ESTABLISH A SYSTEM IDENTIFICATION METHODOLOGY OF MODAL TESTING USING THE ANALYTICAL/COMPUTATIONAL TECHNIQUES IN ROTATING MACHINERY

H. Al-Khazali1, M. Askari21 Faculty of Science, Engineering and Computing, School of Mechanical & Automotive Engineering, Kingston University, London, UK, k0903888@kingston.ac.uk

2 Faculty of Science, Engineering and Computing, School of Aerospace & Aircraft Engineering ,Kingston University, London, UK, M.Askari@Kingston.ac.uk

IntroductionDevices such as engines, motors, disc drives and turbines alldevelop characteristic inertia effects that can be analyzedto improve the design and decrease the possibility of failure.At higher rotational speeds, such as in a gas turbine engine,the inertia effects of the rotating parts must be consistentlyrepresented in order to accurately predict the rotorbehaviorThe experimental technique used thus far is called ModalTesting, a well known and widely used technique in researchand industry to obtain the Modal and Dynamic responseproperties of structures. The technique has recently beenapplied to rotating structures and some research papersbeen published, however the full implementation of ModalTesting in active structures and the implications are not fullyunderstood and are therefore in need of much further andmore in depth investigations.The raw data obtained from experiment was used in

Simulation finite element(FE)model(ANSYS12)for comparison.Since it has good capability for Eigen analysis and also goodgraphical facility, and obtained good result. 3-D modelsresult large number of nodes and elements. This paperdemonstrates how to extract a plane 2-D model from the 3-D model that can be used with fewer nodes and elementsbecause of ease of use, accuracy and performance

Aims & ObjectivesThe aims is to establish a system identification methodologyusing the analytical/computational techniques and updatethe model using experimental techniques alreadyestablished for passive structures but to active rotatingstructures, which subsequently help to carry out healthmonitoring as well as further design and development inrotating machinery.Excess vibration can cause noise and wear in machinery ,it isimportant to identify all the critical speeds within the rangeof operation and analyse the damping effects, massunbalance and other phenomena in rotating machineryand their effects in the safe operation of rotating machinery,which would be the subject of study in this research.In this study the raw data obtained from the experimentalresults are curve-fitted by theoretical data regenerated fromsome of the experimental data. In this project the study ofthe effects from the bearing damping and stiffness will takeplace. It is envisaged that the approach is not limited to thecondition monitoring and predictive failure but could helpthe designers to have better understanding of rotorperformance at the system design stage.

Simulation of model in (ANSYS 12)A model of rotor system one disc with multi degree offreedom (Y and Z directions) has been used to demonstratethe above capability see Fig.4 & 5. A program has beenwritten in (ANSYS 12), Post-processing commands (/POST26).Applying of gyroscopic effect to rotating structure wascarried by using (CORIOLIS) command. This command alsoapplies the rotating damping effect. (CMOMEGE) specifiesthe rotational velocity of an element component about auser–defined rotational axis.(OMEGA) specifies the rotationalvelocity of structure about global Cartesian axes. Model thebearings using a spring/damper element (COMBIN14).Another command which was used in input file(SYNCHRO) that specifies whether the excitation frequency issynchronous or asynchronous with the rotational velocity of a

structure in a harmonic analysis .The finite element (FE)method used in ANSYS offers an attractive approach tomodelling a rotor dynamic system. While it may require morecomputational resources compared to standard analyses.

Fig. 2. Impact test.

Test SetupExperimental modal analysis has grown steadily in popularity since the advent ofthe digital FFT spectrum analyser in the early 1970’s.Today,impact testing (or bumptesting) has become widespread as a fast and economical means of finding themodes of vibration of a machine or structure. The test rotor is shown in picture .1.Basically, the rotor consisted of a shaft with a nominal diameter of 10 mm, with anoverall length of 610 mm. Two journal bearings, RK4 Rotor Kit made by Bentley

Nevada (the advanced power systems energy services company), could be usedto extract the necessary information for diagnostic of rotating machinery, such asturbines and compressor. Been testing the process will be conducted on the rotarymachine as the project is based on rotary dynamics reach practical results for thepurpose of subsequently applied machinery rotary by using (Smart office program,the smart Office is the software which is used in this project. Then, do theexperimental testing using the impact test, installed fixed two accelerometer(model 333B32, sensitivity 97.2 &98.6 mv/g) in Y& Z direction and roving the hammer(model 4.799.375, S.N24492).Configurations of testing on the rotary machines, allnecessary equipment for test was shown with the geometry design wizard (Fig.6).

Fig.6. Geometry design for modal (one disc in the middle ),

experimental test using smart office;

Discussion and Further DevelopmentsInvestigating various aspects of bearing behaviour of an induction motor system for solvingits failure problem of the shaft in machine rotating with high speeds. It is found that thebearing plays the major role in the restraint of the vibration amplitudes of the rotor, ofstationary load that cause reduce load in the bearing and then give the bearing long andsave life.Investigating various aspects of the vibration behaviour of an induction motor system forsolving its failure problem of the shaft.The outcome of the study is to develop a multipurpose modal analysis, using modal testingin vibration analysis, Experimental part. This poster demonstrates how to extract a plane 2-D model (Fig.5 from the 3-D model Fig. 4), which can be analyzed using far fewer nodes,elements, ease of use, accuracy, and performance. 2-D and 3-D model analyses areclosest match each other when comparison with experimental part see Table 1 & Fig.12.From the Fig.10 notes the decreased the damping ratio caused increased naturalfrequency until reach max. amplitude when the system reach resonance ω = ωn ,whendamping ration(ζ =0)(free vibration).From Fig.16 ,17 shows the (FRF) of stationary rotor setup (A), versus achieved unbalancedstationary frequency response function of the setup (B), as it can be seen from the figuresabove, the (FRF) amplitude of the system has reduced by 12 db for first mode, 11db forsecond mode, due to applying an excessive 8 grams of loading to the system at the samespeed. That mean the amplitude is reducing in the shaft after add mass in the disc withangle (135°, 315°) ,we expect the noise is reduce as well due to reduce excess vibration.With further simulation we find the relation between the bending stress with respect timewe can see from the Fig. 19-A,B&C ,the performance of bending stresses at the disc in themiddle, we see before added 8 gram mass and after.We used (ANSYS)to find the relation between the reaction bearing forces(N ) with respecttime we can see from the Fig. 15 the performance of reaction forces in the right and left

bearings with different speed when increasing speed of rotation the reaction forceincreasing for both right and left bearings but from the figure above we see the maximumreaction force in Y direction in left and right bearings when the motor run up, after a fewsecond is become decreases while the reaction in Z direction is began increasing slowly inleft and right bearings until reach maximum value when the speed is increasing. Thatmean we must take care to right bearing when run up the motor because this bearingcarry maximum reaction force at the began.For further studies, increasing discs numbers while ANSYS gives accurate results.

Results

A-At start rotation.

Fig.4.Finite element modal rotating machinery, one disc.

Three dimensions (3-D);

A-ANSYS (APDL). B-(ANSYS workbench).

Fig.5. Finite element modal rotating machinery ANSYS

(APDL).Two dimensions (2-D).

Picture.1.Experimental setup for

the modal testing.

Fig.7.Schematic of a rotor one

disc setup parameter.

.

Fig.8. Stationary load in the middle, one disc, (FRF)

versus Frequency (Hz), (first mode shape). Natural

frequency, 29.79 Hz, rang (0-500) Hz.

B-Natural frequency, 29.95 Hz(3-D)((first mode

shape).

Fig.9. Finite element method simulations, one disc

in the middle, with two bearings;

A-Natural frequency , 29.93 Hz (2–D)(first

mode shape).

Table 1 Comparison Between Natural Frequency Outcomes

From Experiment and ANSYS, (One Disc).

Table 2 Natural Frequency and Damping Ratio (ζ) For

One Disc Rang (0-500) Hz, (Experimental Part).

Modelling Effect; (One Disc).

Reaction forces in the left and right bearings.

Fig.12. Natural frequency experiment with

ANSYS ,(one disc).

Fig.15.Relation between reaction bearing forces versus

time(s) at different speeds of rotations(one disc);

A-Reaction force( Fy)

left bearing .

B-Reaction force( Fy)

right bearing.

C-Reaction force (Fz)

left bearing.

D-Reaction force(Fz)

right bearing.

Fig.13. Comparison between

measured & predicted

frequency (Hz).

Fig.14.Measured & Predicted frequency

(Hz),[From Ref: Ewins, D.J, 1995, Modal

Testing: Theory and practice, John wily,

Exeter, England.].

Fig.10. Damping ratio(ζ)

versus natural frequency

(0-500)Hz.

Fig.11.Variation of amplification ratio

with r ,[From Ref: Singiresu , S . Rao ,

2005, Mechanical Vibrations, Prentice-

Hall, Inc. Si Edition. , Singapore].

A- Before add mass ,one disc, (second

mode shape).

B-After add mass (8 gram at 135°,315° degree) one disc,

in the middle, (First mode shape).

Fig.16 Stationary load (FRF)versus frequency (Hz), speed

of rotation 1000 rpm ;

A-Stationary load (FRF) versus frequency (Hz),

before add mass, one disc, (First mode shape).

B- After add mass (8 gram at 135°, 315° degree) in the

one disc,( second mode shape).

Fig.17. Stationary load in the middle ,(FRF) versus

frequency (Hz) speed of rotation 1000rpm;

ANSYS Results

A-Displacement versus

time before add mass.

B-Displacement versus time

after add mass 8 gram to

the disc.

C- Merge comparison.

Fig.18. Amplitude versus time,(A-With out load ,B-

After add 8 gram mass, C-Merge);

A-Before add

mass. .

B- After add 8 gram mass

at the disc.

C- Merge in Y

direction.

D-Merge in Z

direction.Fig.19.Relation between the bending stress with

respect time (sec.),one disc;

Fig.3. Shaker test.

What is the mode shape ?A vibrating structure by nature deforms into certain

patterns at the natural frequencies of the structure.The deformation patterns and their natural frequencies

depend only on the characteristics of the structure, and noton the input force .

These deformation patterns are the mode shapes of thestructure(Fig.1).

Fig.1.Mode shape.

N

r rrrr

rjri

jij1

22,2

)(

M

φφH

Equation of motion for definition of frequency response functions (FRF)

fxKxCx

M

rrrr

rrr

rr

r

matrixFRF

HHHH

HHH

HH

H

,44,34,24,14

,33,23,13

,22,12

,11

Unbalance Effect; add mass.

Methods

Rotating shaft

B- After rotation (first mode shape).

Mode shape ωn (ANSYS)one disc(Hz) Frequency one disc experiment(Hz) Error %

1 29.939 29.79 0.497678613

2 243.71 242.74 0.398014033

Name

Natural

Frequency (Hz)

Damping

Ratio(ζ) % Modal A[kg/s]

Mode1 29.79 56.701 0.009123329 -(0.01562434)

Mode 2 242.74 7.561 0.001488148 -(0.002922353)

HAMMER