I am trying to determine the required wall thickness of a vessel. There is one hydraulic feed hole in the wall of the vessel. My first prototype failed and cracked between the feed hole and outer diameter at approx 40,000 cycles.

Because of the feed hole I am not sure what formulas to use to calculate the problem or if there is a beneficial relationship to the placement of the feed hole. (Example, thicker wall between feed hole to O.D. or between feed hole to I.D.

Current internal operating pressure of hydraulic fluid in feed hole and bore = 5000psi

Redesigned Values

I.D. of vessel = 1.438 Inch

Max. internal pressure in 1.438 Inch Bore = ?

O.D. of Vessel = 1.995 Inch

Location of hydraulic feed hole in wall = .844 inches    from center

Diameter of hydraulic feed hole in wall = .118 inches

Total wall thickness (No Hole) = .28125

Wall thickness between feed hole and O.D. = .0945 Inch

Wall thickness between feed hole and I.D. = .066 Inch  

Yield strength of material 60000psi

Max. permissible stress 40000psi

Any help would greatly be appreciated.

Thanks in advance

Nick

nickjk,
If you can I would get professional assistance. It sounds that you are joining two pipes under high pressure in a cyclic service. This can be a tough design problem. You will probably need a piping designer, well versed in the air.  

To help clearify the application,

There is one cylinder or pipe.

The I.D. of the pipe is 1.438

The O.D. of the pipe is 1.995

The wall thickness is .28125

There is a .118 diameter hole in the wall.

I hope this helps.

Thank you,

Nick

A short comment;- the hole is not penetrating the wall like a shell opening, is a "hole" along the wall of the pipe. The internal pressure inside the pipe is identical with the pressure in the "hole" along the shell, 5000 psi. That is, the actual shell thickness for pressure calculation is the "Wall thickness between feed hole and O.D. = .0945 Inch" and the shell OD is the pipe OD = 1.995 Inch. That means the minimum thickness between the pipe OD and the "hole" diameter is 0.122 Inch, hence the given 0.0945 Inch is not enough to withstand the pressure, particularly in cycling service. It cracked precisely in the spot predicted by the Pressure Vessel Code.
Cheers,
gr2vesels

gr2vessels

Yes you are correct the hole is along the wall of the pipe, parallel with the I.D. of the pipe.

If I understand you correctly, because the pressure inside the pipe and the pressure inside the hole are equal we are not as concerned with the thickness between the I.D.of the pipe and the hole (does that mean the .066 wall in this area may ok). Therefore our calculated vessel thickness is between the hole and the outer diameter.

My research has found the highest stress to be at the I.D circumference and does the hole having equal pressure cause the deviation of that.

Also, where can I find information on the Pressure Vessel Code you mention.

Thank you so very much
Your view points are greatly appreciated

Nick

Hi Nick,
The minimum shell thickness calculation I used is the very basic one from ASME VIII, which is essentially identical with most of the recognised pressure vessel codes, EN 13445, PD5500, AS 1210, the Russian code, etc...If you need rigorous calculations, you should use an appropriate software, in order to consider other factors, including corrosion allowance, fatigue analysis;- I also doubt the correctness of the allowable design stress you mentioned as it sound excessively high. You should select the apropriate value as listed or calculated in accordance with the design code you are using. Please note that those design stress values differ from ASME to EN 13445 and other codes and is inapropriate to quote a certain value without referring to the code using that allowable stress.
However, from your reply I understand that you have limited knowledge of the pressure vessel design and I suggest to ask for the help of a professional pressure vessel designer who can help you. It will take years of studies to understand and apply the rules of pressure vessel design;- it will also take time to master any pressure vessel design software to the extent required for your little, but interesting exercise.
It would be very instructive to read a bit the posts and replies in this forum, in order to understand the depth of the engineering issues debated and analysed here.
Good luck,
cheers,
gr2vessels

nickjk, your geometry is not allowed by any of the national pressure vessel codes, so, if you had to certify it according to one of them, you would require detailed special calculations.
However, as you don't, you can calculate the basic stress in the wall by this formula, based on simple considerations of equilibrium that are used, without longitudinal holes of course, also in pressure vessel codes:
S=P(0.5+(D+d)/(2t-d)),
where P=5000, D=1.438, d=.118, t=.28125 and S=20 ksi
However this result doesn't account for fatigue: it explains only why you don't get a failure or a permanent deformation at the first pressurization.
For fatigue the notch effect caused by the hole comes into play. For this I suppose it would be beneficial, as the failure occurred at the OD, to put the hole closer to ID, but you are already too close to ID, and the gain would anyway be minimal.
From your quoted allowable stress, I suppose you are with an unwelded construction and a cold worked material, so you could have a too hard material, with low ductility and low fatigue resistance. As you have a good margin with respect to the basic stress, as shown above, I suggest a heat treatment or the choice of a ductile material, simply a standard forging or seamless pipe for pressure vessel construction, that you can easily have with a 20 ksi allowable stress.

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nickjk hydraulic jack is not a vessel and ASME code should not be applied.  nickjk should get a copy of ANSI B30.1 that deals with hydraulic jacking cylinders instead of using ASME codes.
Also, since the hydraulic jack appears to have load reversals,he should identify his max,min,range, endurance and yield stresses as well as a stress concentration factor in order to determine the hoop stress that governs this analysis. A reference would be MF Spotts "Design of machine elements" on working stresses.

Why dont you consider it...second pipe or hole... and the services considered as the cyclic services. For cyclic services make use of the integrally self reinforced contoured nozzle for its better reinforcement.  

I am sorry it took so long to respond, but I have been trying to do some research on my own.

I would like to thank all that responded to my inquiry.

As is evident it has been a long time since these items where covered in collage and I am light in this area, None the less I need to come up with a resolution for my problem.

Exactly what it is I am doing and what I am looking for.

I am designing a hydraulic component that needs to be minimal in size. (Note much room for over design)

I need a method to calculate preliminary wall thickness between the bore and feed hole also between the feed hole and outside diameter. These dimensions will be used in my preliminary design.

I know the answer is somewhere between Roark's Formulas for Thick Shell Pressure Vessels and Petersons Stress Concentration Factors.

I picked up Petersons Stress Concentration Factors today and I am finding very interesting but having a hard time applying my application to come up with a Concentration Factor.

Any help in this area would greatly be appreciated.

Thank You,

Nick