RU2248081C1 - Method for condition inspection of electrical machine rotor shaft - Google Patents

Abstract

FIELD: electrical engineering, electromechanical engineering technology, in-service diagnostics, serviceability check of turbogenerator rotor shafts.

SUBSTANCE: proposed method for diagnosing condition of electrical machine rotor shaft depending on comparison of curves showing vibratory displacement S of rotor shaft (bearing) as function of rotor speed n for flaw-free rotor shaft [S = f(n)] with similar curves of rotor under inspection for speeds from starting moment of electrical machine to full critical speed of first order and close to this speed gained by rotor. Transversal crack formed in rotor shaft in the process is detected by greater slope of (S = f(n) curve, other conditions being equal. Diagnosing by using this method can be conducted both in the course of machine starting and stopping.

EFFECT: ability of preventing turbogenerator rotor shaft destruction in the course of operation.

6 cl, 3 dwg

Description

The invention relates to the technology of electrical engineering, the diagnosis of electrical machines during their operation, in particular to the integrity control of the rotor shafts of turbine generators.

A known method for diagnosing the condition of the rotor shafts of electric machines, for example, synchronous compensators [1], in which the technical condition of the rotor is evaluated based on the results of vibration studies of a working electric machine at a nominal speed, while varying its electromagnetic load by comparing vibration displacements (s) and vibration phase (φ) thrust bearings with the corresponding indicators obtained empirically in the previous time. This method, allowing to evaluate the general vibrational state of the rotor and to identify deviations, mainly related to its electromagnetic and thermal state, does not make it possible to timely identify diagnostic parameters characterizing violation of the integrity of the rotor shaft from these vibration measurements, especially for flexible rotors of turbogenerators with nominal rotation speed (n) is close to the 2nd critical.

The closest analogue to the prototype of the invention is a method for vibration diagnostics of the state of the rotor of an electric machine based on the analysis of the vibration curve at a variable speed s = f (n) [2]. The basis of this method for diagnosing the state of a turbogenerator rotor is the provision that the vibration displacement of the rotor shaft is proportional to the perturbing force (F), and the latter, in the presence of an unbalanced rotor of an electric machine, is determined from the known relation: F = m · r · ω ² , where m is the unbalanced mass , r is the radius of application of the center of gravity of the unbalanced mass, ω is the angular velocity of rotation of the rotor.

Taking into account that the vibration displacement of the rotor shaft under the influence of imbalance varies proportionally to the square of the angular velocity (rotational speed) of the rotor in the considered method, any deviation from the parabolic dependence s = f (n) is diagnosed as a malfunction excluding rotor unbalance. The disadvantage of this diagnostic method is that it does not detail the signs characterizing the defects of the rotor shaft of an electric machine, which appear, for example, when a transverse crack is formed, the influence of which on the change in vibration displacement of the rotor shaft during an increase in the rotational speed is also proportional to the square of the angular velocity of the rotor.

The technical result to which this invention is directed is to diagnose the state of the integrity of the shafts of the rotors of turbogenerators during operation to prevent their destruction.

Figure 1 schematically shows the rotor of a turbogenerator with a barrel 1 in the central part, end parts 2, retaining members 3, supported by bearings 4, and a graphical representation of its dynamic deflection when measuring vibrations in the vertical and transverse directions under the influence of unbalance forces and the influence of the rotor mass . In this case, the following designations of the parameters of the oscillations of the shaft:

S ₁ - dynamic deflection of the rotor of the turbogenerator with the first form of imbalance;

S $\genfrac{}{}{0ex}{}{\text{I}}{\text{2}}$ , S $\genfrac{}{}{0ex}{}{\text{II}}{\text{2}}$ , S $\genfrac{}{}{0ex}{}{\text{III}}{\text{2}}$ , S $\genfrac{}{}{0ex}{}{\text{IV}}{\text{2}}$ - dynamic deflection of the rotor under the influence of the weight of the rotor, respectively, with the position I, II, III and IV of the cross section (AA) of the rotor shaft with a crack (figure 2);

L ₁ and L ₂ - the distance from the transverse axis of the support bearing, respectively, to the end part and the middle of the barrel of the rotor of the turbogenerator;

Figure 2 shows the position of the cross-section AA of the rotor with a crack in 4 positions fixed during rotation, with: maximum crack opening (I), no crack opening (II, IV), partial crushing of the peripheral zone of the crack (III )

υ is the displacement of the center of mass of the rotor shaft in the vertical direction from the axis of rotation of the shaft.

Figure 3 shows the dependence of the vibration displacement (s) and the phase of the vibration displacement (φ) on the rotational speed of the rotor; s = f (n) and φ = f (n) and the following notation is accepted:

s, φ, n - respectively, the vibration displacement, phase and rotor speed;

s ₁ , s ₂ , is the vibration displacement of the rotor shaft (bearings) of the turbogenerator, respectively, for the imbalance of the rotor of the first kind (unbalance) and from the influence of the mass of the rotor:

s ₃ - the total maximum vibration displacement in the vertical direction when the phases of vibration displacement s ₁ and s ₂ coincide;

φ $\genfrac{}{}{0ex}{}{\text{I}}{\text{2}}$ , φ $\genfrac{}{}{0ex}{}{\text{II}}{\text{2}}$ , φ $\genfrac{}{}{0ex}{}{\text{III}}{\text{2}}$ - the dependence of the phases of the vibration displacement of the rotor shaft (bearings) of the turbogenerator, respectively, as the crack in the rotor shaft opens, degrees.

n $\genfrac{}{}{0ex}{}{\text{one}}{\text{one}}$ - 1st critical rotor speed of the first kind.

n $\genfrac{}{}{0ex}{}{\text{one}}{\text{2}}$ - also of the second kind.

In the practice of manufacturing and operating turbine generators in diagnosing the technical condition of rotors by methods of vibration diagnostics, measurements of vibration displacement (vibration velocity) and phase of vibration displacement are made using commonly available instruments, the stationary types of which are equipped with most large turbogenerators. In this case, the vibration of the rotor shaft is measured in the vertical and transverse directions, and that of the thrust bearings also in the axial direction, which are the result, for example, of an imbalance of the rotors causing a dynamic deflection S ₁ shown in FIG. 1. It should be noted that the maximum deflection of the shaft and, accordingly, the magnitude of the vibration displacement of the bearings in the transverse and vertical directions vary in proportion to the magnitude of the imbalance and the square of the rotational speed, differing from each other only due to the different stiffness (compliance) of the bearings and the oil layer between the bearing and the rotor shaft, respectively, in the vertical and transverse directions.

In Fig. 1, the deflections (S ₁ ) for a fixed rotor speed are shown in the transverse and vertical directions the same for the case of equal-circular stiffness of the bearing.

Diagnostics is also practiced with measuring the vibration displacement of the rotor shaft and bearings of the turbogenerator as a function of the rotor speed s = f (n) with the determination of the region of occurrence and the values of vibration displacement of the first and second critical rotation frequencies of the first kind s ₁ (Figure 3). However, the diagnosis of a violation of the integrity of the rotor shaft and the fact of the appearance of a crack is determined only by the increase in vibration, including spasmodic bearings, as a rule, during operation of turbine generators at a nominal speed of rotation along with the manifestation of any other less significant defects of the rotor, leading to a deterioration in its vibration condition. It should be noted that for most modern turbogenerators with flexible rotors, the appearance of a crack in the end parts does not actually lead to a noticeable increase in the vibration of the bearings, since a crack causes a deflection of the shaft of the first form. which, as a rule, does not excite oscillations of bearings since at a working frequency of rotation close to the second critical, they work in the zone of insensitivity to oscillations of the rotor in the first form.

The essence of the proposed method for diagnosing the state of the rotor shaft is to fix vibration signs characteristic of the development of cracks in the rotor shaft in magnitude, phase and direction of vibration displacement, due to the dependence of vibration mixing and the phase of vibration of the rotor shaft and bearings on the rotor speed, as well as the boundaries of the rotor speed, within which the most significant change in rotor vibration occurs, associated with the appearance of a transverse crack.

Two types of oscillatory processes of rotors are known that occur during the operation of electric machines and are caused by the action of forces, the value of which increases in proportion to the change in the square of the rotational speed of the rotor shaft:

- the oscillatory process of the rotor of the first kind is spatial, associated with the rotor imbalance, which during rotation causes its dynamic deflection, which, under certain assumptions, has a sinusoidal shape in the form of a half-wave of a sinusoid (1st form of unbalance), a full period of a sinusoid (2nd form of imbalance) etc.

- the oscillatory process of the rotor of the second kind - occurs in a vertical plane along the longitudinal axis of the rotor and is associated with the influence of the mass of the rotor on the shaft line in the presence of asymmetry of any cross section of the rotor, and the prevailing pitch of the deflection of the rotor is a half-period of a sinusoid.

If we take into account that during the oscillatory process of the first kind, only constant bending stresses appear in the rotor shaft, and during the oscillatory process of the second roll, the most dangerous alternating bending stresses cause transverse cracks, as well as the fact that the formation of cracks in the shaft leads to its asymmetry cross-section, it becomes reasonable to use in the diagnosis of the state of the rotor shaft of an electric machine signs characteristic of the oscillatory process of the second kind, the proposed in the invention, distinguishing them from signs of vibration diagnostics, characteristic of other types of rotor defects, affecting the vibrational state of the machine (mechanical, thermal, and other types of imbalance, defects of coupling halves, bearings, etc.)

The oscillatory process of the rotor of the turbogenerator of the 2nd kind is characterized by some features.

The mass distribution of the rotor along its longitudinal axis is uneven. Most of the mass is concentrated in the middle part (barrel) of the rotor, especially for turbogenerator rotors, the retaining nodes of which 2 (Figure 1) are cantilevered, and the ratio L ₁ / L ₂ (Figure 1) is near 0.5. For such rotors, the influence of the rotor mass as the factor causing the second-order oscillatory process with the highest value of the static deflection of the shaft, causing increased alternating bending stresses and creating the prerequisites for the appearance of cracks in the shaft and the appearance of the dynamic components of the shaft deflection in the second-kind oscillatory process with by moving the center of mass of the rotor in a vertical plane by υ (Figure 2)

It is known that the formation of cracks in the rotor shaft, as a rule, is a long process, which can last several months, and its growth rate depends both on the design features of the rotor and on operating conditions, for example, the frequency of starts and stops. Examination of the rotor shafts at the fracture site shows at least the presence of a fatigue crack and a peripheral grinding zone of adjacent adjacent shaft sections, which, as a rule, accounts for about 30% of the total crack area. As the position of the cross section of the shaft with the crack changes during the rotation of the rotor (Figure 2), the moments of inertia of the cross section of the shaft with the crack change. So the smallest moment of inertia of the shaft cross section takes place in position I, when the crack opens under the influence of the rotor mass in the lower position. The moment of inertia of the cross section of the rotor shaft with a crack in positions II, III and IV is much larger due to the fact that position II and IV, the transverse axis, relative to which the moment of inertia is calculated, are located in the same way as the undamaged shaft, and in position III are in contact the flap flaps in the upper position objectively prevent a decrease in the moment of inertia of the cross section of the shaft with the crack.

Figure 1 shows the curves of the dynamic deflection of the rotor shaft of the second kind for all 4 positions of the cross sections of the shaft with a crack during rotation of the rotor. If we take as an option the case when the moment of inertia of the cross sections of the rotor shaft in position III is less than or close to the moments of inertia of the cross section of the shaft with a crack in positions II and IV, then in the process of rotation of the rotor shaft with a crack in the shaft an oscillatory process of the second kind occurs with a reverse the frequency determined by the difference in moments of inertia of the cross section of the shaft with a crack in positions I and II. In another case, when the moment of inertia of the rotor shaft in position III is significantly greater than the moment of inertia when the rotor shaft is cracked (II and IV), when the rotor rotates in addition to the second-order oscillatory process with a reverse frequency of rotation, a second-kind oscillatory process can be detected with double reverse frequent.

If we compare the previously mentioned dependences S = f (n) in the analysis of vibrational processes of the first and second kinds that occur with the rotor of a turbogenerator, which has a transverse crack in the shaft and a defect-free rotor, and determine the magnitude of the perturbing forces acting on the rotor, then with increasing frequency rotation from zero to the first critical first kind for a rotor with a crack in the shaft, the dependence S = f (n) passes much steeper (large values of vibration displacement) at the same values of rotation frequency) than for the same Specifications for defect-free rotor shaft.

Forces (F) causing oscillations of the rotor of the first and second kind are described by the dependence F = m · r · ω ² , i.e. the forces causing rotor vibrations for the same rotational speed are proportional to the unbalanced mass (m _n ) and the radius of its center of gravity (r) for the oscillatory process of the rotor of the first kind and, accordingly, the mass of the rotor (m _mouth ) and a change in its center of gravity υ (figure 2) with the oscillatory process of the rotor of the second kind.

If we take into account that m _mouth · υ >> m _n · r even with the initial occurrence of a transverse crack, as well as the fact that the magnitude of the acceleration in the oscillatory process of the second kind is higher than in the oscillatory process of the first kind for the same values of the rotor speed, then it becomes justified when ascertaining the formation of a transverse crack in the rotor shaft the use of a diagnostic sign of a more intensive increase in the vibration displacement of the shaft (bearings) with increasing rotor speed.

Therefore, in order to early identify transverse cracks in the rotor shaft, the dependences S = f (n) and φ = f (n) should be analyzed for the rotational speed from the moment the electric machine starts up until the rotor reaches the first critical rotational speed of the first kind and near it, and diagnose the appearance of a transverse crack in the rotor shaft, ceteris paribus, at the greater steepness of the dependence S = f (n).

Another diagnostic feature, which indicates the formation and development of a transverse crack in a flexible rotor, which at nominal conditions works near the 2nd critical speed, can be obtained based on the analysis of the dependence S = f (n), which, after the rotor exceeds the 1st critical the rotational speed tends to decrease as the second form rotor prevails in the vibrations. This tendency with an increase in the rotational speed from the first critical rotational speed to the nominal rotational speed is determined by the approximation of the oscillatory process of the bearings to the dead zone of rotor vibrations in the first form.

Due to the fact that the oscillatory process of the rotor of the second kind is planar in nature, and this plane is oriented vertically in the direction of the perturbing force of the weight of the rotor, to diagnostic signs indicating the appearance of transverse cracks in the rotor shaft, one can also judge by the increase in the steepness of the dependence S = f (n), which is most intensively manifested in the vertical direction of vibration displacement of the rotor shaft, vertical and axial directions of vibration displacement of the bearings and practically does not appear in the transverse pressure ION vibrosmescheniya rotor shaft and bearings.

From calculations of the critical rotation frequency of the oscillations of the shafts of rotors of the second kind with mass distributed along their length and various moments of inertia in at least one cross section of the shaft, it is known that each point located on the axis of the shaft moves along a circle of radius and with a double angular velocity, and the center the circle is shifted vertically by the amount of static deflection of the rotor. For close values of the moments of inertia in the mutually perpendicular axes of the cross section of the rotor shaft, the critical rotational speed of the rotor of the second kind is close to half the critical frequency of rotation of the first kind. Moreover, the phase of vibration displacement in the process of overcoming the rotor of the 1st critical rotation frequency of the second kind changes by 180 degrees. Considering that the oscillations of the rotor of the second kind are planar in nature (the vertical plane along the longitudinal axis of the rotor), while the oscillations of the rotor of the first kind from the influence of imbalance are spatial, and the measured vibration displacements are total values from the influence of the mass of the rotor and its imbalance, depending on the ratio of these forces and the phase of vibration displacement of an unbalance, which does not change at a critical number of revolutions, the total measured phase of vibration displacement becomes less than 180 degrees in. Therefore, the appearance of a transverse crack in the rotor shaft can also be judged by the intensive increase in the vibration displacement of the rotor shaft (bearings) at the first critical rotor speed of the second kind (near or below half the value of the first critical rotor speed of the first kind), and the change in the vibration displacement phase φ from insignificant in the initial stage of its development to 160-180 degrees at the final stage of crack development.

If we take into account the fact that a defect-free rotor of a turbogenerator with a certain degree of imbalance in the process of increasing the rotational speed during the passage of the first critical rotational speed changes the vibration displacement phase by 180 degrees, then the appearance of a transverse crack in the rotor shaft can be judged by the decrease in the magnitude of the phase displacement φ (from 180 degrees for a defect-free rotor) when the rotor overcomes the 1st critical rotation frequency of the first kind, and this decrease in the phase shift is determined by a change in the ratio influence of the unbalance forces of the rotor and the forces of the impact of the mass of the rotor.

If we take into account that during the oscillatory process of the rotor of the second kind, before the first critical frequency of rotation is reached, the vibration displacement phase has zero values with respect to the vibration displacement vector, then the area of the crack that appears in the rotor shaft and the zone of its maximum opening can be determined relative to the zero phase mark fixed on the rotor vibration displacement φ, measured in the region of the rotational speed of the rotor shaft to the 1st rotation frequency of the second kind.

When diagnosing the condition of the rotor shafts of electric machines by the proposed method, it is possible to use both individual diagnostic signs of cracks in the shaft and a set of diagnostic signs set forth in 6 claims, which will increase the reliability of the diagnosis of the condition of the rotor shaft, because the presence of any of the signs confirms the formation of cracks and the degree of its development, and the absence of any of the signs can only indicate a failure of the development of the crack and an unfavorable ratio of the oscillatory processes of the first and second genera.

Diagnosis of the state of the rotor shafts by the features of the claims can be carried out both during the start-up of the electric machine and in the process of stopping it. Preference should be given to diagnosing the condition of the shafts of an electric machine when starting it from a cold state, because In this case, distortions of the dependences s = f (n) and φ = f (n) related to variable instability during thermal instability of the rotor barrel and winding, displacement of retaining rings, etc., will be excluded.

Sources of information

1. Avrukh V.Yu. About some reasons for the increased vibration of synchronous compensators, "Electric Stations", 1960, No. 1.

2. Samoilov V.A. Power unit vibration and rotor balancing. - M., Gosenergoizdat, 1949, 160 pp.

Claims (6)

1. A method for diagnosing the condition of the rotor shaft of an electric machine containing a rotor barrel with end parts, retaining parts and bearings, including the detection of a rotor defect by comparing the experimentally determined dependence of the vibration displacement (S) of the rotor shaft (bearings) as a function of speed (n) [S = f (n)] for a rotor that did not have a defect with the same dependence of the diagnosed rotor, characterized in that the dependence S = f (n) of the diagnosed rotor is analyzed for the values of its rotational speed from the moment of starting electrically machines until the rotor reaches the first critical frequency of rotation of the first kind and near it and, at the same time, the formation of a transverse crack in the rotor shaft is diagnosed, ceteris paribus, at the greater steepness of the dependence S = f (n) of the diagnosed rotor after the occurrence of a defect compared to that the same characteristic of a rotor that did not have a defect. 2. The method according to claim 1, characterized in that the formation and development of a transverse crack in a flexible rotor, which at nominal mode operates near the second critical speed, is diagnosed by the form of the dependence S = f (n), which in the presence of a crack after exceeding the rotor the first critical rotational speed tends to decrease as the prevailing oscillations in the rotor in the second form. 3. The method according to claim 1, characterized in that the formation of a transverse crack in the rotor shaft, determined by the greater steepness of the dependence S = f (n) of the diagnosed rotor in comparison with the same characteristic of the rotor that did not have a defect, is diagnosed by a greater intensity of vibration displacement rotor shaft in the vertical direction and vibration displacement of the bearings in the vertical and axial directions. 4. The method according to any one of claims 1 and 3, characterized in that it further analyzes the dependence of the vibration displacement phase (φ) as a function of rotor speed (n) [φ = f (n)], and the formation of a transverse crack in the rotor shaft according to the intensive growth of the vibration displacement of the rotor shaft (bearings) at the first critical frequency of rotation of the second kind near or below half the value of the first critical frequency of rotation of the first kind, while the phase of vibration displacement φ can vary from insignificant in the initial stage of crack development to 160-180 degrees and the final stage of its development. 5. The method according to any one of claims 1 and 3, characterized in that it further analyzes the dependence of the vibration displacement phase (φ) as a function of rotor speed (n) [φ = f (n)], and the formation of a transverse crack in the rotor shaft to reduce the magnitude of the phase shift of the vibration displacement φ of the diagnosed rotor when the rotor passes the first critical frequency of rotation of the first kind, and this decrease in the magnitude of the phase shift from 180 degrees for a defect-free rotor is determined by a change in the ratio of the influence of the unbalance forces of the rotor and the force of sy rotor. 6. The method according to any one of claims 1, 3 and 4, characterized in that the region of location of the crack in the rotor shaft and the zone of its maximum opening are determined relative to the zero mark fixed on the rotor by the vibration displacement phase φ measured in the region of the rotational speed (n) rotor shaft from the moment of start up to the first critical speed of the second kind.

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