Circ. Stress at Horn of Saddle
Circ. Stress at Horn of Saddle
(OP)
How can the Circ. Stress at Horn of Saddle be reduced?
Actual= 227.41 MPa Allowable= 135.31 MPa
Otherthings are fine.
Shell length= 15 m
Shell Dia= 3 m
Thanks n Regards
Actual= 227.41 MPa Allowable= 135.31 MPa
Otherthings are fine.
Shell length= 15 m
Shell Dia= 3 m
Thanks n Regards





RE: Circ. Stress at Horn of Saddle
S =-Q/(4*t*(b+1.56*Sqr(Ro*t)))-3*K*Q/(2*t2)
Assuming that you are fixed for most variables such as:
shell thickness, t
shell radius, R
saddle load, Q
That means you can fiddle with the saddle width, b, and the saddle and wear plate angle in order to affect Zick factor K. So if you increase the saddle width and/or increase the saddle and wear plate angles you should be able to reduce the calculated stress.
If this step doesn't reduce the stress sufficiently (and it seems like you have a long way to go to get this stress down) then you'll have to try adding a stiffener ring, or move the saddles closer to the heads so they act as stiffeners, or increase the thickness of the shell.
RE: Circ. Stress at Horn of Saddle
best regards,
Mandeep Singh
RE: Circ. Stress at Horn of Saddle
Very informative.
RE: Circ. Stress at Horn of Saddle
What's the logic behind that: Unequal and uncontrolled division of weight between the three saddles?
RE: Circ. Stress at Horn of Saddle
So... if you run across a three saddle design, ensure that the foundation is monolithic, and even then, be suspicious that the loading may not be what you expect. A two saddle weight distribution is far more reliably known.
jt
RE: Circ. Stress at Horn of Saddle
It is turning out that I will also need to work on a vessel with 3 saddles in some time.
Can you please share as to what approach did you use for anaylsing vessel on 3 saddles.Somebody from my office suggested using STAADPRO.Since I have no experience dealing with such vessels,can you please suggest how to tackle this case?
Starrproe
RE: Circ. Stress at Horn of Saddle
Results of the three methods do not seem to me to agree very well.
Regards,
Mike
RE: Circ. Stress at Horn of Saddle
I used a rather large model, one which I would only have dreamed of a decade ago. Used SolidWorks to build a quarter model including contact conditions for the wear plate to shell and baseplate to pedestal (the vessel had an unusual baseplate / pedestal setup). You might be able to do something like that with shell elements, but I used high order solids. Without getting into details, I'll just point out that I was a bit perplexed by the result until I ran each loading individually and realized that their combinations resulted in a situation which I did not anticipate but was reasonable.
I guess I'd say there's more going on here than I anticipated, even as a somewhat experienced FEA / vessel design guy.
Mike-
I'd suspect that the FEA - if properly done - is most likely the better result, but clearly some "back of the envelope" hand calc's should back up the order of magnitude of the results. Wondering how your beam program and FEA models incorporate the weight / stiffness of the tube bundle and tubesheet. Seems that the relative deflection of the (cantilevered off the tubesheet) tube bundle vs the shell may be playing a role. [Have you considered writing up a paper regarding your analysis methods to be published / presented at the ASME Pressure Vessels and Piping Conference? Rumor has it it'll be US central east coast next summer...]
jt
RE: Circ. Stress at Horn of Saddle
Regarding my beam program, it does not account for tube bundle stiffness, it just treats the exchanger as a tubular beam with distributed loading, no doubt a source of error. The in-house program I'm sure does the same.
This discussion well illustrates why more than two supports should be avoided if possible. Two supports, reactions found in seconds with a calculator, more than two, uh-oh...
As for papers, what I've seen is not nearly formal enough to write up:)
Regards,
Mike