Fixed Fastener without Projected Tol

[dthom0425]

Hi all,

I am attempting to use Y14.5-2009, Appendix B.5 "provision for tilting of the axis or center plane when projected tolerance zone is not used" formula to calculate the max shaft dia on my screw.

The only problem is I'm getting an unexpected result. I may be choosing the wrong variables or maybe the equation needs to be modified for my part.

I attached a cartoon of what I'm working with.

Black = housing, tapped for a 6-32 (1.5D) helicoil.
Green = the 1.5D helicoil (6-32)
Orange = top plate. The hole in the center is tapped 6-32 (to essentially captivate the screw once you thread it through the top plate).
Blue = captive washer
Red = 6-32 threaded captive fastener with a necked down portion. The necked down portion is free to move diametrically in the Orange (top plate) 6-32 thread minor diameter.

I am using this formula: H=F+T1+T2(1+(2P)/D) .

F = max diameter of fastener. I am trying to calculate this value
H = minimum dia of clearance hole. For this, I was using the tap drill size (orange plate 6-32) of a 6-32 thread since the tapped hole in the orange part essentially becomes a clearance hole that the necked down shaft of the screw must lie within.
T1 = positional tol dia of the clearance hole. For this I was using the pos tol of the 6-32 thread in the Orange top plate as that is what the screw needs to clear in the mating condition.
T2 = pos tol diameter of tapped hole. For this I used the pos tol of the helicoil.
P = max projection of fastener. For this I used the distance between the top of the helicoil to the bottom side of the head of the fastener (top of helicoil to top of washer).
D = min depth of thread. For this I used the 1.5D length of the 6-32 helicoil.

Does anyone see anything wrong with this approach? The max diameter I end up calculating is way small.

Thanks

 

[AndrewTT]

F, P, D should all be known values. You have selected a faster (#6-32). Use the MMC of F in the equation.

You should be solving for T1, or T2, or H. Obviously you will have to assign values to two of these variables and then solve for the remaining. If you know the positional tolerances that can be held by the processes that will make the holes then you can solve for H.

Look at appendix B.5 in ASME Y14.5-2009 for an example of this equation in use.

 

[chez311]

Orange = top plate. The hole in the center is tapped 6-32 (to essentially captivate the screw once you thread it through the top plate).
(...)
Red = 6-32 threaded captive fastener with a necked down portion. The necked down portion is free to move diametrically in the Orange (top plate) 6-32 thread minor diameter.

I had to re-read this in order to see what your design requirements are. This seems to be the limiting portion.

Typically as AndrewTT hinted at H is your unknown and can increase or decrease proportionally based on your constants T1/T2 dictated by expected process limits. Based on your specific design, H has a relatively stringent limitation as it must capture the fastener (presumably for assembly/service requirements). If you didn't have this requirement your design would have more leeway for larger position tolerances. Since you are limited by a maximum diameter for H, this puts a squeeze on things.

What I would do is set H to the maximum diameter you can allow while still capturing the fastener - tap drills for aluminum/steel usually have 75% and 50% thread respectively, you could probably get away with less than 25% thread since it just needs to capture it and not hold any torque, though some experimentation might help determine this limit. Your diameter F is perhaps not a constant, but it does have a minimum limit based on your applied torque ie: you do not want the fastener to yield or fail after applied preload or in service. You should determine this minimum limit, apply what you think is a reasonable tolerance to either T1 or T2, and solve for the other. From there you can increase your diameter F until you get to a value for T1 which can be held by manufacturing. You may have to play with the value for T2 or T1 (whichever one you set as a constant) in order to achieve this, as well as have a conversation with your manufacturing department (or supplier) as to what tolerances for T1/T2 can be achieved with their processes, or what kind of increased precision is possible (and at what cost). You could also set T1/T2 as constants and solve for F and change T1/T2 until your value for F is above your minimum - either way its going to take some iteration.

Fair warning, depending on how the values for F and H shake out (as well as your projected height) these T1/T2 may be very tight and come at a commensurate cost. At that point, it may be worth looking into an alternate method of making that fastener captive - or an alternate method for the threaded portion of the mating part (ie: some sort of captive, floating nut).

 

[drawoh]

dthom0425,

I claim that the orientation of your screw will be due to the top face of the top plate and the top face of the washer. A 1.5D[ ]helicoil insert is not going to control the tilt angle of your screw. Problem solved?

Try doing this from first principles. Use CAD and draw your fastener at the maximum angle allowed by your projected tolerance. You have no control over the clearance hole. Is this a custom screw that lets you control the shaft diameter?

Again, I do not think you have a problem.



--
JHG

 

[3DDave]

Fixed fastener calculations are most applicable to hardended pins installed with heavy interference fits into strong materials.

When applied to threaded fasteners they become a bit weak as a predictor. The problem is that during initial installation, starting with the first thread engaged, the thread has nearly no orientation limiting. As more threads are engaged, as the fastener is turned, there is more orientation limiting, but usually some clearance remains and the control is not absolute.

Usually there is a transition where the fastener is tightened, causing deformation of the threads and bending as the forces on the underside of a canted fastener head comes into contact with the mating part.

Unfortunately, GD&T experts don't normally inquire as to the stresses induced and go just on rigid body predictions for performance. I have done such calculations, though not for threaded fasteners. It's not been a cost effective analysis to perform.