Flat Head Calculation

What is the formula in ASME Sec 8 Div 1 UG-34?

t min= (d √(CP/ SE)

I forget what C means in that formula. What value are you using for E? If you illlustrate with the numbers you are using, we can see if 35mm makes sense for the head thickness. (Off the top of my head, I think it does.)

Thanks guys for all reply.....But still....

The contactor need a flat head design.the vessel is not a high pressure vessel.As i told earlier the design pressure is 20 PSI.My concern is for a vessel with less internal pressure but with a bigger internal dia (like 3.5m) why we require a high thickness flat head.when we use a high thickness flat head will there be any buckling effect on the shell(since it is of only 12 mm thick)

Valuses i have been using is C factor=0.2

inside dia of unstayed flat head=146.61 inch(3724mm)

Joint efficiency E =0.85

Maxi allowable stress=20000 PSI

thats how i got a value of 48 mm thick

Looks very surprising that 48mm thick plate to cover a low pressure pressure vessel

with a shell thickness of 12mm

Please advice

Antony, using joint Efficiency E=1 (Increasing NDE´s like RT or UT),

Significantly reduce the thickness required for

the flat head plate.

Hai

It is not unusual for an ASME pressure vessel flat head thickness to be 3 to 5 times the calcculated thickness of the shell.

Flat head closures only make economic sense for small diameter vessels.

You should not be surprised at your result of 35 mm.

Consider a concave vessel head.

Tell us more about your vessel contents, service and why you desire a flat head design.

Thanks guys for all reply.....But still....

The contactor need a flat head design.the vessel is not a high pressure vessel.As i told earlier the design pressure is 20 PSI.My concern is for a vessel with less internal pressure but with a bigger internal dia (like 3.5m) why we require a high thickness flat head.when we use a high thickness flat head will there be any buckling effect on the shell(since it is of only 12 mm thick)

Valuses i have been using is C factor=0.2

inside dia of unstayed flat head=146.61 inch(3724mm)

Joint efficiency E =0.85

Maxi allowable stress=20000 PSI

thats how i got a value of 48 mm thick

Looks very surprising that 48mm thick plate to cover a low pressure pressure vessel

with a shell thickness of 12mm

Please advice

Have you considered that 146.61" Diameter (73.3" radius) has (∏R^2)=16,900 square inches of area, which means 338,000 pounds, or 169 tons of force at only 20 psi?

thats what i am also asking?

A flat head without the stiffeners will be very thick and uneconomical. Use Ms Flat stiffeners which reduces the span. The exact design details can be found in a book by Timoshenko, which I have used during my haydays.

Hope it helps

looks like a good idea.can you expain further...i didnt get completly

I can try to explain more. You have to weld MS Flats of a adequate size (50 X 6 should be fine) at a spacing of 700 mm. This means you will have 4 nos. Longitudinal Stffeners. Pls ensure that the 50 mm dimension is kept vertical to the plane of the Flat Plate. In between the adjacent rows of these stiffeners, weld another set of stiffeners perpendicular to the previously welded rows. You can keep the same distance. It will form a Grid like structure with the middle grids of size 700 mm x 700 mm. The required plate thickness with this kind of arrangement will be much less.

Also, Pls check and confirm whether a vessel with 20 psi Internal pressure falls under ASME SECT VIII. Your vessel looks like a storage vessel and .

We used to design the stoarge vessels as per API codes.

I have tried to explain and I hope it is conveyed in the true sense.

thanks a lot....But i cant corelate your suggetion to to code...if possible please tell where this option is mentioned in the code.i read through Sec 8.Vessels with internal pressure greater than 15 psi can comes under sec 8.

Unfortunately the ASME Code did not mention anything on this type of designs. I am not aware of any amendments that may have been added later on. Although, the use of Stiffeners for External Pressure for Shell design is very much covered in the codes.

Basically, the design formulae given in the codes is derived from the STRENGTH OF MATERIAL CONCEPTS, which are based on the assumption of thin shells. The code mainly defines the allowable stress values of the material being used, Welding parameters, Joint Efficiencies, Stress correction factor due to thick shells and so on.

Also, the custom design can be submitted to ASME divisions for their approval.

Hope it explains.

Flat heads aren't good for much. Beer bottles even fall off.

Flat heads are the weakest type of head covered under Section VIII. You should get a head thickness greater than the shell thickness. I think that's what you were asking for right?

Your calculation is OK, see the attached example:

As per ASME VIII-1, UG-34 c(2), The minimum required thickness of flat unstayed circular heads, covers and blind flanges shall be calculated by the following formula:

Example:

Fig. UG-34 Some Acceptable Types of Unstayed Flat Heads and Covers

For the same conditions, except the vessel diameter d = 350 mm instead of 3500 mm:

Conclusion: Flat head thickness is about 7 mm for vessel diameter 350 mm, where it was about 70 mm in case of vessel diameter 3500 mm.

Because flat head with larger diameter will experiences greater normal stress due to bending moment. For the body of your pressure vessel, hoop stress is dominant, but for flat end, bending stress is dominant. Remember, larger the diameter, greater the bending stress. Therefore requires higher thickness.

Dear All,

Please give me idea about flat cover of head design under external pressure.

Means Design of flat head under external pressure. (In ASME CODE Sec 8 Div 1 UG 34 is only for Internal Pressure Only?)