RU2654306C1 - Method of controlling technical condition of the car - Google Patents

Abstract

FIELD: defectoscopy.

SUBSTANCE: invention relates to non-destructive testing of elastic solids by acoustic methods, namely to methods of monitoring the technical state of machines (power mechanical equipment), and can be used for diagnostics of mainly all types of rotating power-mechanical equipment, including gas-pumping units, turbo-aggregates, pumps, compressors, fans, transmissions driven by an electric motor, an internal combustion engine, etc. In the method of monitoring the technical condition of a machine, a machine operating under load measures the parameters of natural oscillations of its structural elements, judging by the presence of natural oscillations about the appearance of defects. In this case, the parameters of polyharmonic oscillations of the structural elements of the machine are measured, on the constructive data of the product and its kinematic scheme produce a mathematical model of rotating equipment through a computer program, calculate the natural frequencies of the torsional oscillations of the shaft line and the parameters of their description (energy forms, kinetic and potential energy spectra). Then periodically compute the calculated parameters of natural frequencies with experimental information (vibration spectra) based on the parameters of polyharmonic oscillations and failure statistics, by the presence of natural oscillation frequencies in the experimental vibration spectra, the fact of the appearance of defects in the structure is fixed. By the combination of the experimental levels of the spectral components of vibration and the calculated levels of energy capacity of parts in the energy forms and spectra of kinetic and potential energies of the machine receive the result of the object's control as a kind of its technical state with indication of the location, type and causes of the defects.

EFFECT: establishment the fact of the presence of defects and the determination of their location with an accuracy to the next communication node in structurally complex systems; ensuring early detection of defects, their type and causes of occurrence by simplifying the allocation of the vibration signal due to the presence of a malfunction, determination of the most probable places of occurrence of defects at natural frequencies of oscillations of structural elements of machines related to structurally complex systems.

1 cl, 11 dwg, 3 tbl

Description

The invention relates to non-destructive testing of elastic solids by acoustic methods, and in particular, to methods for monitoring the technical condition of machines (power-mechanical equipment). And it can be used to diagnose mainly all types of rotating energy-mechanical equipment, including gas pumping units, turbine units, pumps, compressors, fans, transmissions driven by an electric motor, internal combustion engine, etc.

Defects of structural parts (couplings, gears, shafts, vanes and others) that carry a technological load are a frequent cause of emergency stops and downtime of rotating power-mechanical equipment.

At the initial stages of development, these defects lead to the appearance of vibration-shock processes due to mechanical reasons associated with the parameters of torsional vibrations. Then, the structural parts are destroyed with the loss of operability of the energy-mechanical equipment. For example, the occurrence of a defect in the engagement (grazing) of the teeth of the gear transmission (clutch) leads to the excitation of angular vibration of the elastic system of the shaft of the gas pumping unit (GPU). First of all, shafting parts that are sensitive to the parameters of near-resonant torsional vibrations are destroyed. The GPU’s performance is impaired and the unit is accidentally stopped by standard emergency protection (for example, by significantly reducing the oil pressure in the GPU oil supply system due to the destruction of the hydraulic pump parts) or the emergency and warning alarm parameters of the GPU standard automatic control systems (ACS) (by linear vibration, pressure , temperature) are within acceptable limits and damage to parts is detected during scheduled repairs of the gas compressor unit. These circumstances make urgent the problem of diagnosing incipient defects in parts of power-mechanical equipment by the parameters of torsional vibrations.

A known method for detecting defects in an elastic construction material (RU2190207C2, published September 27, 2002). According to this method, improving the reliability of detection of a defect is achieved due to the fact that excite the oscillations of the reference and the investigated design and select several forms of vibration. For the selected vibration modes, the equivalent masses of structures corresponding to the observation point, which is chosen near the nodal line or nodal point, are determined by the calculation-experimental method. The occurrence of a defect is judged by the difference in the values of the equivalent masses for the reference and investigated designs. For the chosen modes of vibration, several equivalent masses of the reference and studied structures are additionally determined at the observation points, which are chosen near the most probable places of the defect to occur, and the place of its occurrence is determined by the largest difference in the equivalent masses of the reference and studied structures from among the additionally determined ones. The most probable places of occurrence of a defect are determined, for example, using the ANSYS application package by numerical calculations of stresses at various points of the structure that arise during its operation (ANSYS [electronic resource]. - URL: http://www.ansys.com (accessed date 10.05 .2017); Basov K.A. ANSYS for designers. - M.: DMK Press, 2009. - P. 248). The most probable points of occurrence of the defect are those points at which the stresses are maximum.

The disadvantage of the analogue method is that it is necessary to use labor-intensive operations:

- for the selected vibration modes, the equivalent mass of the structures is determined by the calculation-experimental method;

- in order to determine the most probable places of occurrence of the defect, a powerful finite-element application package ANSYS is used to simulate forced oscillations.

Also, in the analogue method, the technology for selecting the desired vibration modes of the structures for their subsequent analysis has not been disclosed.

In other well-known methods for monitoring the technical condition of energy-mechanical equipment (Non-Destructive Testing: Ref .: 8 vol. / Under the general editorship of V.V. Klyuyev. - M.: Mechanical Engineering, 2003–2005; A. Reshetov Non-Destructive Testing and technical diagnostics of energy facilities: textbook / A.A. Reshetov, A.K. Arakelyan; edited by prof.A.K. Arakelyan, Cheboksary: Publishing House of Chuvash University, 2010. - 470 p .; Sokolova A.G., Balitsky F. Ya., Dolaberidze G.V. et al. Vibromonitoring of the state of a DG-90 gas turbine engine according to multivariate discriminant analysis // estnik technological development National Technology Group -.. 2011. - № 2 (42) - pp. 47-56) requires a thorough application of the methods of correlation and regression analysis due to the fact that the vibration signal parameters dependent on a significant number of factors.

The closest analogue adopted for the prototype is a method for monitoring the technical condition of an electric machine (RU2304837C2, published on 08.20.2007), where the parameters of natural vibrations of its structural elements and the presence of natural vibrations at frequencies are measured on an electric machine operating under load, not having the property of multiplicity with respect to the frequencies of the main coercive forces, they judge the appearance of defects and their form. To increase the reliability of diagnosis and recognition of defects, measurements are carried out in various load conditions.

The disadvantages of the prototype are as follows:

- significant economic costs for the search for defects in structurally complex systems;

- the lack of the ability to determine the location of defects up to the next communication node in structurally complex systems;

- the lack of early detection of defects by simplifying the allocation of the vibration signal due to the presence of a malfunction;

- the impossibility of determining the most likely places of occurrence of a defect at the natural frequencies of vibrations of the structural elements of machines related to structurally complex systems.

The technical result of the application of the invention are as follows:

- reduction of economic costs for the search for defects in structurally complex systems;

- establishing the fact of the presence of defects, determining their location
up to the next communication node in structurally complex systems;

- ensuring early detection of defects, their type and causes by simplifying the allocation of the vibration signal due to the presence of a malfunction;

- determination of the most probable places of occurrence of defects at natural frequencies of vibration of structural elements of machines related to structurally complex systems.

The technical result of the invention is achieved by the fact that in the method of monitoring the technical condition of the machine, on a machine running under load, the parameters of the natural vibrations of its structural elements are measured, the presence of natural vibrations is judged on the occurrence of defects, characterized in that the parameters of polyharmonic vibrations of the structural elements of the machine are measured , according to the structural data of the product and its kinematic scheme, a mathematical model of the rotating equipment is constructed by means of a computer program, calculate the natural frequencies of torsional vibrations of the shaft line and the parameters of their description (energy forms, spectra of kinetic and potential energies), periodically compare the calculated parameters of natural frequencies with experimental information (vibration spectra) based on the parameters of polyharmonic oscillations and failure statistics, according to the presence of natural vibration frequencies in experimental vibrational spectra record the fact of the appearance of defects in the structure, according to the set of exp The experimental levels of the spectral components of vibration and the calculated energy levels of the parts in the energy forms and spectra of the kinetic and potential energies of the machine receive the result of monitoring the object as a type of its technical condition with an indication of the location, type and causes of defects (Reshetov A.A., Zakharov N.A. . Software and hardware for improving the efficiency of vibration-diagnostic control of energy-mechanical equipment: Certificate of state registration of computer programs dated 01.09.2014 No. 2014610101. - M., 2014 .; Reshetov A.A., Zakharov N.A., Artemyev I.T. Technology for improving the efficiency of diagnosing the technical condition of gas pumping units // Bulletin of Chuvash. un-that. - Cheboksary: Publishing house of Chuvash. University, 2015. No. 3. - S. 186–192).

This method is characterized in that the calculated parameters are used to describe the natural frequencies of torsional vibrations of the machine shaft (energy forms, spectra of kinetic and potential energies), the instrument measurements of the parameters of polyharmonic oscillations are repeated every 600 hours of machine time, and the measurements are carried out in various modes (start, load, coasting).

To achieve the technical result of the present invention, new parameters have been introduced and used to describe the diagnostic signs of incipient damage (natural frequencies) —parameters of energy forms and energy spectra of natural vibrations of structurally complex systems with a significant number of degrees of freedom, which allow predicting the onset of defects (ranking the risks of damage to parts of power-mechanical equipment in resonance and near-resonance modes).

=54 сочленения ротора электродвигателя СТД-12500 и вала-колеса (16)

=88 мультипликатора;
17 – вал-шестерня

=51 мультипликатора; 18 – рабочее (насосное) колесо ГДМ

=49; 19 – рабочее (турбинное) колесо ГДМ

=48; 20 – зубчатая муфта

=56 сочленения вторичного вала ГДМ и ротора ЦБН; 21 – рабочее колесо 1-й ступени ЦБН Н-235-21-1

=14; 22 – рабочее колесо 2-й ступени ЦБН Н-235-21-1

=14; 23 – думмис; 24 и 25 – цилиндрические шестерни

=78 и

=79; 26 и 27 – конические шестерни

=33 и

=33; 28 – малое зубчатое колесо

=14; 29 – планетарная шестерня

=11 (3 шт.); 30 – рабочее колесо

=11 лопастного насоса.In FIG. 1 - shows the kinematic diagram of a gas-pumping unit EGPA-12500 with a hydrodynamic coupling (GDM) of type R18K480MINV from Voith Turbo, Germany, where the following notation is used: 1 ÷ 13 - plain bearings;
14 - rotor of an electric motor STD-12500; 15 - gear clutch

= 54 joints of the rotor of the electric motor STD-12500 and shaft-wheels (16)

= 88 multiplier;
17 - gear shaft

= 51 multiplier; 18 - working (pumping) wheel

= 49; 19 - working (turbine) wheel

= 48; 20 - gear clutch

= 56 articulation of the secondary shaft of the pulp mill and the rotor of the pulp and paper mill; 21 - impeller of the 1st stage of the pulp and paper mill Н-235-21-1

= 14; 22 - impeller of the 2nd stage of the Central Bank N-235-21-1

= 14; 23 - dummis; 24 and 25 - spur gears

= 78 and

= 79; 26 and 27 - bevel gears

= 33 and

= 33; 28 - small gear

= 14; 29 - planetary gear

= 11 (3 pcs.); 30 - impeller

= 11 vane pump.

=88 мультипликатора; 7 – вал-шестерня

=51 мультипликатора; 8 – рабочее (насосное) колесо ГДМ

=49; 9 и 10 – цилиндрические шестерни

=78 и

=79; 11 и 12 – конические шестерни

=33 и

=33; 13, 14 и 15 – инерционные массы вала насоса; 16 – шестеренчатый насос жидкой смазки; 17 – рабочее колесо

=11 лопастного насоса рабочего масла; М₁ ÷ М₅, М₈, М₁₆, М₁₇ – моменты, воздействующие на ротор ГЭД, зубчатую муфту, рабочее (насосное) колесо ГДМ, насосы ГДМ (моменты, возбуждающие крутильные колебания); М_с1, М_с8, М_с16, М_с17 – моменты, демпфирующие крутильные колебания (моменты сопротивления).  In FIG. 2 - shows the model of torsional vibrations of the EGPA-12500 shafting with a hydraulic module, where the following designations are accepted: J_m and C_k - axial moments of inertia of concentrated m-x masses (parts) and rigidity during torsion of k-th shaft sections; GED - the main electric motor; ЗМ - gear coupling; MP - multiplier; 6 - shaft-wheel

= 88 multiplier; 7 - gear shaft

= 51 multiplier; 8 - working (pumping) wheel

= 49; 9 and 10 - spur gears

= 78 and

= 79; 11 and 12 - bevel gears

= 33 and

= 33; 13, 14 and 15 - inertial masses of the pump shaft; 16 - gear pump fluid lubrication; 17 - the impeller

= 11 vane pump of working oil; M_one ÷ M₅, M₈, M₁₆, M₁₇- moments acting on the rotor of the EDL, gear coupling, working (pumping) wheel of the hydraulic machine tools, pumps of the hydraulic machine tools (moments exciting torsional vibrations); M_s1, M_c8, M_s16, M_s17- moments damping torsional vibrations (moments of resistance).

In FIG. 3 - shows the energy forms of torsional vibrations of the GPA shafting (forms of relative kinetic energies of inertial parts), where the following notation is accepted: T_m2and T₂Is the maximum value of the kinetic energy of the mth inertial mass and the entire system with natural vibrations with a frequency f₂= 258.02 Hz; q_m2Is the generalized coordinate of the mth shaft line parts on the 2nd form of natural vibrations at f₂= 258.02 Hz.

In FIG. 4 - shows the energy forms of torsional vibrations of the GPA shafting (forms of relative potential energies of elastic parts), where the following notation is used: W _k2 and W ₂ are the maximum values of the potential energies of the k-th section and the entire system with natural vibrations with a frequency f ₂ = 258, 02 Hz; k ₂ is the number of the kth elastic section of the shaft line on the 2nd form of natural vibrations at f ₂ = 258.02 Hz.

In FIG. 5 - shows the energy forms of torsional vibrations of the GPA shafting (forms of relative kinetic energies of inertial parts), where the following notation is accepted: T_m4and T_fourIs the maximum value of the kinetic energy of the mth inertial mass and the entire system with natural vibrations with a frequency f_four= 392.53 Hz; q_m4Is the generalized coordinate of the mth shaft line parts on the 4th form of natural vibrations at f_four= 392.53 Hz.

In FIG. 6 - shows the energy forms of torsional vibrations of the GPA shafting (forms of the relative potential energies of elastic parts), where the following notation is used: W _k4 and W ₄ are the maximum values of the potential energies of the kth section and the entire system with natural vibrations with a frequency f ₄ = 392, 53 Hz; k ₄ is the number of the k-th elastic section of the shaft line on the 4th form of natural vibrations at f ₄ = 392.53 Hz.

In FIG. 7 - shows the energy forms of torsional vibrations of the GPA shafting (forms of relative kinetic energies of inertial parts), where the following notation is accepted: T_m5and T₅Is the maximum value of the kinetic energy of the mth inertial mass and the entire system with natural vibrations with a frequency f₅= 611.75 Hz; q_m5Is the generalized coordinate of the mth shaft line parts on the 5th form of natural vibrations at f₅= 611.75 Hz.

In FIG. 8 - shows the energy forms of torsional vibrations of the GPA shafting (forms of relative potential energies of elastic parts), where the following notation is used: W _k5 and W ₅ are the maximum values of the potential energies of the k-th section and the entire system with natural vibrations with a frequency f ₅ = 392, 53 Hz; k ₅ is the number of the kth elastic section of the shaft line on the 5th form of natural vibrations at f ₅ = 611.75 Hz.

In FIG. 9 - shows the direct vibration spectrum of the EGPA fluid coupling bearing housing without defect of engagement (grazing) of the teeth of the gear clutch z ₁ = 54 GPA (station No. 1) (thrust bearing of the drive shaft of the hydraulic drive, control point No. 3, axial direction), where the following designations are adopted: V is the amplitude of the RMS vibration velocity of the GPU body in m / s; f is the frequency in Hz.

= 1,72 мм/с – амплитуда СКЗ виброскорости корпуса подшипника ГПА в осевом направлении, мм/с, в диапазоне частот 10 ÷ 1 500 Гц. In FIG. 10 - shows the direct vibration spectrum of the EGPA fluid coupling bearing housing with a defective engagement (grazing) of the teeth of the gear clutch z ₁ = 54 GPA (station No. 7) (thrust bearing of the drive shaft of the hydraulic drive, control point No. 3, axial direction), where the following designations are adopted: V is the amplitude of the RMS vibration velocity of the GPU body in m / s; f is the frequency in Hz;

= 1.72 mm / s - the amplitude of the RMS vibration velocity of the HPA bearing housing in the axial direction, mm / s, in the frequency range 10 ÷ 1,500 Hz.

= 3,03 мм/с – амплитуда СКЗ виброскорости корпуса подшипника ГПА в осевом направлении, мм/с, в диапазоне частот 10 ÷ 1 500 Гц.In FIG. 11 - shows the direct vibration spectra of the EGPA fluid coupling bearing housing with a defect of engagement (grazing) of the teeth of the gear coupling z ₁ = 54 GPA (station No. 7) (thrust bearing of the drive shaft of the hydraulic drive, control point No. 3, axial direction), where the following notation is used: V is the amplitude of the RMS vibration velocity of the GPU body in m / s; f is the frequency in Hz;

= 3.03 mm / s - the amplitude of the RMS vibration velocity of the HPA bearing housing in the axial direction, mm / s, in the frequency range 10 ÷ 1,500 Hz.

The inventive method is carried out, for example, on a gas turbine of the EGPA-12500 type with a centrifugal supercharger TsBN N-235-21-1 and a hydrodynamic coupling (GDM) of the type R18K480MINV from Voith Turbo (Germany) with the following technical characteristics (Fig. 1):

- rated power of the STD-12500 electric motor - 12.5 MW;

- the nominal rotational speed of the rotor of the electric motor is 3,000 rpm.

Also, the parameters of the excitation of forced torsional-bending-axial vibrations of the EGPA-12500 shaft shaft with the GDM, corresponding to the frequencies of the engagement (passage) of the first tooth (blade) harmonics of the exciting moments (Table 1), were calculated.

Table 1

Excitation parameters of forced torsional-bending-axial

of oscillations of the EGPA-12500 shafting with gas turbine

EGPA Details Gear
relation to the speed of rotation of the electric drive Frequencies
rotation
parts, Hz Engagement frequencies
(passage) first
gear (scapular)
exciting harmonics
moments, Hz

=88
мультипликатораshaft-wheel

= 88
multiplier one fifty 4400 gear shaft

= 51
multiplier 88/51 ≈ 1.725 ~ 86,274 4400

Continuation of table 1

=49working (pumping)
paper wheel

= 49 88/51 ≈ 1.725 ~ 86,274 ~ 4227.45 spur wheel

= 78 one fifty 3900 cylindrical
gear

= 79 78/79 ≈ 0.987 ~ 49,367 3900 bevel wheel

= 33 78/79 ≈ 0.987 ~ 49,367 ~ 1629.11 bevel gear

= 33 78/79 ≈ 0.987 ~ 49,367 ~ 1629.11 cylindrical small
gear

= 14 78/79 ≈ 0.987 ~ 49,367 ~ 691.14 cylindrical
planet gear

= 11 (3 pcs.) 78 ∙ 14 / (79 ∙ 11) ≈ 1,256 ~ 62,831 ~ 691.14 Working wheel

= 11
vane pump 78/79 ≈ 0.987 ~ 49,367 ~ 543.04

To determine the a priori basis of the technical diagnostics system (energy forms and energy spectra) of the gas compressor unit, a mathematical construction of the mathematical model of the diagnostic object was performed based on the calculation of its dynamic parameters according to technical data and drawing design documentation (Fig. 2, Table 2) (A. Reshetov ., Zakharov N.A. Software and hardware for increasing the efficiency of the vibration-diagnostic control of energy-mechanical equipment: Certificate of state registration of the program for Computer from 09.01.2014 № 2014610101. - M., 2014).

table 2

Dynamic characteristics of the model of torsional vibrations of a GPA EGPA-12500 shaft shaft with a gas turbine engine (reduced to the rotational speed of the rotor HED, gear ratio i = 88/51 ≈ 1.725)

№№
inertial masses Nodes, details Axial moment of inertia
J _m , kg * m ² №№
elastic sections Torsional rigidity C _k , N * m / rad

Continuation of table 2

one Rotor STD-12500 275.0 1-2 1.22710 ⁷ 2 Tooth. sleeve + rotor part 2,225 2-3 1 · 10 ¹⁰ 3 Tooth. coupling half + prom. shaft 1,871 3-4 2.25110 ⁸ four Part prom. shaft + tooth. coupling half 1,871 4-5 1 · 10 ¹⁰ 5 Tooth. sleeve + shaft part 1,047 5-6 2,59510 ⁷ 6 Shaft-wheel + shaft part 27,757 6-7 1 · 10 ¹⁰ 7 Pinion shaft + shaft part 9,632 7-8 5,71810 ⁷ 8 Pump wheel 46,652 6-9 2,76110 ⁷ 9 Cylindrical Tooth. wheel +
shaft part 0.113 9-10 4,98310 ⁶ 10 Cylindrical Tooth. gear + shaft part 0.01418 10-11 8,00610 ⁴ eleven Conical tooth. wheel +
drive shaft 0,00304 11-12 3,05710 ⁶ 12 Conical tooth. gear +
pump shaft part 0,00449 12-13 1.019 · 10 ⁵ 13 Pump shaft part 0,00108 13-14 9.19110 ⁴ fourteen Pump shaft part 0.00125 14-15 7,10010 ⁴ fifteen Pump shaft part 0,00142 15-16 7,000 · 10 ⁴ 16 Gear pump 0,00087 16-17 1,350 · 10 ⁵ 17 Vane pump 0.00140

According to (Reshetov A.A., Zakharov N.A. Software and hardware for increasing the efficiency of vibrodiagnostic control of energy-mechanical equipment: Certificate of state registration of computer programs dated 09.01.2014 No. 2014610101. - M., 2014), the calculation of energy forms and energy spectra of generalized torsional vibrations of the GPA shafting system. The sensitivity of the natural frequencies of torsional vibrations of the GPA shafting to the characteristics of inertial and elastic elements (the sensitivity of the parameters of parts to the generation of defects as the energy rating of the shafting elements) is determined numerically through the energy forms of vibrations (Fig. 3 ÷ 8). The value of each sensitivity coefficient corresponds to the percentage change in the natural frequency of the system when the corresponding characteristic of the element changes by 100%.

At a gas compressor unit operating in the route of gas pipelines according to the dispatch schedule in a load mode close to the nominal one, the frequencies and amplitudes of polyharmonic oscillations in the frequency range 10 ÷ 1,500 Hz are measured. Due to the fact that the torsional vibrations of the shafting of this installation are transmitted to the bearings (the presence of a gear coupling and gears) and cause linear vibration of the machine's support units, then vibration transducers of the VP-3 type are used to measure its parameters in the vertical, horizontal-transverse and axial directions magnetic fasteners mounted on the GPA bearing housing. The vibration sensors are connected to a data collector / vibration analyzer SK-2300 (Orgenergogaz JSC, Moscow) for vibration monitoring of the GPA technical condition and calculation of the SKZ vibration velocity of the bearing housings, mm / s, in the frequency range 10 ÷ 1,500 Hz.

The results of measurements of vibration of the GPU casing without a defect in graphical form are presented in FIG. 9.

The defect of engagement (engagement) of the teeth of the gear coupling z ₁ = 54 GPA, which manifested during its operation in damage and destruction of energy-consuming parts of the hydraulic pumping unit (drive shaft of a mechanical hydraulic pump; teeth of cylindrical wheels of a reduction gear pair of a drive of a hydraulic hydraulic pump), was determined by comparing the experimental information with a theoretical spectral portrait.

The results of measurements of vibration of a GPU casing with a defect in graphical form are presented in FIG. 10.

As can be seen from the measurement results (Fig. 10), together with the harmonics of oscillations at the frequencies of the main driving forces 50 × 0.987; 50 × 1.256; 50 × 1.725 Hz there are harmonics at frequencies ~ 258.02 and ~ 392.53, which do not have the property of multiplicity with respect to frequencies 50 × 0.987; 50 × 1.256; 50 × 1.725 Hz, but corresponding to the frequencies of natural torsional vibrations of the GPA shaft line f ₄ = 392.53 Hz (with the exception of the natural frequency f ₂ = 258.02 Hz ≈ 50 × 1.725 × 3). Based on the detection of harmonics of intrinsic torsional vibrations of the GPA shafting, the fact of the appearance of a vibration-shock defect is recorded (Fig. 3 ÷ 10).

In this way:

- based on a comparison of the calculated parameters (natural frequencies, energy forms of kinetic and potential energies) with experimental information (vibration spectra) on the presence of natural frequencies in the vibration spectrum, the fact of the appearance of a vibration-shock defect in the structure is recorded;

- according to the calculated level of the parameters of the energy intensity of the parts in the energy forms of kinetic and potential energies, they make a conclusion about the location of the defects of the parts according to the structure of the product (defect of gear teeth of the gear clutch z ₁ = 54 GPA).

To clarify the diagnosis, the instrumental measurements of the parameters of the polyharmonic vibrations (vibration) of the GPU body are repeated every 600 hours of machine time, the results of which are presented in FIG. eleven.

As can be seen from the measurement results (Fig. 11), together with the harmonics of the oscillations at the frequencies of the main driving forces 50 × 0.987; 50 × 1.256; 50 × 1.725 Hz there are harmonics at frequencies of ~ 258.02; ~ 392.53; ~ 611.75 Hz, not having the property of multiplicity with respect to frequencies of 50 × 0.987; 50 × 1.256; 50 × 1.725 Hz, but corresponding to the frequencies of natural torsional vibrations of the GPU shaft line f ₄ = 392.53 Hz; f ₅ = 611.75 Hz (except for the natural frequency f ₂ = 258.02 Hz ≈ 50 × 1.725 × 3). Based on the detection of harmonics of intrinsic torsional vibrations of the GPA shafting, the fact of the appearance and development of a vibration-shock defect is recorded (Fig. 3 ÷ 11).

In this way:

- based on a comparison of the calculated parameters (natural frequencies, energy forms of kinetic and potential energies) with experimental information on the presence of natural frequencies in the vibration spectrum, the fact of the appearance and development of a vibration-shock defect in the structure is recorded;

- on the calculated level of the energy intensity parameters of the parts in the energy forms of kinetic and potential energies, make a conclusion about the location of the defects of the parts according to the structure of the product (defect of gear teeth of the gear coupling z ₁ = 54 GPA; drive shaft of a mechanical hydraulic pump; gear wheels of a gear pair of a mechanical pump drive hydraulic systems).

The final technical diagnosis (conclusion on the causes of defects). Natural frequencies of torsional vibrations f ₂ = 258.02 Hz; f ₄ = 392.53 Hz; f ₅ = 611.75 Hz are determined respectively by the energy consumption parameters of the gear coupling parts z ₁ = 54 GPA and the pump insert of the GPU hydrodynamic coupling (drive shaft of a mechanical hydraulic pump; gear wheels of a gear pair of a mechanical hydraulic pump drive) - FIG. 3 ÷ 11:

- are absent in the direct vibration spectra of the bearing housing of the GPU fluid coupling (gear coupling z ₁ = 54 GPA without defect in gear engagement) - FIG. 9;

- are manifested in the direct vibration spectra of the bearing housing of the GPU fluid coupling (the gear coupling z ₁ = 54 GPA has a tooth gear defect) - FIG. 10, 11;

- the cause of the repeated destruction of the nodes of the pumping gearbox insert of the hydraulic drive gear is the excitation of near-resonant torsional-bending-axial vibrations of the elastic system electric motor - gearbox pair of the hydraulic motor gearbox - pump gearing insert of the hydraulic gearbox;

- the source of the excitation of dangerous torsional vibrations of the GPU shaft line is the technological installation defect of the gear teeth (engagement) of the teeth of the gear coupling z ₁ = 54 GPA, causing forced oscillations of the system in a wide frequency range on a number of harmonic components of the working frequencies such as the comb spectrum and natural frequencies;

- type of technical condition of the control object - emergency condition.

Management of the diagnostic object based on the results of the conclusion on the causes of defects. In order to increase the reliability of the GPU, technical measures have been introduced to replace the parts of the gear coupling z ₁ = 54 GPA and the control of the manufacturing process, fault detection and installation of couplings has been strengthened.

Data on the results of the practical implementation of this invention when monitoring the technical condition of power-mechanical equipment are systematized and summarized in table. 3.

Table 3

Data on the results of the practical implementation of the developed method

GPA type GPA problem nodes Natural frequencies of torsional vibrations GPU-intensive parts EGPA-12500
with paperwork The defect of gearing (grazing) of the teeth of the gear clutch z ₁ = 54 GPA.
Destruction of parts
pump insert f ₂ = 258.02 Hz;
f ₄ = 392.53 Hz;
f ₅ = 611.75 Hz Gear Coupling Parts
z ₁ = 54 GPA;
details of pump insert

Continuation of table 3

EGPA-12500
without paperwork

=56 (разрушение 40% зуба зубчатой втулки мультипликатора по его длине, скол корневой части
2-х зубьев)Gear coupling of a joint of a gear shaft of a multiplier and a rotor of TsBN

= 56 (destruction of 40% of the tooth of the gear sleeve of the multiplier along its length, cleaved root part
2 teeth) f ₃ = 338.75 Hz Gear Coupling Parts

= 56 EGPA2-12.5-76 / 1.5 Gear coupling of articulation of ED and shaft-wheel of the animator:
- heating and destruction of rubber elements;
- seizing (welding, setting) of teeth
gear clutch f ₃ ≈ (160.3 ÷ 169.5) Hz due to non-linear parameters of the gear coupling with rubber elements Gear Coupling Parts
with rubber elements

Claims (2)

1. The method of monitoring the technical condition of the machine, in which the parameters of the natural vibrations of its structural elements are measured on a machine operating under load, the presence of defects is judged by the appearance of defects, characterized in that the parameters of polyharmonic vibrations of the structural elements of the machine are measured, according to the structural data of the product and its kinematic diagram, a mathematical model of the rotating equipment is constructed using a computer program, the calculation of the eigenfrequencies cr of useful shafting vibrations and parameters for their description (energy forms, spectra of kinetic and potential energies), periodically compare the calculated parameters of natural frequencies with experimental information (vibration spectra) based on the parameters of polyharmonic vibrations and failure statistics, according to the presence of natural frequencies in the experimental vibration spectra, the fact of the appearance of defects in the structure, according to the set of experimental levels of the spectral components of vibration and calculated x levels of energy intensity of parts in energy forms and spectra of kinetic and potential energies of the machine receive the result of monitoring the object as a type of its technical condition with an indication of the location, type and causes of defects. 2. The method according to claim 1, characterized in that the calculated parameters are used to describe the natural frequencies of torsional vibrations of the shaft shaft of the machine (energy forms, spectra of kinetic and potential energies), instrument measurements of the parameters of polyharmonic vibrations are repeated every 600 hours of machine time, and measurements are carried out in various modes (start, load, coast).

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