EP2912272B1 - Method for misaligning a rotor blade grid - Google Patents

Description

The invention relates to a method for detuning a blade lattice.

A turbomachine has rotor blades arranged in the rotor blades, which can be regarded as firmly clamped at their blade roots and can oscillate during operation of the turbomachine. Depending on the operating state of the turbomachine, this can lead to oscillation processes in which oscillation states occur with high and critical stresses in the rotor blade. When the blade is stressed for a long time due to critical stress conditions, material fatigue occurs, which can ultimately lead to a reduction in the service life of the blade, which necessitates replacement of the rotor blade.

Due to centrifugal forces acting on the blade during operation of the turbomachine, a bias is generated in the blade. Due to this and the high temperature of the blade during operation, the natural frequencies of the blade during operation are different from the natural frequencies of the stationary and cold blade. As a quality assurance measure during production, only the natural frequencies at standstill of the turbomachine are measurable, but it is necessary for the design of the blade to know the natural frequencies under centrifugal force, so that the vibration processes in which the vibration states with the high and critical voltages in the Blade occur, can be avoided.

In the document EP 1 589 191 For example, a method of detuning a blade lattice is disclosed.

The object of the invention is to provide a method for detuning a blade lattice of a turbomachine, wherein the rotor blades have a long service life during operation of the turbomachine.

The inventive method for detuning, in particular the rotor-dynamic detuning, of a turbine blade having a plurality of blades has the steps of: a) setting for each of the blades of the blade grid at least one desired natural frequency ν _{F, S} , which the blade for at least one predetermined vibration mode in normal operation of the turbomachine under a centrifugal force has such that the vibration load of the blade lattice under the centrifugal force is below a tolerance limit; b) setting up a table of values ν _F (m, _S) r with selected discrete mass values m and radial center of gravity r _S, resulting from variations of the Nenrigeometrie of the blade and determining the respective natural frequency ν _F of the centrifugal force for each selected pair of values m and rs; c) measuring the mass m _I and the radial center of gravity position r _{S, I of} one of the moving blades; d) determining an actual natural frequency ν _{F, I of} the blade under the centrifugal force by interpolating the measured mass m _I and the measured radial center of gravity position r _{S, I} in the table of values ν _F (m, r _S ); e) in the case that ν _{F, I is} outside a tolerance of ν _{F, S} , selecting from the table of values ν _F (m, r _S ) of a pair of values m _S and r _{S, S} such that ν _{F, I} at ν _{F, S,} at least, and removing material from the blade such that m _I and r _{S, I} correspond to the value pair m _S and r _{S, S} ; f) repeating steps c) to e) to ν _{F, I is} within the tolerance of ν _{F, S.}

By measuring the mass m _I and the radial center of gravity position r _{S, I} and by interpolating these values in the value table ν _F (m, r _S ), the natural frequency ν _{F, I} under the centrifugal force can advantageously be determined with high accuracy. It is the same with the method according to the invention advantageously possible to set this natural frequency ν _{F, I} with a high accuracy and to approximate the specified target natural frequency ν _{F, S.} Thus, the vibration load of the blade during operation of the turbomachine can be reduced, thereby extending the life of the blade. In addition, the method is simple to perform because _, surprisingly enough, for an accurate determination of the actual natural frequency ν _{F, I} , m ₁ and r _{S, I of} the blade are measured without their full geometry. In addition, m _I and r _{S, I are} easily measured variables, for example, m _I can be determined by means of a balance.

The predetermined vibration modes are preferably selected such that the natural frequencies ν _{F, S} associated with the vibration modes are equal to or lower than a multiple harmonic of the rotor rotational frequency, in particular eight times the harmonic, one value table ν _F (m, r _s ) for a plurality or for all of the vibration modes is set up, the actual natural frequency ν _{F, I is determined} for each table of values and the value pair m _S and r _{S, S is} selected such that the determined ν _{F, I} to the set ν _{F, S} at least approximate.

The inventive method for detuning, in particular the rotor-dynamic detuning, of a turbine blade having a plurality of blades has the steps of: a) setting for each of the blades of the blade grid at least one desired natural frequency ν _{F, S} , which the blade for at least one predetermined vibration mode in normal operation of the turbomachine under a centrifugal force has such that the vibration load of the blade lattice under the centrifugal force is below a tolerance limit; b) establishing a table of values ν _F (m, r _S ) and a table of values ν _S (m, r _S ) with selected discrete mass values m and radial centroid r _s , resulting from variations of the nominal geometry of the blade, and determining the respective natural frequency ν _F under the centrifugal force and the respective Natural frequency ν _S at standstill of the blade for each selected value pair m and r _S ; c) measuring the mass m _I and the radial center of gravity position r _{S, I of} one of the moving blades; d) determining an actual natural frequency ν _{F, I of} the blade under the centrifugal force by interpolating the measured mass m _I and the measured radial center of gravity position r _{S, I} in the table of values ν _F (m, r _S ); e) in the case that ν _{F, I is} outside a tolerance of ν _{F, S} , selecting from the table of values ν _F (m, r _S ) of a pair of values m _S and r _{S, S} such that ν _{F, I} at ν _{F, S,} at least, and removing material from the blade such that m _I and r _{S, I} correspond to the value pair ms and r _{S, S} ; f) in case material has been removed, measuring the natural frequency ν _{S, I of} the blade at standstill; g) repeating steps e) to f) or c) to f) to ν _{F, I} within the tolerance of ν _{F, S} and ν _{S, I} within a tolerance corresponding to the tolerance by ν _{S, S.}

By additionally measuring the natural frequency ν _{S, I} , the actual natural frequency ν _{F, I} under the centrifugal force can advantageously be determined with an even higher accuracy. It is also possible to use only the measurement of the natural frequency ν _{S, I} at standstill to control the ablation, without repeating the measurement of m _I and r _{S, I.}

The predetermined oscillation modes are preferably selected such that the natural frequencies ν _{F, S} associated with the oscillation modes are equal to or lower than a multiple harmonic of the rotor rotational frequency, in particular the eightfold harmonics, one value table ν _F (m, r _s ) and one each Value table ν _S (m, r _S ) is set up for a majority or all of the vibration modes, the actual natural frequency ν _{F, I} and the actual natural frequency ν _{S, I is determined} for each table of values, the value pair m _S and r _{S , S is} selected such that the determined ν _{F, I} approach the fixed ν _{F, S} at least and the natural frequencies ν _{S, I} are measured for the predetermined vibration modes.

The variations in nominal geometry preferably include thickening and / or thinning of the blade in each radial cut or in radial sections. It is preferred that the variations in the nominal geometry have a linear variation in the thickness of the blade over the radius. It is advantageously possible to set up the value table by thickening and thinning the nominal geometry with an accuracy sufficient for determining the natural frequencies ν _F and ν _S.

The desired natural frequencies ν _{F, S} are preferably set such that adjacent blades arranged in the blade lattice have unequal nominal natural frequencies ν _{F, S} and that the desired natural frequencies ν _{F, S are} different from the rotor rotational frequency during normal operation of the turbomachine up to and including a multiple harmonic of the rotor rotational frequency, in particular the eightfold harmonics of the rotor rotational frequency. Thereby, it is prevented that a vibrating blade can excite a blade adjacent thereto to a vibration and that it comes to a coupling of the rotation of the blade lattice with the vibrations of the blades. Thus, the vibration loads of the blades are low and their lifetimes are long.

It is preferred that the measurement of the mass m ₁ and the radial center of gravity position r _{S, I} takes place relative to a reference blade, which has been measured three-dimensionally, in particular by means of a coordinate measuring machine and / or by means of an optical method. The accuracy of a measurement depends on the size of the measuring range, with a larger measuring range resulting in a lower accuracy. By measuring m _I and r _{S, I} relative to the reference blade, a small measurement range can be used with high accuracy. It is therefore only necessary to take a single blade as the reference blade and to characterize it once with a costly three-dimensional process, whereby also m _I and r _{S, I of} all other blades can be measured with high accuracy.

It is preferred that the value pair m _S and r _{S, S is} selected such that the imbalance of the rotor is reduced and / or that the effort for removal is minimal. The knowledge of the value pair m _S and r _{S, S} is sufficient for a balancing of the rotor, so that advantageous by the removal of the material can be done detuning and balancing of the blade grid in a common process step. The removal of the material can also be done so that the amount of material to be removed is minimized.

The predetermined vibration mode is preferably selected such that the natural frequency ν _{F, S of} the predetermined vibration mode is equal to or lower than the multiple harmonic of the rotor rotational frequency, in particular the eightfold harmonic of the rotor rotational frequency. The natural frequencies ν _F and / or ν _I are preferably determined by calculation, in particular by means of a finite element method.

It is preferable that when measuring the natural frequency ν _{S, I,} the blade is clamped to its blade root, the vibration of the blade is excited and the vibration is measured. The vibration is preferably measured by means of vibration sensors, acceleration sensors, strain gauges, piezoelectric sensors and / or optical methods. This is a simple method for determining the natural frequency.

By means of a comparison of the measured natural frequency ν _{S, I} with an actual natural frequency determined by interpolating m _I and r _{S, I} in the value table ν _S (m, r _S ), it is preferable to adapt the model for determining the natural frequencies ν _F and ν _S performed. As a result, influences of the material on the natural frequencies can be taken into account with advantage.

Figur 1

Längsschnitte von drei Laufschaufeln mit einer Nenngeometrie der Laufschaufel und Variationen der Nenngeometrie,

Figur 2

eine zweidimensionale Auftragung von Eigenfrequenzen ν_S der Laufschaufel im Stillstand und eine zweidimensionale Auftragung von Eigenfrequenzen ν_F der Laufschaufel unter Fliehkraft als Funktion der Masse m und der radialen Schwerpunktslage r_S der Laufschaufel und

Figur 3

ein Ablaufschema des erfindungsgemäßen Verfahrens.

In the following the invention will be explained in more detail with reference to the accompanying schematic drawings. Show it:

FIG. 1

Longitudinal sections of three blades with a nominal geometry of the blade and variations of the nominal geometry,

FIG. 2

a two-dimensional plot of natural frequencies ν _{S of} the blade at a standstill and a two-dimensional plot of natural frequencies ν _{F of} the blade under centrifugal force as a function of mass m and the radial center of gravity r _{S of} the blade and

FIG. 3

a flow chart of the method according to the invention.

FIG. 1 shows three blades 1 of a turbomachine, wherein the first blade in its nominal geometry 5, the second blade both in its nominal geometry 5 and in a first variation 6 and a second variation 7 and the third blade both in their nominal geometry 5 and in a third Variation 8 and a fourth variation 9 are shown. The rotor blades 1 have a blade root 2, which is fixedly mounted on a rotor shaft 4 of the turbomachine, and a blade tip 3 facing away from the blade root 2. In an oscillation of the blade 1 during operation of the turbomachine, a vibration node is arranged on the blade root 2. The radius r of the blade 1 is directed from the blade root 2 to the blade tip 3.

The second blade shows variations 6, 7 of the nominal geometry 5, in which, starting from the nominal geometry 5, the mass m is not changed, however, the radial center of gravity position r _S of the blade. In the first variation 6, the mass m is increased by uniformly thickening the second blade at each radial distance r from the axis of rotation and in the second variation 7, the mass m is reduced by uniformly diluting the second blade at each radial distance r.

In the variations 8, 9 of the third blade, starting from the nominal geometry 5, the thickness of the blade in the circumferential direction and / or the axial direction is varied linearly over the radius r. According to the third variation 8, starting from the nominal geometry 5, the blade is thickened at its blade root 2 and thinned at its blade tip 3, and according to the fourth variation 9, starting from the nominal geometry 5, the blade is thinned at its blade root 2 and thickened at its blade tip 3. As a result, in the third variation 8, the radial center of gravity position R _{S is} displaced radially inward and in the fourth variation 9 radially outward, whereas the mass m does not change. However, the variations 8, 9 can also be carried out such that both the mass m and the radial center of gravity r _{S are} changed. In addition, it is possible to carry out the mass m and the radial center of gravity position r _S by thickening and / or thinning the rotor blade 1 in selected radial sections.

A multiplicity of variations of the nominal geometry 5 are carried out and for each variation a natural frequency ν _S of the lowest frequency bending vibration of the blade 1 clamped at its blade root 2 and at a standstill is calculated by means of a finite element method. Furthermore, the natural frequency ν _{F of} the same bending vibration is calculated for each variation, taking into account the centrifugal force acting on the moving blade 1 during normal operation of the turbomachine. Optionally, when calculating ν _F , an increased temperature and thus changing material properties can also be taken into account. For a given blade lattice, it is advantageously only necessary to perform the variations of the nominal geometry once.

Then, for each variation of the nominal geometry 5, the mass m and the radial center of gravity r _{S of} the blade 1 are determined and a value table ν _S (m, r _S ) with value triplets ν _S , m, r _S and a table of values ν _F (m, r _S ) with value triplets ν _F , m, r _S. In the left application in FIG. 2 is the table of values ν _S (m, r _S ) and in the right-hand plot in FIG. 2 the value table ν _F (m, r _S ) represented by the respective natural frequency ν _S 10 and ν _F 11 is plotted against the mass m 12 and the radial center of gravity position r _S. The natural frequencies ν ν _S 10 and _F 11 are in arbitrary units and the nominal geometry 5 is in each case m = 0 and r _S = 0 is applied. Out FIG. 2 It can be seen that a reduction in the mass m and a shift in the radial center of gravity position R _S inwardly are accompanied by an increase in the natural frequencies ν _S 10 and ν _F 11.

In FIG. 3 the method according to the invention is shown in a flow chart. It is set for each of the blades 1 of the blade lattice a nominal natural frequency ν _{F, S} 14, which has the blade 1 for the lowest frequency bending vibration of the blade 2 fixedly clamped blade 1 during normal operation of the turbomachine under a centrifugal force, such that the Vibration load of the blade lattice below the centrifugal force is below a tolerance limit. This is achieved by having rotor blades adjacently arranged in the blade lattice having unequal nominal natural frequencies ν _{F, S} and that the nominal natural frequencies ν _{F, S are} different from the rotor rotational frequency during normal operation of the turbomachine up to and including 8 times the rotor rotational frequency.

Subsequently, for each nominal eigenfrequency ν _{F, S,} a corresponding desired natural frequency ν _{S, S is} determined 15, which has the blade 1 for the lowest-frequency bending vibration of the blade 1 firmly clamped to its blade root 2 at standstill. Subsequently, as described above, through the variations of the nominal geometry 5, the value table ν _S (m, r _S ) and the value table ν _F (m, r _S ) are set up 16.

After manufacturing 18 of the blade 1, its mass m and radial center of gravity r _{S are} measured 19. Subsequently, an actual natural frequency ν _{F, I of} the blade 1 under the centrifugal force by interpolating the measured mass m _I and the measured radial center of gravity r _{S, I} in the value table ν _F (m, r _S ) determines 17.

An actual target adjustment 21 is performed by comparing ν _{F, I} with ν _{F, S.} In the case where ν _{F, I} lies outside a tolerance of ν _{F, S} , a value pair m _S and r _{S, S is} selected from the value table ν _F (m, r _S ) such that ν _{F, I} an ν _{F, S is} at least approximated, and material is removed from the blade 1 such that m _I and r _{S, I} correspond to the value pair m _S and r _{S, S.} As it is from the right application FIG. 2 As can be seen, a plurality of value pairs m _S and r _{S, S} are generally available in order to achieve a certain natural frequency ν _{F, S.} From the plurality of value pairs, a pair of values m _S and r _{S, S} can be selected such that the rotor of the turbomachine is balanced and / or that the effort for removal is minimal. The removal 24 can be done for example by grinding.

To control the removal 24, the natural frequency ν _{S, I of} the blade 1 can be measured 20 at a standstill. For this purpose, the blade 1 is clamped to its blade root 2, the vibration of the blade 1 is excited, for example by a beat, and the sound emitted by the blade 1 is measured. Alternatively, to control the removal 24, the mass m and radial center of gravity r _{S of} the blade 1 can be measured 19. With a particularly high accuracy, the control can be performed by both the natural frequency ν _{S, I} 20 and the mass m and radial Center of gravity r _S 19 are measured.

It is also possible to measure both the mass m and the radial center of gravity r _S 19 and the natural frequency ν _{S, I} 20 before ablation 24 of the material, in order to obtain the actual natural frequency ν _{F, I} with a particularly high accuracy to eat. By means of a comparison of the measured natural frequency ν _{S, I} with an actual natural frequency determined by interpolating m _I and r _{S, I} in the value table ν _S (m, r _S ), an adaptation of the model for determining the natural frequencies ν _F and ν _{S be} performed.

In the event that ν _{F, I is} within a tolerance of ν _{F, S} , optional process steps 22 may be performed on the blade 1, such as applying a coating. Subsequently, the blades 1 is installed in the blade grid 23rd

While the invention has been further illustrated and described in detail by the preferred embodiments, the invention is not limited by the disclosed examples, and other variations can be derived therefrom by those skilled in the art without departing from the scope of the invention.

Claims (13)

1. Method for detuning a rotor-blade cascade, comprising a multiplicity of rotor blades (1), of a turbomachine, having the steps:

   a) establishing (14) for each of the rotor blades (1) of the rotor-blade cascade at least one setpoint natural frequency ν_F,S which the rotor blade (1) has for at least one predetermined oscillation mode during normal operation of the turbomachine under the effect of centrifugal force, such that the oscillation load of the rotor-blade cascade under the centrifugal force lies below a tolerance limit;

   b) compiling (16) a value table ν_F(m, r_S) with selected discrete mass values m and radial center-of-mass positions r_S, which result from variations (6 to 9) of the nominal geometry (5) of the rotor blade (1), and determining the respective natural frequency ν_F of the predetermined oscillation mode under the centrifugal force for each selected value pair m and r_S;

   c) measuring (19) the mass m_I and the radial center-of-mass position r_S,I of one of the rotor blades (1);

   d) determining (17) actual natural frequency ν_F,I of the rotor blade (1) under the centrifugal force by interpolation of the measured mass m_I and the measured radial center-of-mass position r_S,I in the value table ν_F(m, r_S);

   e) in the event that ν_F,I lies outside a tolerance around ν_F,S, selecting from the value table ν_F(m, r_S) a value pair m_S and r_S,S such that ν_F,I at least approximates ν_F,S, and removing (24) material of the rotor blade (1) in such a way that m_I and r_S,I correspond to the value pair m_S and r_S,S;

   f) repeating steps c) to e) until ν_F,I lies within the tolerance around ν_F,S.
2. Method according to Claim 1,
   wherein in addition to step b), a step b1) is carried out with the following features:

   b1) compiling (16) a value table ν_S(m, r_S) with selected discrete mass values m and radial center-of-mass positions r_S, which result from variations (6 to 9) of the nominal geometry (5) of the rotor blade (1),

   and determining the respective natural frequency ν_F,S of the predetermined oscillation mode with the rotor blade (1) at rest for each selected value pair m and r_S;

   f) in the event that material has been removed, measuring (20) and natural frequency ν_S,I of the rotor blade (1) at rest;

   g) repeating steps e) to f or c) to f) until ν_F,I lies within the tolerance around ν_F,S and ν_S,I lies within a tolerance around ν_S,S corresponding to the tolerance.
3. Method according to Claim 1,
   wherein the predetermined oscillation modes are selected in such a way that the natural frequencies ν_F,S associated with the oscillation modes are equal to or of lower frequency than a multiple harmonic of the rotor rotation frequency,
   in particular the eighth harmonic,
   wherein a value table ν_F(m, r_S) is respectively compiled (16) for a multiplicity of or all the oscillation modes, the actual natural frequency ν_F,I is determined (17) for each value table and the value pair m_S and r_S,S is selected in such a way that the determined ν_F,I are at least approximated to the established ν_F,S.
4. Method according to Claim 2,
   wherein the predetermined oscillation modes are selected in such a way that the natural frequencies ν_F,S associated with the oscillation modes are equal to or of lower frequency than a multiple harmonic of the rotor rotation frequency,
   in particular the eighth harmonic,
   wherein respectively a value table ν_F(m, r_S) and respectively a value table ν_S(m, r_S) are compiled (16) for a multiplicity of or all the oscillation modes, the actual natural frequency ν_F,I and the actual natural frequency ν_S,I are determined (17) for each value table and the value pair m_S and r_S,S are selected in such a way that the determined ν_F,I are at least approximated to the established ν_F,S and the natural frequencies ν_S,I are measured (20) for the predetermined oscillation modes.
5. Method according to one of Claims 1 to 4,
   wherein the variations (6 to 9) of the nominal geometry (5) comprise thickening and/or thinning of the rotor blade (1) in each radial section or in radial sections.
6. Method according to one of Claims 1 to 5,
   wherein the variations (6 to 9) of the nominal geometry (5) comprise a linear variation (8, 9) of the thickness of the rotor blade (1) over the radius.
7. Method according to one of Claims 1 to 6,
   wherein the setpoint natural frequencies ν_F,S are established in such a way that rotor blades arranged next to one another in the rotor-blade cascade have unequal setpoint natural frequencies ν_F,S, and that the setpoint natural frequency ν_F,S are different to the rotor rotation frequency of the turbomachine up to and including a multiple harmonic of the rotor rotation frequency,
   in particular the eighth harmonic of the rotor rotation frequency.
8. Method according to one of Claims 1 to 7,
   wherein the measurement of the mass m_I and of the center-of-mass position r_S,I is carried out relatively in a difference measurement with respect to a reference blade which has been three-dimensionally measured, in particular by means of a coordinate measuring device and/or by means of an optical method.
9. Method according to one of claims 1 to 8,
   wherein the value pairs m_S and r_S,S are selected in such a way that the unbalance of the rotor is reduced and/or that the outlay for the removal is minimal.
10. Method according to one of claims 1 to 9,
    wherein the predetermined oscillation mode is selected in such a way that the natural frequency ν_F,S of the predetermined oscillation mode is equal to or of lower frequency than a multiple harmonic of the rotor rotation frequency,
    in particular the eighth harmonic.
11. Method according to one of Claims 1 to 10,
    wherein the natural frequencies ν_F and/or ν_I are determined computationally,
    in particular by means of a finite element method.
12. Method according to one of Claims 2, 4 to 11,
    wherein, during the measurement of the frequency ν_S,I, the rotor blade (1) is clamped at its blade root (2), and the oscillation of the rotor blade (1) is excited and measured.
13. Method according to one of Claims 2, 4 to 12,
    wherein adaptation of the model for determining the natural frequencies ν_F and ν_I is carried out by means of a comparison of the measured natural frequency ν_S,I with an actual natural frequency determined by interpolation of m_I and r_S,I in the value table ν_S(m, r_S).

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