Thread Engagement

[daparojo]

I have a problem and need some guidance.
I have a bolt that is stretched hydraulically to 1700 kn. It's an M76x6 thread. I have always used 0.8 d to calculate the engagement on the nut/bolt.

Nut width = 0.8 x 76 = 60.80 = 61mm
Material Yield 680 N/mm2
Material US 850 N/mmm

I ran this through a thread strip calculator and it came out with a minimum engagement of 45 mm. This seems a little low. What I need is a nut width of 56mmy max.
Is this possible?

Thanks

 

[RolMec]

Hi,

can't know what your thread strip calculator does, somewhere inside
But, as you are on metric ground, I propose you base your considerations on german standard VDI 2230 edition 2014; chapter 5.5.5 Length of Engagement. This should cover yuor needs.
Regards

 

[desertfox]

Hi daparojo

If the nut and bolt are of the same material then 45mm might well be correct however you haven't told us what the bolt material is or what material the nuts clamping down onto.
If the material your clamping down on does have a lower yield than the nut material then the 1700kN force might not be achieved due to embedding.

“Do not worry about your problems with mathematics, I assure you mine are far greater.” Albert Einstein

 

[daparojo]

Thanks people.

The material of bolt and but are the same. The flange is of similar material.

 

[CoryPad]

Your 1 700 kN prestress is quite low for a thread of this size. The screw has a nominal stress less than 200 MPa, so the nut does not have to be very thick to withstand that force. Your 0.8 ratio would be true if the nut had to resist the full ultimate tensile force of the screw, which is much higher than 1 700 kN.

 

[racookpe1978]

We do NOT have the "guts" (the interior theory and equations and approximations) of your program, so we don't know whether the results are even close or not.

But, most likely, they calcuate (approximate) stresses across the thread roots against tear-out forces caused by elongation. And, in actual fact, almost every nut is over-sized to resist tear-out (so the joint doesn't fail) but also to yield before the stud yields - because it is usually easier to replace a nut than a stud if failure occurs. And certainly much, much cheaper. SO, most nuts have enough thread to resist pull-out in the first 2/3 of their length.

So, the "real" nuts are thicker (taller) than the theoretical nuts due to manufacturer's choice of total yield (safety) margins, and tolerances in the standard nut sizes against weak material and manuf. tolerances.

 

[RolMec]

Just for the big picture, I'd not feel comfortable with a ratio engagement to diameter lower than 0,8. For your setting of thread to strength of material, I'd rather wanted to see a ratio of 1 than a ratio of 0.74 which you strive for. Anything out of the margins of common ground should be analyzed very carefully indeed, + perhaps assessed by someone outside your mill. Safety margins are there to guarantee safety due to imminent incertainties, not to be melted down "because..."
If you consider the safety related background of your application (which is hidden to us): Could you live with a thread strip accident?
R.

 

[daparojo]

Many thanks for your replies, it's appreciated.

The application is a rotating coupling running about 90 rpm.

The problem is that there is an obstruction stopping the nut going on the bolt. We cannot reduce the length of thread protruding from the bolt.

We use normally a ratio of 0.8 x thread diameter for engagement.

We have tried this formula in the past to give us guidance.

Fmax,nut = τ · Ashear,nut · C1 · C3

where

Fmax,nut is the maximum force the nut can withstand

τ is the nut material shear strength - 0.60 x 850 N/mm2 = 510 N/mm2

Ashear,nut is the nut shear area
Ashear,nut = (π d meff)/P · [P/2 + (d - D2)/√3]

π = 3.141 592 654

d is the thread major diameter - 76

P is the thread pitch -6

meff is the thread engagement length - 56

D2 is the nut pitch diameter - 72.478

C1 is a nut dilation factor - you can use 0.91

C3 is a nut thread bending factor - you can use 0.897


Max. Load Approx + 4666 Kn -> This seems rather excessive based on a thread length of 56mmm, personally I would of thought it would of been less.

 

[tbuelna]

Is axial space on the nut side of the bolt an issue? Since you are using a hydraulic tool to pretension the bolt, I'd assume you need more than one nut height of bolt threads protruding beyond the nut for the hydraulic tensioner to attach to.

 

[RolMec]

Just some thoughts:
If the situation is as per att. sketch, then:
- the bolted connection would not be securized against loosening (refer also CoryPads post)lest you prestress it to ~ 90% YS
- if it's the sectional area of the bolt you need for transmission of torque, then why not loose on the thread diameter and add a sufficiently strong washer?
- there should be several bolts on the circumference, any chance to increase their number & loose on the diameter?

 

[tbuelna]

RolMec, after reading that the bolts would be pre-tensioned using a hydraulic device, the amount of axial space at the nut end of the bolt presented an obvious concern. The total axial space required for the nut, the bolt threads needed to attach the hydraulic tensioning fixture, and the axial space required to remove the hydraulic tensioning fixture must be accounted for. The total axial space required would likely be at least 3X the nut height.

 

[RolMec]

The pic was given just for the sake of an illustration of those ideas I tried to convey. For me, sketches always give something to discuss, or clarify possible misunderstandings.
However #1, is there perhaps a way out, if looking at the issue from a different angle? If this solution M76 + hydraulic stretching doesn't work out? At so low a tensioning torque, why not deviate to another tensioning principle? Or changing to a tapped hole, and so on.
However #2, what about securizing the connection? If this is a rigid coupling, there will be a bending moment i. e. an oscillation of the axial forces on the bolt. So loosening would imo present an issue.
Regards
R.