Разработка метода определения величин зазоров в подшипниках скольжения Текст научной статьи по специальности «Механика и машиностроение»

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DEVELOPMENT OF A METHOD FOR DETERMINING THE SIZE OF CLEARANCE IN SLIDING BEARINGS

Nikola P. Zegarac

cm Serbian Academy of Inventors and Scientists,

cm Belgrade,Republic ofSerbia,

of e-mail: zegaracnikola@vektor.net,

^ ORCID iD: https://orcid.org/0000-0002-1766-8184

§ DOI: 10.5937/vojtehg68-26107; https://doi.org/10.5937/vojtehg68-26107

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FIELD: Mechanical Engineering, Energetics, Shipbuilding ARTICLE TYPE: Original Scientific Paper

x Abstract:

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w Introduction/purpose: The purpose of this paper is to present the

>- importance of applying a new diagnostic method with a monitoring system

< and its capability to reliably determine when and where the problem will occur related to the wear of plain bearings in further operation of the plant, to offer a quality assessment of how the system will continue to function over time, to predict causes of failures and how to eliminate them as well as to provide time for planned maintenance of technical systems.

S

< Method:The new method solves the problem of sliding bearing diagnostics u by measuring the dynamic trajectories (orbit) of the sleeve in the sliding g bearing. Modern methods of technical diagnostics based on the

measurement of dynamic parameters and electrical quantities and their

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analysis enable the support of measuring systems and measurement sensors from different manufacturers.

Results: Measurements of the dynamic trajectories of sleeves in sliding ^ bearings determine the values characterizing: normal condition, initial

clearance size, further clearance increase, bearing clearance sizes, the moment when the state parameters are close to the upper limit of the allowable clearance, and the specific clearance size when further plant exploitation can cause system failure.

Conclusion: The new diagnostic method and monitoring systems can be widely applied in all technical fields: internal combustion engines, hydroelectric power plants, thermal power plants, process plants, and in many other areas. Many monitoring systems, compatibility of devices and equipment of different manufacturers are supported by hardware and software. The calculations of theoretical and experimental dynamic parameters were verified. The method has a wide range of possibilities of application.

Keywords: sliding bearing, bearing clearance, bearing wear, bearing sleeve, dynamic trajectory.

Introduction

The new procedure and measuring systems (hereinafter referred to as monitoring systems) for determining the clearance size, i.e. the degree of wear of sliding bearings, is a new diagnostic method in the technical field. It is is characterized by the following: it allows direct measurement of the movement parameters of the sleeve in the sliding bearing, thus enabling automatic quality control of plants in which sliding bearings are built into. The monitoring system has an almost unlimited lifetime, has no transferable and wear-exposed elements, does not depend on the speed of rotation and overload of a plant, in the case of internal combustion engines, and does not depend on the engine type and stroke. The new method is used as a precautionary measure to prevent possible system crashes. The monitoring system can be built into machinery during production, so there is no limit if it is installed on a plant already in operation. Monitoring systems can be implemented in the online version, i.e. as a constant monitoring system or in the off-line version, i.e. as an occasional monitoring system, depending on the plant type and the requirements of the plant user (Cernej et al, 1986).

Theoretical settings, dynamics of sliding bearings, and the calculation of dynamic parameters

There is a range of impacts such as defective lubrication, dirty and diluted lubricating oil, incorrect bearing geometry due to manufacturing and assembly errors, or large bearing deformations caused by dynamic forces during plant operation, which causes damage to sliding bearings. High temperatures in sliding bearings are most often due to:

- transient mixed friction with a longer duration, mixed friction in a larger zone along the bearing range or due to mixed friction of higher intensity,

- lack of lubricating oil for a certain period of time or a permanent lack of oil in the sliding bearing.

The causes of mixed friction may be different or a combination of multiple impacts, such as (Licen & Zuber, 2003):

- mistakes in making liners and sleeves,

- small and large bearing clearances,

- excessive mounting or working deformation of the bearing,

- clogging of oil channels with dirt in the lubrication system,

- small lubricant pump capacity,

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- transmission of dynamic forces, and

- filtration and cooling of lubricating oil.

«3 Fatigue of the material in the sliding bearing occurs if the limit of the

° dynamic endurance of the bearing material exposed to cyclically variable pressures in the oil film is exceeded. Today's limits, with regard to fatigue and bearing wear, are (Cernej et al, 1986): of - minimum oil film thickness obtained by calculation - 2 ^m,

- maximum pressure in the oil film up to 250 MPa, g - maximum temperature in the oil film up to 140°C, and o - hydrodynamic specific friction force in the bearing up to 0.15

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CH The bearing pressure of the bearing lubricating oil is determined by

m the bearing capacity characteristic and is in the range from 90° to 130° of the bearing circularity. In the rest of the clearance, the lubricating oil

< pressure is approximately equal to the oil supply pressure in the sliding bearing.

Knowledge of the parameters of the dynamic orbit of the sleeve in the sliding bearing leads to quality indicators about the relative w performance of certain drive and design parameters such as: bearing dimension, the size of the clearance in the bearing, speed of rotation, 2 viscosity of lubricating oil, and the like.

1 The dynamic orbit (trajectory) parameters of the sleeve in the bearing are determined by:

- the minimum thickness of the oil film, which is closely related to the o bearing to^

^ - maximum pressure in the oil film,

- the friction magnitudes in the bearing (energy losses) depending on the bearing temperature regime, and

- the amount of oil flow in the bearing.

In order to theoretically obtain the dynamic trajectory of the sleeve in the sliding bearing, the input data requires a dynamic load force of the bearing and the angle at which it acts. The basis for obtaining the dynamic trajectory is based on the balance of the dynamic force (F) loading the sleeve and the hydrodynamic forces due to the pushing of the sleeve (Fp) and turning the sleeve in the bearing (Fo), (Figure 1), (Zegarac, 1989).

f = y-180

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Figure 1 - Schematic presentation of a cylindrical sliding bearing and the equilibrium

position of dynamic forces Рис. 1 - Схематическое изображение цилиндрического подшипника скольжения и

положение равновесия динамических сил Слика 1 - Шематски приказ цилиндричног клизног лежаjа и равнотежног положаjа

динамичких сила

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Based on the balance of dynamic forces, the exact position of the sleeve motion in the sliding bearing is obtained at each moment of the plant cycle time.

For realistic operating conditions and different directions of rotation of the sleeve in the bearing in Figure 2, possible positions of the non-stationary loaded bearing (Lang & Steinhilper,1978) are shown.

From equation (1), the value for the dynamic force is obtained:

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Fp = F

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From equation (2), the value for the dynamic force is obtained:

Fo = F

sign(S - r)sin(S - r) sin 3

(4)

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Figure 2 - Possible balance positions of the sleeve of a non-stationary loaded sliding

bearing

Рис. 2 - Возможные положения равновесия рукава нестационарно нагруженного

подшипника скольжения Слика 2 - МогуПи положаи равнотеже рукавца нестационарног оптереПеног

клизног лежала

By introducing known ratios for Sommerfeld's bearing characteristics for components due to the rotation and thrusting of the sleeve, the parameters of the dynamic trajectory of the angular velocity of the sleeve in the bearing are obtained (Lang & Steinhilper,1978):

So0 =

SoP =

©

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(5)

(6)

The effective angular velocity of the sleeve

(7):

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m — m - msp — m i + m 2 - 2 5

is given by equation

(7)

where is:

©j - angular velocity of the sleeve,

©2 - angular bearing speed (©2 = 0, in the case of the non-rotating main sleeve bearing),

© - relative angular velocity of the sleeve in the bearing,

© = - 2 dL - is a hydrodynamic effect in the bearing, stopping the

sp dt

sleeve and causing it to rotate in the opposite direction, 5 - angular velocity of the minimum thickness of the oil film,

y- relative bearing clearance calculated according to equation (8):

R - r

W — where is:

Z_

r

(8)

R - bearing radius, r - bearing sleeve radius, and Z - radial bearing clearance.

Finally, the value for ©

sign(5 - y)F sin(5 - y)y2

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m — m-25

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The value for the radial velocity of the sleeve center (s) is obtained from equation (10), (Lang & Steinhilper,1978):

s =

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sin S - y |

tg p

(10)

The value for the angular velocity of the minimum thickness of the oil film (S), is obtained from equation (11):

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F y/2 sin( S - y) BD iSo 0 sin p

(11)

An internal combustion engine of the type 6ASL-25D, manufactured by Jugoturbina - Karlovac, licensed by Sulzer company, Switzerland, was chosen as an example of modeling and calculating dynamic parameters in order to determine the size of clearance in the sliding bearing.

Determining clearance in sliding bearings of internal combustion engines is very complex. In this case, the main bearings of the engine crankshaft were the subject of investigation.

Figure 3 shows the kinetostatic model of the motor mechanism, on the basis of which the calculation of all dynamic parameters will be performed: the speed of the parts of the motor mechanism, acceleration, the dynamic forces that load the sliding bearing and the calculation of the parameters of the dynamic orbit, on the axis, which will determine the clearance size, i.e. bearing wear. A dynamic calculation was performed in a completely new way. The results of the calculations were obtained by the help of the software developed for this purpose (Zegarac, 1989).

Figure 4 shows the change in the intensity of the dynamic loading force (F) acting on the sleeve of the 1st main bearing, depending on the rotation angle of the crankshaft (a = 0° to 720°), during the engine operating cycles, (Zegarac, 1989).

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Figure 3 - Kinetostatic model of an internal combustion engine

Рис. 3 - Кинетостатическая модель двигателя внутреннего сгорания

Слика 3 - Кинетостатички модел моторног механизма са унутрашшим

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Figure 4 - Changing the intensity of the dynamic force (F) acting on the sleeve of the 1st main bearing depending on the angle of rotation of the crankshaft engine (a=0° to 720°)

Рис. 4 - Изменение интенсивности динамической силы (F), действующей на рукав первого главного подшипника, в зависимости от угла поворота коленчатого вала двигателя (a=0° до 720°),

Слика 4 - Промена интензитета динамичке силе (F) ща делуе на рукавац првог главног лежаjа у зависности од угла заокрета коленастог вратила мотора (a=0°

до 720°)

Figure 5 shows the change in the angle (y) at which the load force (F) acts on the main bearing sleeve depending on the crankshaft rotation angle (a = 0° to 720°):

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Figure 5 - Changing the angle (y) at which the force (F) acting on the main bearing sleeve depends on the the crankshaft rotation angle (a=0° to 720°)

Рис. 5 - Изменение угла (у), под которым сила (F), действующая на рукав главного подшипника, зависит от угла коленчатого вала двигателя (a=0° до

720°)

Слика 5 - Промена угла (у) под щим сила (F) делу^е на рукавац главног лежала у зависности од угла заокрета коленастог вратила мотора (a=0° до 720°)

Figure 6 shows the hydrodynamic forces (Fo) and (Fp) depending on the crankshaft rotation angle (a) on the 1st main bearing of the engine. It presents the sum of the forces (EFy), (EFx), the influence of the dynamic forces from all engine cylinders on the 1st main bearing of the engine in the horizontal axis (y) and the vertical axis (x).

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The radial clearance in the bearing is Z = 124 ^m, (Zegarac, 1989):

Figure 6 - Changing the intensity of the hydrodynamic forces ZFx, SFy depending on the the crankshaft rotation angle (a = 0° do 720°) a) vertical axis of the motor, b) horizontal axis of the motor

Рис. 6 - Изменение интенсивности гидродинамических сил ZFx, ZFy, в зависимости от угла поворота коленчатого вала (a = 0° do 720°) а) вертикальная ось двигателя, б) горизонтальная ось двигателя

Слика 6 - Промена интензитета хидродинамичких сила ZFx и ZFy у зависности од угла заокрета (a = 0° до 720°) коленастог вратила мотора: а) вертикална ос мотора, б) хоризонтална ос мотора

Figure 7 shows the calculated dynamic trajectory of the sleeve on the 1st main bearing of the crankshaft, at 100% engine load and an engine speed of n = 720 min-1 when the value of the radial clearance is Z = 124^,m. In the dynamic trajectory, the crankshaft rotation angle (a = 0° to 720°) is indicated. The angle value (a = 0°) indicates the start of the expansion in the engine, (Zegarac, 1989).

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Figure 7 - View of the calculated dynamic trajectory of the main sleeve on the 1st main

bearing on the engine crankshaft Рис. 7 - Изображение измеренной динамической траектории главного рукава в первом главном подшипнике двигателя коленчатого вала Слика 7 - Приказ прорачунате динамичке путане главног рукавца на првом главном лежаjу коленастог вратила мотора

Experimental research

The experimental studies were performed on a marine diesel engine type 6ASL-25D, manufactured by Jugoturbina - Karlovac, licensed by Sulzer company, Switzerland, at different engine modes and engine speeds up to a maximum speed of n = 720 min-1. All the preparations for the measurement were made, the operating parameters of the engine were adjusted, the monitoring system was installed and the measurement system calibrated (Easy-Laser, 2020), (Zegarac,1993).

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If the plant is in operation and there is a need to install a new monitoring system, the installation and centering of contactless encoders to measure the center displacement of the sleeve center can be S3 performed with a laser centering system of the sleeve center relative to ° the bearing center (Zegarac, 2016), (Licen & Zuber, 2003).

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of maximum speed of n = 720 min-1.

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Dynamic trajectories at partial engine loads were also measured, up

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Figure 7 shows the measured dynamic trajectories of the main 5 sleeve on the 1st main bearing of the crankshaft engine at 100% engine

m load and a speed of n = 720 min-1.

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¡5 The size of the radial clearance in the sliding bearing is Z = 124^,m.

The figure also shows the size (e) that represents the displacement of the sleeve center relative to the bearing center, i.e. the eccentricity of w the sleeve center, which is directly related to the bearing clearance (Z).

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2 In order to compare the calculated and measured sizes of the

bearing clearance, the term of so-called relative eccentricity (s) was introduced and determined by the equation:

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e

° * = y (12)

Figure 8 shows the angle (S) of the minimum thickness of the oil film.

On the basis of the measured values (e) and (S), the clearance in the sliding bearing and the assessment of further plant exploitation are determined in order not to cause a plant failure.

Figure 8 - Display of the measured dynamic sleeve trajectory on the 1st main bearing of the crankshaft, at the engine load of 100%, n = 720 min-1, and the bearing clearance Z=124 (/m)

Рис. 8 - Изображение измеренной динамической траектории рукава в первом главном подшипнике двигателя коленчатого вала, при нагрузке на двигатель (100%), n = 720 мин-1, зазор подшипника Z = 124 (/um)

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Figure 9 displays the measured dynamic trajectory of the sleeve on the 1st main bearing of the crankshaft engine at the load of 0%,n = 720 min-1, and the bearing clearance Z = 124(^,m):

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Figure 9 - Display of the measured dynamic trajectory of the sleeve on the 1st main bearing of the crankshaft, at the engine load of 0%, n = 720 min-1, and the bearing

clearance Z = 124 (/am) Рис. 9 - Изображение измеренной динамической траектории рукава в первом главном подшипнике двигателя коленчатого вала, при нагрузке на двигатель (0%), n = 720мин-1, зазор подшипника Z = 124 (/им)

Отображение измеренной динамической траектории втулки на 1-м главном подшипнике коленчатого вала при нагрузке двигателя 0%, n = 720 мин-1 и зазоре

подшипника Z = 124 (иm) Слика 9 - Приказ измерене динамичке путане рукавца на првом главном лежаjу мотора коленастог вратила мотора, при оптереПеъу мотора (0 %), n=720min-1,

зазор лежаjа Z=124 (иm)

Based on the results of measuring the displacement of the center of the sleeve (e) for this engine type, the dependence (Z) of (e) was determined (Zegarac, 1989):

Z=1.01- e+7,48±4 (13)

The results are presented in Figure 10.

Figure 10 - Graphic depiction of the clearance size (Z) and the displacement of the center of the sleeve (e) for different rotation speeds and engine loads

Рис. 10 - Гоафическое изображение размера зазора (Z) и смещения центра рукава (e) для разных скоростей вращения и нагрузок на двигатель

Слика 10 - Графички приказ зависности величине зазора (Z) и помака средишта рукавца (e), за различите бризине вртше и оптереПеъа мотора

Comparison of the theoretical calculation results and the experimental research results

Figure 11 gives a graphic representation of the calculated and measured values of the sleeve eccentricity (e) depending on the angle of rotation of the crankshaft (a) for various bearing wear degrees, at a rotation speed of n = 720 min-1 and at 100% engine load, (Zegarac, 1993):

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Figure 11 - Display of the calculated and measured eccentricities of the sleeve (e) on the

1st main bearing depending on the crankshaft rotation angle (a) for different sizes of bearing clearance Z=84 (/m), Z=124(/m), Z=144 (/m), engine speed n = 720 min-1 and

100% engine load

Рис. 11 - Изображение рассчитанных и измеренных эксцентриситетов рукава (e)

первого главного подшипника, в зависимости от угла поворота коленчатого вала (а) при разных размерах зазора подшипника Z = 84 (um), Z = 124(/m), Z = 144 (/m), частота вращения двигателя n = 720 мин-1 и нагрузка на двигатель 100%

Слика 11 - Приказ прорачунатих и измерених ексцентричности рукавца (e) на првом главном лежаjу у зависности од угла заокрета коленастог вратила мотора (а) за различите величине зазора у лежаjу: Z=84 (um), Z=124(/m), Z=144 (/m), брзини вртше мотора n =720 min-1 и оптереПеъу мотора 100%

In Figure 12, there is a graphic representation of the measured values of the sleeve eccentricity (e) depending on the angle of rotation of the crankshaft (a) for various bearing wear degrees Z=84 (^m), Z=124 (^.m), Z=144 (^m), at different speeds of rotation, without loading the engine.

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Figure 12 - Display of the measured eccentricities of the sleeve (e) on the 1st main bearing, depending on the crankshaft rotation angle (a) for different sizes of the clearance in the bearing Z=84(/m), Z=124 (/m), Z=144 (/m) and speed, no engine load

Рис. 12 - Изображение измеренных эксцентриситетов рукава (e) первого главного подшипника в зависимости от угла поворота коленчатого вала (a) при разных величинах зазора в подшипнике Z = 84 (/m), Z = 124 (/m), Z = 144 (¡um) и скорость, без нагрузки на двигатель

Слика 12 - Приказ измерених ексцентричности рукавца (е) на првом главном лежаjу, у зависности од угла заокрета коленастог вратила (а), за различите величине зазора у лежаjу: Z=84(/m), Z=124 (/m), Z=144 (¡um) и брзинама вртъе,

без оптереПеъа мотора

Figure 13 gives a graphic representation of the calculated and measured values of the eccentricity angle of the bearing hose (S depending on the angle of rotation of the crankshaft of the engine (a) for various bearing wear degrees, at 100% engine load and the engine

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speed of n = 720 min (Zegarac, 1993):

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Figure 13 - Display of the calculated and measured values of the angle of eccentricity of the bearing hose (S) depending on the crankshaft rotation angle (a) for various bearing

wear degrees, 100% engine load, speed of rotation engine n = 720 min-1 Рис. 13 - Изображение расчетных и измеренных значений угла эксцентриситета рукава подшипника (S) в зависимости от угла поворота коленчатого вала двигателя (а) при различном износе подшипника, при 100% нагрузке на двигатель, скорости вращения двигателя n = 720 мин-1 Слика 13 - Приказ прорачунатих и измерених вредности угла ексцентричности руквца лежаjа (S) у зависности од угла заокрета коленастог вратила мотора (а) за разна истрошеша лежаjа, при 100% оптреПе^у мотора, брзини вртше мотора

n=720 min-1

Based on the comparison of the calculated dynamic sizes and the measurement results, it can be concluded that the correspondence of the results of the calculations and measurements is in the domain of 5%, which is completely acceptable. The repeatability of the measurement values during the measurement was obtained.

The method of calculation and measurement of the functional parameters of the new diagnostic method has been completely verified.

Diagnostics of sliding bearings can be carried out for all modes, rotation speeds and plant loads.

The author of this scientific paper has spent many years of research on many projects related to technical diagnostics of machine and electrical plants, development, research, design and implementation of monitoring systems (Zegarac, 1994), (Zegarac, 2019).

Conclusion

The application of the new diagnostic method and the monitoring system gives an opportunity to reliably determine when and where the problem will arise related to the wear of sliding bearings in further plant operation, to offer a quality assessment of how the system will continue to function over time, as well as to predict the causes of failures and how to remedy them, and to provide time for routine maintenance of technical systems. Modern methods of technical diagnostics based on the measurement of dynamic parameters and electrical sizes and their analysis allow the support of measuring systems and measurement sensors from different manufacturers. The capabilities of a very advanced program configuration for system diagnostics and monitoring are presented. The values of the functional parameters were measured and software processed as well as their limit values and alarms displayed in different ways.

There are options for remote control and monitoring, shutting off particular parts from further exploitation to prevent damage to various plants where vital machine elements are installed. High accuracy and reliability of measurements of all relevant measuring sizes are ensured, on the basis of which the technical correctness of the plant can be qualitatively determined. Multi-channel, modular, software-hardware monitoring, control and protection systems allow the optimization of plant operating parameters. The new diagnostic method and monitoring systems can be widely applied in all technical fields: internal combustion engines, hydropower plants, thermal power plants, process plants, and in many other areas. Many systems for monitoring, compatibility of devices and equipment from various manufacturers are supported by hardware and software. Mechanical and electrical sizes measured on different types of plants are very important for further operations of drive systems. In this paper, a special attention is given to the diagnostics of sliding bearings.

The new method of diagnostics of sliding bearings, by measuring the dynamic trajectory of the sleeves in the bearing, does not depend on the plant type, and can be applied to all types of internal combustion engines and various other plants. The author of this very important project proposes that this new method and its accompanying monitoring systems be installed in drive systems. In the past, many important systems did not have similar monitoring systems in place, and many problems arose. For system installations used since the 1950s, simpler systems for monitoring basic functions were designed, which was not enough,

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especially in recent times, as the development of new techniques has progressed.

New plant monitoring systems have alarm functions. The alarm can be visual or accoustic or in the form of SMS notifications. The field of technical diagnostics is a very interesting field for many scientists (Braut et al, 2008), (Licen & Zuber, 2003). This area is under-researched, as is the area of medical diagnostics.

References

Braut, S., Zigulic, R., & Butkovic, M. 2008. Numerical and experimantal analysis of a shaft bow influence on a rotor to stator contact dynamics. Strojniski vestnik-Journal of Mechanical Engineering, 54(10), pp.693-706 [online]. Available at: https://www.sv-jme.eu/article/numerical-and-experimantal-analysis-of-a-shaft-bow-influence-on-a-rotor-to-stator-contact-dynamics/ [Accessed: 7 April 2020].

Cernej, A., & et al. 1986. Radni uslovi podmazivanja lezaja dizel motora. Goriva i maziva, 26(5-6), pp.229-236 (in Serbian).

-Easy-Laser. 2020. Laser measurement & alignment systems. Technical documentation. Damalini, Sweden: Easy-Laser.

Lang, O.R, & Steinhilper, W. 1978. Gleitlager. Berlin, Heidelberg: SpringerVerlag. Available at: https://doi.org/10.1007/978-3-642-81225-5.

Licen, H., & Zuber, N. 2003. Vibrodiagnostics as an element for quality and reliabilty assurance: SPIDER 8 - universal measuring device, concept and application. In: Quality 2003: 3rd Research/expert conference with international participation, Zenica, Bosnia and Herzegovina, pp.299-306. November 13-14 [online]. Available at:

http://www.quality.unze.ba/zbornici/QUALITY%202003/045-R-037.pdf (in

Serbian) [Accessed: 7 April 2020].

Zegarac, N. 1989. Dijagnostika kliznih lezajeva u dizel motoru (in Serbian). Ph.D. thesis. Zagreb: University of Zagreb, Faculty of Mechanical Engineering and Naval Architecture.

Zegarac, N. 1993. Postupak dijagnostike lezajeva merenjem dinamickih putanja glavnih rukavaca, kolenastog vratila. Serbian Patent number P-640/93.

Zegarac, N. 1994. Analysis of the wear of vital parts of turbocharged engines of high specific power. Scientific Technical Review, 44(6), pp.14-24.

Zegarac, N. 2016. Experience in the development of the invention from the aspect of scientific verification and market creation. Vojnotehnicki glasnik/Military Technical Courier, 64(1), pp.92-109 Available at: https://doi.org/10.5937/vojtehg64-9315 (in Serbian).

Zegarac, N. 2019. The Application of Modern Diagnostic Systems on Machine and Electric Power Plants. Tehnika, 74(1), pp.73-87. Available at: https://doi.org/10.5937/tehnika1901073Z.

Вывод: Новый метод диагностики и системы мониторинга может иметь широкое применение во всех технических областях, таких как: двигатели внутреннего сгорания, гидроэлектростанции, тепловые электростанции,

технологические установки и пр. Аппаратное и программное обеспечение поддерживает множество систем мониторинга, контроля совместимости устройств и оборудования от разных производителей. Проведена верификация расчета

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представлении важности применения нового метода диагностики и системы мониторинга, возможности надежного определения времени и точного места возникновения проблемы, связанной с износом подшипников скольжения при дальнейшей эксплуатации установки, а также качественного прогноза при продолжении работы системы с течением времени, предусмотрение причин отказа и способов их устранения, а также планировки регулярного технического обслуживания системы. Метод: Новый метод решает проблему диагностики подшипников скольжения путем измерения траекторий перемещения (орбита) вала в подшипнике скольжения. "§ Современные методы технической диагностики, основанные на измерении динамических параметров и электрических величин и их анализе позволяют поддерживать измерительные системы и датчики от различных производителей.

Результаты: Измеряя траектории передвижения (орбиты) валов в подшипниках скольжения, определяется значения, которые характеризуют: нормальное состояние, начальную величину зазора, его дальнейшее увеличение, величину зазора подшипника, когда параметры приближаются предельно допустимой величине зазора и определяется величина зазора, в случае если дальнейшая эксплуатация установки может привести к отказу системы.

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анализе, омогуПава]у подршку мерних система и сензора за мереше разних произво^ача. Резултати: Мерешем динамичких путаша (тра]ектори]е) рукаваца у клизним лежа]има утвр^у]у се величине ко}е карактеришу: нормално сташе, почетну величину зазора, шегово дале повеЯаваше, величине зазора лежа}а, када су параметри сташа близу горше границе дозволеног зазора, и утвр^иваше величине зазора када дала експлоатаци}а постро]еша може проузроковати хавари}у система.

Заклучак: Нова ди}агностичка метода и мониторниг система могу се широко применити у свим техничким областима: моторима са унутрашшим сагоревашем, хидроелектранама, термоелектранама, процесним постро]ешима и многим другим областима. Хардверски и софтверски подржава}у се многи системи за надгледаше и проверава компатибилност уре^а]а и опреме разних произво^ача. Извршена }е верификаци]а прорачуна

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Paper accepted for publishing on / Дата окончательного согласования работы / Датум коначног прихвата^а чланка за об]ав^ива^е: 20.05.2020.

© 2020 The Author. Published by Vojnotehnicki glasnik / Military Technical Courier (www.vtg.mod.gov.rs, втг.мо.упр.срб). This article is an open access article distributed under the terms and conditions of the Creative Commons Attribution license ф

(http://creativecommons.org/licenses/by/3.0/rs/). °

© 2020 Автор. Опубликовано в «Военно-технический вестник / Vojnotehnicki glasnik / Military Technical Courier» (www.vtg.mod.gov.rs, втг.мо.упр.срб). Данная статья в открытом доступе и распространяется в соответствии с лицензией «Creative Commons» (http://creativecommons.org/licenses/by/3.0/rs/).

© 2020 Аутор. Обjавио Воjнотехнички гласник / Vojnotehnicki glasnik / Military Technical Courier (www.vtg.mod.gov.rs, втг.мо.упр.срб). Ово jе чланак отвореног приступа и дистрибуира се у £

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