My company is looking to use a Ø3/4" pipe nipple to hang a load of less than 100lbs in a particular assembly. Some concern exists over the strength of the pipe threads being loaded in tension. I looked for a way to analyze this problem but cannot really come up with anything. Most of the formulas for tensile stress area are assuming a straight thread. I'd like to be able to calculate the stress on the threads and then compare that to the strength of the pipe material to determine a safety factor? Can anyone help?
Tek-Tips is the largest IT community on the Internet today!
Members share and learn making Tek-Tips Forums the best source of peer-reviewed technical information on the Internet!
-
Congratulations cowski on being selected by the Eng-Tips community for having the most helpful posts in the forums last week. Way to Go!
Pipe Thread Stength 1
- Thread starter jaydigs
- Start date
jaydigs,
Check out this link. You can work backwards from the pressure limit to the axial force on the thread. This will give you a ballpark limit strength.
Timelord
Check out this link. You can work backwards from the pressure limit to the axial force on the thread. This will give you a ballpark limit strength.
Timelord
- Thread starter
- #4
The reason for the pipe thread is that we have components attached to both ends of the pipe (neither require water tight connections, but they do use tapered threads).
Also, the equation in the article would allow you to back out the tensile strength of the material but I can look that up. I think the problem is in calculating the tensile stress area for tapered threads. I need this area to get the stress on the threads.
Thanks
Also, the equation in the article would allow you to back out the tensile strength of the material but I can look that up. I think the problem is in calculating the tensile stress area for tapered threads. I need this area to get the stress on the threads.
Thanks
jaydig,
Just use the actual burst pressure for your size pipe and multiply by the inside area which will be the axial force at failure (from tables at end of article). If your material is not one listed, you know your material strength, just ratio from a listed material.
Timelord
Just use the actual burst pressure for your size pipe and multiply by the inside area which will be the axial force at failure (from tables at end of article). If your material is not one listed, you know your material strength, just ratio from a listed material.
Timelord
The minimum Pitch Diameter of the 1st external thread of a 3/4-14 NPT thread is nominally 0.968", the maximum height of thread is nominally 0.062"; that gives a minimum Minor Diameter of 0.844. I would use the Minor Diameter calculated for the OD value below:
Stress = F/A = (F / ((pi/4)(OD^2-ID^2)))
Is the load constant or does it cycle?
- Thread starter
- #8
If you multiply the burst pressure for the smallest Ø3/4" pipe (with my ID of .824") you get an area of .533in^2 so mulitplying by 25400psi gives you an axial force at failure of 13544lbs. I doubt 1/2" to 3/4" thread engagement of Ø3/4" pipe will support that much weight.
The load is static only.
I don't think using a= piD^2/4 where D is the difference between 2 diameters is sufficient either. You would have to take into consideration the length of engagement. From simple area formula you get something like .0003141in^2 which seems very low, and it doesn't consider thread engagment length.
The load is static only.
I don't think using a= piD^2/4 where D is the difference between 2 diameters is sufficient either. You would have to take into consideration the length of engagement. From simple area formula you get something like .0003141in^2 which seems very low, and it doesn't consider thread engagment length.
jaydigs,
Why do you doubt the answer? The paper is reporting the results of actual tests. An assumption I think you can make is that the threads are cut per spec and made up per spec (see Machinery's Handbook for specs). The ~6.5 ton answer is quite believable to me, especially for the higher strength materials listed and 3/4 in pipe. You need to ratio the answer by the ratio of the ultimate strength of your material to the ultimate strength of the listed material to correct the answer for your case.
Another argument you can make is that only a few of the failures were by thread shear, most cases the pipe burst first. Calculate the burst pressure of your pipe with your material and the calculate the axial force that the burst pressure will exert. Should be close to the ultimate thread shear strength. I suspect the number will be at least an order of magnitude or more than your load, probably 2 or 3 orders of magnitude. In other words your load is trivial unless you are making the threads of some extremely weak material.
Timelord
Why do you doubt the answer? The paper is reporting the results of actual tests. An assumption I think you can make is that the threads are cut per spec and made up per spec (see Machinery's Handbook for specs). The ~6.5 ton answer is quite believable to me, especially for the higher strength materials listed and 3/4 in pipe. You need to ratio the answer by the ratio of the ultimate strength of your material to the ultimate strength of the listed material to correct the answer for your case.
Another argument you can make is that only a few of the failures were by thread shear, most cases the pipe burst first. Calculate the burst pressure of your pipe with your material and the calculate the axial force that the burst pressure will exert. Should be close to the ultimate thread shear strength. I suspect the number will be at least an order of magnitude or more than your load, probably 2 or 3 orders of magnitude. In other words your load is trivial unless you are making the threads of some extremely weak material.
Timelord
This pipe nipple is being used as a rod hanger, right?
The allowable stress for a low grade Carbon Steel pipe is 15ksi at ambient, that would allow a 390 lb load. If it is Mallable Iron, 10ksi would be an allowable stress which would allow you a 260 lb load. For a sch40 nipple. Use a sch80 for safety.
Allowable Stress does not equal Ultimate Stress.
jaydig,
An additional comment that may make you more receptive to the results. A single (grade 8) 3/8-16 bolt has an minimum ultimate tensile strength of 13,900 lbs (right off my screw selector from Holo-Chrome). Can you believe that a 3/4" pipe joint is about the same axial strength as a single 3/8 bolt? The thread shear area is certainly a lot less in the bolt.
Timelord
An additional comment that may make you more receptive to the results. A single (grade 8) 3/8-16 bolt has an minimum ultimate tensile strength of 13,900 lbs (right off my screw selector from Holo-Chrome). Can you believe that a 3/4" pipe joint is about the same axial strength as a single 3/8 bolt? The thread shear area is certainly a lot less in the bolt.
Timelord
RCHandy,
If jaydig is concerned about thread pullout strength, going from schedule 40 to schedule 80 pipe gets you nothing. The threads are the same. Is the 390 lbs that you calculated the thread pullout strength or the tensile strength of the pipe? If it is the actual pullout strength of the threads, show jaydig how you arrived at it to convince him his load is trivial. Working vs ultimate strength has a sizable safety factor in it.
Timelord
If jaydig is concerned about thread pullout strength, going from schedule 40 to schedule 80 pipe gets you nothing. The threads are the same. Is the 390 lbs that you calculated the thread pullout strength or the tensile strength of the pipe? If it is the actual pullout strength of the threads, show jaydig how you arrived at it to convince him his load is trivial. Working vs ultimate strength has a sizable safety factor in it.
Timelord
Timelord, pullout strength will not be an issue with ANSI threads as they are designed to alleviate that concern. Jaydig should get out Machinery's Handbook and/or ASME B1.20.1-1983 Pipe Threads, General Purpose INCH... run the numbers and convince his/herself. I did years ago.
The attached is fastener oriented but is a good overview for jaydig to study to eliminate the pullout strength concerns.
I would still use sch80 to accommodate 3rd world suppliers.
The attached is fastener oriented but is a good overview for jaydig to study to eliminate the pullout strength concerns.
I would still use sch80 to accommodate 3rd world suppliers.
RCHandy,
You are correct, ANSI pipe thread is designed to hold at least the axial force resulting from the pipe's rated pressure. But here, thread pullout is exactly the issue. It is the question that jaydig originally asked. The 100 lb load on a 3/4" pipe nipple itself is almost trivial, so jaydig must be worried about the female threads pulling out and he doesn't give us the material. The real question is how to calculate the strength of the threads in axial loading because we do not know what material the female threads are put in. jaydig wants to know how to calculate it with his material. My answer was an attempt to pose a simple thought experiment to show him he could arrive at a quick answer by modifying the actual failure strength of a same size pipe thread in the reference paper by the ratio of ultimate strength of his female material to the material used in the reference. It was an attempt to show him his problem is trivial in a thought experiment. If you want an actual calculation, it is the shear strength times the shear area which is the conical surface defined by the thread roots and the thread engagement length. Shear failure in the female threads would be the method of failure when you have a steel pipe threaded into a significantly softer or weaker material and loaded axially. Enough said!
jaydig, Did we answer your question?
Timelord
You are correct, ANSI pipe thread is designed to hold at least the axial force resulting from the pipe's rated pressure. But here, thread pullout is exactly the issue. It is the question that jaydig originally asked. The 100 lb load on a 3/4" pipe nipple itself is almost trivial, so jaydig must be worried about the female threads pulling out and he doesn't give us the material. The real question is how to calculate the strength of the threads in axial loading because we do not know what material the female threads are put in. jaydig wants to know how to calculate it with his material. My answer was an attempt to pose a simple thought experiment to show him he could arrive at a quick answer by modifying the actual failure strength of a same size pipe thread in the reference paper by the ratio of ultimate strength of his female material to the material used in the reference. It was an attempt to show him his problem is trivial in a thought experiment. If you want an actual calculation, it is the shear strength times the shear area which is the conical surface defined by the thread roots and the thread engagement length. Shear failure in the female threads would be the method of failure when you have a steel pipe threaded into a significantly softer or weaker material and loaded axially. Enough said!
jaydig, Did we answer your question?
Timelord
-
1
- #15
Calculate thread shear area: either calculate of the truncated cone created by the thread pitch line or calculate the area of the engaged threads.
Wrench tight thread engagement is 0.5533 in.
Minimum pitch diameter is 0.95429 in.
Maximum pitch diameter at last engaged thread is 1.00179 in.
Average pitch diameter of engaged threads is 0.9780 in.
Area of the trucated cone is 1.7062 in^2
Half the thread is female, half the thread is male, therefore female or male thread shear area is 0.8531 in^2
Using average pitch diameter method:
Thread width at the pitch line is p/2 = 0.0357 in.
Number of engaged threads is 14/0.5533 = 7.74
Thread shear area is pi*7.74*0.978*0.0357 = 0.849 in^2
Not significant error in area calculation: 0.4%
Thread shear stress = 100/0.849 = 118 psi
Tension area at last engaged external thread is 0.0854 in^2
Tension stress at the sch40 pipe thread root is 100/0.0854 = 1171 psi
No allowance made for stress concentration. Assume 4x then the stress is 4*1171 = 4,684 psi.
Compare these stresses with your material properties.
Ted
Wrench tight thread engagement is 0.5533 in.
Minimum pitch diameter is 0.95429 in.
Maximum pitch diameter at last engaged thread is 1.00179 in.
Average pitch diameter of engaged threads is 0.9780 in.
Area of the trucated cone is 1.7062 in^2
Half the thread is female, half the thread is male, therefore female or male thread shear area is 0.8531 in^2
Using average pitch diameter method:
Thread width at the pitch line is p/2 = 0.0357 in.
Number of engaged threads is 14/0.5533 = 7.74
Thread shear area is pi*7.74*0.978*0.0357 = 0.849 in^2
Not significant error in area calculation: 0.4%
Thread shear stress = 100/0.849 = 118 psi
Tension area at last engaged external thread is 0.0854 in^2
Tension stress at the sch40 pipe thread root is 100/0.0854 = 1171 psi
No allowance made for stress concentration. Assume 4x then the stress is 4*1171 = 4,684 psi.
Compare these stresses with your material properties.
Ted
- Thread starter
- #16
- Status
Similar threads
- Question
- Locked
- Question