressure of shaft–mandrel journal with the bearing rigidly fixed in the housing with a clearance of 10 μm.

Reserves for reducing the vibration levels of bearing units in determining the optimal shaft journal bearing preload for most frequencies are within 5–8 dB (Figures 6 and 7). With a gap of 40 μm in conjunction «the bearing outer race housing», the maximum vibration reduction is achieved as expressed by curve 2 in Figure 7(b).

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Figure 7.

Dependence of vibration levels of bearing units on fitting pressure to the neck of shaft–mandrel with different fixing of the bearing in the housing (primer): (a) rigid, (b) with an axial clearance, and (c) elastic, with the wave spring; 1, 2, and 3—fitting clearances of the bearing in the housing: 10, 40, and 60 μm, respectively.

3.4. Determination of rational parameters forbearing installation into the housing

Studies to determine the optimal clearance of bearing installation into the housing and its press fit on the shaft were carried out with roughness parameters of the preload surfaces of the housing Ra = 1.23 mm. In case of deviations in ovality and taper rate to a maximum of 4 μm, with a total power of the bearing preload-850 N, the dimensions of the housing fitting surfaces and the fitting dimensions of test bearings allowed the installation of bearings with diametrical clearances of 10, 40, and 60 μm. The bearings and the fitting areas were lubricated with grease CIATIM-221. The test bearings were mounted into the bearing unit housing according to three fixing schemes: bearing installation into the housing with an axial clearance between the end of the bearing outer race and the cap, rigid fixation of the bearing outer race by the cap, and elastic fixation of the bearing outer race by the wave spring with a power of 800 N.

Vibration spectra (Figure 8) are constructed based on the arithmetic mean values of vibration levels of six measurements. Comparison of the third-octave vibration spectra of separate bearings (1) and the bearing unit (2) points to a significant difference of the latter. The displacement of resonance frequency of a separate bearing (2000 Hz) and bearing units (3150 Hz) is explained by the parameters of the «shaft–bearing», «bearing–housing» systems in the bearing units, taking into account the bearing press fit on the shaft and rigid fixation in the housing.

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Figure 8.

Vibration spectrograms: 1—of a separate bearing, 2—the bearing unit, 3—the stand, and 3—VNIPP-508κ.

Figure 9 shows the dependences of vibration levels of separate bearing units on the diametrical clearance of the bearing fitting into the housing at certain frequencies. The points indicate the maximum, arithmetic mean, and minimum vibration levels of 15 tested separate bearing units.

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Figure 9.

Dependence of vibration levels of the bearing unit with the rotation frequency of: (a) 1450 r/min and (b) 2000 r/min on the bearing fitting clearance in the housing with the frequencies of: 1—160 Hz, 2—200 Hz, 3—250 Hz, and 4—630 Hz.

The maximum effect of the bearing clearance fit into the housing on the levels of bearing unit vibration is manifested in the frequency region of the periodic change in bearing stiffness and the natural frequencies of its parts. The variation of vibration levels of separate bearing units at these frequencies, depending on the bearing clearance fit into the housing and the rotation speed, reaches 10 dB. The maximum reduction in the vibration levels of the bearing unit within the range of 160–250 Hz frequency caused by the periodic change in the bearing stiffness, due to the rolling element misalignment and various dimensions of the latter, was obtained with a diameter of 20–23 μm, and at alternate frequencies, it was obtained with a clearance of 30–40 μm.

From Figure 9, it follows that for each vibration frequency there is an optimum bearing clearance fit into the housing, at which the maximum reduction of bearing unit vibration levels is achieved. The rational clearance of bearing installation into the housing, in terms of minimizing the vibration across the entire spectrum, is the diameter clearance of 20–40 μm. With such a clearance, in most cases, the maximum reduction in the levels of vibration of bearing units is attained.

Thus, the bearing clearance fit into the housing considerably affects the vibrodiagnostic characteristics of the bearing unit. Reserves for lowering the vibration levels of bearing units in determining an optimal bearing clearance fit into the housing can reach from 5 to 10 dB at most of the vibration frequencies.