Not necessarily. It depends on the configuration of datum features from which datum axes for position of holes in all 3 flanges are derived. Could you post a sketch showing at least simplified assembly configuration?
Your bolt, with a diameter D[sub]B[/sub], is located at the exact nominal position. Any hole with a diameter D[sub]H[/sub] exceeding D[sub]B[/sub] at the exact nominal position, will clear the bolt. Any clearance over the diameter of the bolt gives you a positional tolerance.
Clearance: C = D[sub]H[/sub] - D[sub]B[/sub]
Your positional tolerance for each hole is the clearance, C.
This is equivalent to a hole called up as ØD[sub]MAX[/sub]/D[sub]B[/sub] with a positional tolerance of zero at MMC. D[sub]MAX[/sub] should be substantially larger than D[sub]B[/sub] since it provides your positional tolerance, as well as the feature of size tolerance.
I should note that these three parts are cylindrical. There are planar face datums and cylindrical face datums and hole datums for clocking the hole pattern.
At the moment I can only say that this is not that simple. In my opinion the formula you want to use, nor the idea of treating all three holes independently will not work. Before giving any answer to this, I think it is necessary to know...:
1. Datum precedence in the positional callout for the hole(s) in the left flange (is flat face primary datum feature, or secondary with datum axis being primary datum?)
2. Datum precedence in the positional callout for the hole(s) in the right flange (is flat face primary datum feature, or secondary with datum axis being primary datum?)
3. What callout(s) exactly is/are applied to the hole(s) in the central flange and what is datum precedence in this/these callout(s)?
4. How is locational relationship between both short cylinders (datum features) in the central part controlled?
5. What controls the relationship between both flat surfaces in the central flange? Is it just a size dimension, or is there something else?
ASME Y14.5M-1994, Appendix B (pg. 205) says: "Any number of parts with different hole sizes and positional tolerances may be mated, provided the formula H=F+T or T=H-F is applied to each part individually."
The very same Appendix B (in para. B.1) says: "Consideration must be given for additional geometric conditions that could affect functions not accounted for in the following formulas."
Believe it or not, but the example being discussed here may contain (and most like contains) those addtional geometric conditions, thus my request for additional clarification. One of the additional conditions may be the fact that it looks that the left flange (and in consequence the left hole) is located in the assembly by the left short cylinder in the central part, whilst the right flange (and in consequence the right hole) is located in the assembly by the right short cylinder in the central part. So if calculations do not take axis offset between both short cylinders into account (and the formulas proposed by 3DDave and given in the standard do not do that), the result will not be correct.
Or saying it differently, floating fastener formula could be used here, if all three holes were located relative to the very same datum axis. And to complicate things even more, this formula would work then only if the datum axis was referenced as primary in positional callouts for these holes.
Yeah, I know, but the original formula was so ridiculously, entirely wrong** it didn't seem easy to explain how the entire stack-up needs to be dealt with.
The formulas assume, and this is always the case, that the related datum reference frames are perfectly aligned. It's too bad the standard has an explanation for addition and subtraction, but no explanation that the DRF on one part is assumed to be located and oriented exactly to the DRF on mating parts for the FCFs on both parts to align. Without alignment, none of the calculations for clearance make any sense.
To look at it another way, if they aren't aligned, there is no general tolerance allocation solution. One can always install the first fastener of a group, tighten it, and find that no other fasteners can be installed. Or one can not bother to line up the parts to begin with, or be able to align the parts.
In this case, the shoulder on one side may be misaligned to the other; the possible forced displacement needs to be accounted for. It's possible the faces aren't parallel; this will add to the displacement.
** Make one hole 3 times the fastener diameter, the other two holes of 0 diameter and locate all 3 holes with 0 position tolerance. Equation shows equality so it's good?
I agree - it is too bad that the standard does not offer more details about application of floating (and fixed) fastener formula.
I also agree that parallelism error between both faces in the central flange will have impact on the result of the stack.
However, I do not agree with such absolute opinion that the formula given by dlinthicum is "so ridiculously, entirely wrong". It will not work - that is true, but in very similar way absolutely valid classic floating fastener formula (for two parts bolted together) can be expressed - if hole sizes and positional tolerance values for holes in both parts are different, the formula T=H-F becomes T1+T2=H1+H2-2F. So it does not surprise me at all that the OP simply added T3, H3 and multiplied F by 3 in that equation. The ** argument that the other two holes can be of 0 diameter is not the way to proove that the formula is invalid, because there is additional condition to that formula that must be met - each H must be greater than or equal to F, otherwise we can't even consider any assembly.
I wasn't proving the equation invalid, but incomplete.
The latest picture shows controls for runout and parallelism which will drive slightly more clearance in the holes in the end pieces; the piece on the right would require more on the basis that it is affected by both runout and parallelism.
The question to ask is if the part must never allow loads on the fasteners.
The MMC allowances on the shoulders means potentially allowing the assembler to shift the part to get the fasteners to fit. If that's OK, then there need be no additional allowance for the left piece. If not, then that datum shift amount would be added to the clearance requirements in the left piece holes.
The holes in the right piece have both runout and datum shift to contend with and a tiny parallism contribution; datum shift applies only if the part should not be guided in position by the bolts.
There's no harm, but no benefit to using one hole as a tertiary datum; they already share a common frame of reference. If it is important, then remember to mark the selected holes and add a note on the drawing so the assembler and inspector know they must be aligned.
The better choice of a primary datum is one that exhibits the most orientation leverage on the part. The stubby part of the shoulder doesn't do this as well as the face of the flange.
Even though it's an interference fit, or especially because interference fits means the parts must be elastic, the shoulder doesn't control orientation of the part very well. A demo is as easy as taking a paper card and jamming a pencil through it. Even though it's an interference fit, the card angle relative to the pencil axis is easily changed; the fit offers little leverage to control mating part orientation over the thickness of the card. Of course, clearance fits have even less control.
It is also difficult to establish the orientation of stubby shoulders in order to verify the orientation of related features. Small imperfections in the surfaces and measurement technique produce relatively large changes in apparent orientation.
I think now there is enough info to give an answer, although I will still be guessing in some points.
First, I would like to say that I agree with 3DDave that faces of the flanges are better primary datum features here, because they will do much better job in orienting these parts in assembly than the short cylinders. (By the way, very nice real-life example with a paper card and a pencil).
If this application is what I think it is (I have some level of confidence, because I work with very similar stuff on daily basis), the fit between secondary datum features, while interference after the parts have been assembled, won't probably be interference during assembly operation - in order to assemble the parts together, either the central piece will be cooled down, or the left and the right pieces will be heated up. So again, this means that flat faces of the flanges will be orienting the parts in assembly, and short cylinders in the central piece are probably "just" for locating all pieces radially.
Furthermore, I have some doubts about referencing datum feature S at MMC/MMB in positional callouts for considered holes. In general, referencing datum features at MMC/MMB means that potentially (when the datum features have not been produced at their MMCs/MMBs) parts can shift in assembly relative to each other. In this case, when a bolt goes through all 3 holes after the parts have been assembled, there is no possibility that the parts will shift radially relative to each other during insertion of the bolt, thus referencing S at RFS/RMB (no modifier) is better choice, in my opinion.
And getting to the point of the discussion... The fact that flat faces are used as primary datum features in the positional callouts for holes complicates tolerance analysis a lot. If one wants to include absolutely every factor that may have any influence on simultaneous ability of holes to accept a floating fastener, trigonometry must be used. It is because the parallelism tolerance between both faces of the central part has to be taken into account (the possible maximum parallelism error, even though it is only .002, makes datum reference frame for position of holes in the right piece not aligned in terms of orientation with datum reference frames for position of holes in the left and in the central part).
As a workaround, I would recommend doing 2 separate tolerance stacks (the number of stacks could be reduced to 1, if size limits of all holes were identical, and all the positional tolerance values were the same):
- Stack I - between the left part (#1) and the central part (#2). Because both components have common mating face, and they are radially located relative to common datum axis (due to the interference fit), datum reference frames for position of holes in these 2 parts are locationally and orientationally aligned. Using simple floating fastener formula, T1+T2=H1+H2-2F or better yet the inequality H1+H2-2F-T1-T2>=0, will give you an idea whether a bolt of max diameter F can pass through the holes of certain size that are located within certain position tolerances.
- Stack II - between the left part (#1) and the right part (#3). Because there is no common mating face between the components, and they are radially located by different features in the central part, datum reference frames for position of holes in piece (#1) and piece (#3) are not locationally and orientationally aligned. The inequality H1+H2-2F-T1-T2>=0, which for parts #1 and #3 would be H1+H3-2F-T1-T3>=0, has to be modified by including a factor coming from the circular runout tolerance (CR) and from a safety margin (SM) reflecting the fact that the parallelism tolerance has not been considered in the stack. So the inequality becomes: H1+H3-2F-T1-T3-CR>=SM, where SM is a number greater than 0.
I am not 100% sure, but to me a separate stack between the central part (#2) and the right part (#3) is not needed - if bolt of max diameter F passes through parts (#1) and (#2) AND through parts (#1) and (#3), it will always pass through parts (#2) and (#3).
