Vibration problems on long shaft

[kief99]

Hi hope anyone can help

I have encountered a problem during the commission of a back to back test rig for industrial gearboxes. The test rig is of mechanical type so the driveline is looped by means of two transfer boxes connected by a long tube, actually two tubes. The operating speed of rotation of these is 1600RPM.
The design of the tubes is such that is made out of a standard tube lenght 4.2m OD 216mm an ID196mm. There are solid hubs welded at the ends which step down the OD and are supported in two pilllow block bearings. The tubes are connected to the tranfer boxes and between each other using gear couplings.

The problem that we have is the vibrations are high on the bearing supports. We have increased the support rigidity with some sucess but not yet acceptable levels. One of the tubes was balanced at high speed up to 1000 RPM but was vibrating at 1300 RPM, And from the balancing machine data, vibrations increase with speed and unbalance also changes with speed. It seems that we are approching first order vibration as the phase of the unbalance force at the meassuring points are almost in phase.

Can anyone with experince sugget an approach to solve the problem?

Some of the options have been

To add a bearing on the middle of the tube restricting from whirling .i.e having the torque tube on three bearings.

To reduce the torque tube length and have three instead of two.

Try to balance the torque tube at operating speed of 1600RPM but would need to balnce it on the test rig as the balancing machine will not reach that speed.

Also other related questions: if the torque tube is balanced under operating speed and no load. when the load is applied the twist from end to end can be up to 0.7 degrees. Would this have a significant effect on the balance.

We suspect that the manufacturing tolerances of the tube section would make it very hard to balance since the tube is behaving as a flexible rotor.

If any one can suggest what would be the best approach to solve the problem or share some knowledge, would appreciate your input.

Thank you

 

[Strong]

A simple impact test or variable speed test would confirm that the tube shaft is probably operating near resonance. If that is the case, then you have to change tube dimensions or span between bearings. A 2-plane balance at low sped is probably not adequate if the speed is near or above the balance resonance speed, since the shaft is a flexible rotor and not rigid.

Walt

 

[kief99]

Hi walt thanks for the response, i think you are right we are operating near resonance, honestly i was expecting this to be lower than the calculated but the balancing machine shows there is phase change at ~700RPM and the calcualted critical speed is 2000RPM.

In the end we have chosen to modify the design by adding a bearing at the centre to contsrain it from whirling, so effectively we have the shaft supported on 3 bearings. Also the shaft will be changed to solid bar of a smaller diametre

Thanks again, any further comments are welcomed.

 

[FeX32]

If you are sure you are operating at close to resonance, a tuned dynamic vibration absorber (DVA) would work very well when tuned at that same frequency. It could simply be a small beam attachment ect. You could simply bolt/clamp/weld it directly to the tube.


Fe

 

[GregLocock]

Nice try but I wouldn't touch that with a bargepole.

Cheers

Greg Locock


New here? Try reading these, they might help FAQ731-376

http://eng-tips.com/market.cfm?

 

[electricpete]

Also other related questions: if the torque tube is balanced under operating speed and no load. when the load is applied the twist from end to end can be up to 0.7 degrees. Would this have a significant effect on the balance.

I was suprised to find that according to the reference below, the transmitted torque can lower shaft stiffness and critical speeds... apparently becomes more significant for long thin rotors:

The influence of axial torque on bending can be significant. The
equations of motion for a rotor in bending (no gyroscopic effect is
included) subjected to secondary effects such as a constant axial torque
have been given in [17], where it is shown that the stiffness of the shaft
decreases as the axial torque increases. This effect is predominant over
the other secondary effects if the slenderness ratio R/2L is lower than
0.0025. Experiments have been performed to show the influence of the
combined constant and pulsating axial torque [13, 18, 571

They give further discussion of finite element approach to calculating the effect.


=====================================
Eng-tips forums: The best place on the web for engineering discussions.

 

[Tmoose]

If your calculated bending critical is 2000 and it tested
(bump test) near that value, rules like "stay 10% or even 20% away from criticals" offer false sense of security. It is quite possible to find the initiation of whip at 50% or even below from a big unbalance due to wall thickness variation at the middle. in 1934 Den Hartog said it flat out, and by now all the rolls know it too.

Could have (1) bored the tube ID, or (2) prebalanced it with weights on the ID near the quarter points, or (3) field or machine balanced the assembled tube with scary weights on the OD near the quarter points.

When I was starting as a balance technician/specialist, the first few times we conscientously Low speed balanced a thick bellied, mean spirited assembled roll we missed the wall thickness problem every time.

http://icanhascheezburger.files.wordpress.com/2007/10/128347186987656250unsuspectingcat.jpg

 

[kief99]

Hi Timoose,
thanks for the input and the information, lesson learnt.

Regarding your options, balancing the shaft in field with weights on the OD could have been possible at quarter points but i am not confident enough in the methodology and the DAQ sytem required, would you be able to share a reference or a good book. As for the data aquisition i would have recorded a acceleration in one diredction on each bearing support and use a proximity probe pointed radially at the shaft at at a noth or something that could have worked as a reference of 1 revoultion. then work out the phase difference of the maximum acceleration and angle of rotation. then would add trial masses at each of the balancing points individually and look at the effect on the accelreations. from then move on to add masses on all points. but not sure how to work out the mass values. would that be a correct approach

regards

 

[tensor47]

Hi all,
I have field balanced flexible shaft lightly damped fans with a rotor between 2 bearings and I can assure you a single weight midspan will solve your problem.

You need to start slow and balance it at ever increasing speeds (determined by when the vibration becomes significant) using a balance ring with 36 small holes to minimize spliting weights between holes.

The theoretically required 1/2 mid span weight at each end at 180 for low speed balance is not a big issue in practice because the speed and therefore centrifugal forces are low AND most importantly the final mid span weight will be very small.

Cheers

John