Conveyors-Ugh!!!
Originally Posted by trophywalleye
I would appreciate if someone could “shed some light” on why CEMA suggests in CEMA B105.1, the use of the ASME B105.1-83 fully reversed bending and steady torsion equation when calculating minimum pulley shaft diameters.
In summary my conveyor will cycle (accelerate, run const speed, drift to rest) approximately every 2 minutes for 20 years (give or take down times). Total cycle time is about 72 sec. During each cycle, acceleration takes up about ~5%, runs at constant speed ~90%, and drifts to rest ~5%. The conveyor drive is a planetary gearbox actuated by a hydraulic motor. The forces applied to the shaft are almost twice as much as the running forces, so both bending and torsion is increased approximately by a factor of two. The shaft is equipped with shoulders, and keyways, of which I have setup a calculation spreadsheet to help generate minimum diameters at my critical points.
I challenge the use of the fully reversed bending and steady torsion formula for three main reasons:
1.Torsion is not steady (from accelerating to constant speed)
2.Although bending can be considered fully reversed (only in one state, either accelerating or running), the amplitude does changes between the states
3.The fully reversed bending and steady torsion formula does not apply a stress concentration factor to the torsion portion of the formula (Norton’s book refers to this as Kfs)
4.CEMA also leads me to believe the shafts should be sized using constant speed (running speed) torsions and bending moments
The only reason I can think of why CEMA suggests the use of the fully reversed bending and steady torsion formula is that (usually) the acceleration cycles are significantly lower than the constant speed (running) cycles.
As many opinions as possible would be appreciated.
J
FYI – I have included the formulas in an attachment
Have you forgotten that on the head pulley, drive and driven pulley
and take up pulley(if hydraulic-where two would be needed for a mechanical take up)
on just one of of your conveyor flights has only six or eight points that are connected
in unison with the pulley drums themselves?
The six to eight taper locks/locking rings have only so much square area in
contact with the six to eight 5-15/16,6-15/16, 6-15/16, 7-15/16, 8-15/16....................,
size shafts and also in contact with the pulley bores in use for a conveyor
flights drive system wherein the head pulley and take up pulley actually are
passive and just along for the ride?????
This leaves you with only four taper locks/locking rings contacting the drive and
driven pulley doing the job of moving and ramping up the speed of a single
conveyor flight with the speed switches and flyball governors of the speed
switches-(that dates me).
lzaharis
Moderator ■
Load Assessment Systems
Dear J,
i would think that in the preamble (or the discussion sheets of the standard committee) of mentioned standard there will be wording in the line of "characteristic load assumption for ... conveyor pulley shafts ... dynamic loads at start-up and braking ... to be characteristically approx. X% of overall op - time ... occuring approx./max. 5 times per day (or so, i'm at home and cannot look it up)... etc. Your machine seems far away from a conventional bulk material belt conveyor, if not to the eye, then decidedly from its start - stop characteristics. So then CEMA or equal can imo not apply at all and one has to go back to basic mechanic design rules / standards for shaft design.
Ultimately one has to consider damage accumulation and load collective calculations according to true operational assumptions in order to get footing on factors / formulae.
Concluding: Imo your assumption concerning the reason is correct, but that leaves you with the more basic design desicions to be made based on more specific load assumptions as per your case.
The point is, that perhaps thus you might arrive at nonstandard equipment requirements (Not only shafts, don't forget bearings etc. .. and taper locks as Mr. Izaharis points out). Please be very thorough in transmitting your design parameters to suppliers etc. 20 years stop and go every 2 minutes, that's not going to be easy (and lightweight, at that). Perhaps a cut to 10 ys. and replacement of the drive are the ticket? Interesting project, from an engineering pov.
Best success!
Regards
R. ■
Re: Conveyor Pulley Shaft Sizing
Originally Posted by Roland Heilmanni would think that in the preamble (or the discussion sheets of the standard committee) of mentioned standard there will be wording in the line of "characteristic load assumption for ... conveyor pulley shafts ... dynamic loads at start-up and braking ... to be characteristically approx. X% of overall op - time ... occuring approx./max. 5 times per day (or so, i'm at home and cannot look it up)... etc. R.
Thanks for the reply Roland. Can you provide me or show me where to look for this committee discussion sheet that describes this? ■
Shafts At Engineered Pulley
Hello J,
i'm very sorry, but i really can't, i do not have the appropriate documents under my hand. At the moment, not even the full 105.1.
However, just in order to point, pls. cf. to the equation that leads to the safety factor(s) SF. There's
kg --> to be set to 1 for normal conveyor service --> pls. check, if not already done so.
I'm not sure, whether this covers all of your application, as the calculation of the moments relies on steady - state belt tension forces and considers thus the intermittent bending from shaft rotation but then not the number of extra loads from start-up, you mentioned it earlier. Ultimately, you could use the the maximum belt tensions as de facto design base, or on your own design decision, make up a load spectrum. You check the results of those calculations, perhaps it is possible already at that point to come to a safe & acceptable conclusion. Consider FEA.
If not, there's specialists around who do "engineered pulleys" and who might be of service.
Best regards
R. ■
Re: Conveyor Pulley Shaft Sizing
I believe this is the article that Roland is referring to. This is a snip it from CEMA 5 page 235.
Pulley Overloads
Normal running loads for standard pulleys should not exceed ratings in the ANSI
B105.1 and 501.1 load tables. Starting and occasional peak loads should not exceed
ratings by more than 50 percent. Overloads may result from such causes as starting,
jam-ups, maladjusted take-ups, brakes, misalignment and excess amounts of material
on the belt. Overload conditions may result in premature failure of either pulleys,
shafting, or bearings.
Normal running loads for engineered class pulleys should not exceed the loads
used for design. Starting and occasional peak loads should not exceed design loads by
more than 50 percent. If starting or peak loads are expected to exceed design loads by
more than 50 percent, that information should be provided to the pulley manufacturer
to be considered in the pulley design. ■
Gary Blenkhorn
President - Bulk Handlng Technology Inc.
Email: garyblenkhorn@gmail.com
Linkedin Profile: http://www.linkedin.com/in/gary-blenkhorn-6286954b
Offering Conveyor Design Services, Conveyor Transfer Design Services and SolidWorks Design Services for equipment layouts.
Conveyor Pulley Shaft Sizing
Hello,
I would appreciate if someone could “shed some light” on why CEMA suggests in CEMA B105.1, the use of the ASME B105.1-83 fully reversed bending and steady torsion equation when calculating minimum pulley shaft diameters.
In summary my conveyor will cycle (accelerate, run const speed, drift to rest) approximately every 2 minutes for 20 years (give or take down times). Total cycle time is about 72 sec. During each cycle, acceleration takes up about ~5%, runs at constant speed ~90%, and drifts to rest ~5%. The conveyor drive is a planetary gearbox actuated by a hydraulic motor. The forces applied to the shaft are almost twice as much as the running forces, so both bending and torsion is increased approximately by a factor of two. The shaft is equipped with shoulders, and keyways, of which I have setup a calculation spreadsheet to help generate minimum diameters at my critical points.
I challenge the use of the fully reversed bending and steady torsion formula for three main reasons:
1.Torsion is not steady (from accelerating to constant speed)
2.Although bending can be considered fully reversed (only in one state, either accelerating or running), the amplitude does changes between the states
3.The fully reversed bending and steady torsion formula does not apply a stress concentration factor to the torsion portion of the formula (Norton’s book refers to this as Kfs)
4.CEMA also leads me to believe the shafts should be sized using constant speed (running speed) torsions and bending moments
The only reason I can think of why CEMA suggests the use of the fully reversed bending and steady torsion formula is that (usually) the acceleration cycles are significantly lower than the constant speed (running) cycles.
The second formula in the attachment below is an equation for fluctuating bending and fluctuating torsion, of which I believe should be used.
As many opinions as possible would be appreciated.
J
Attachments
formulas (JPG)
■