Topic: Calculating Tolerances for a Press Fit
Had a quick question for you, we're trying to add GD&T to an
assembly that gets a bushing pressed in. the hole diameter and
bushing are rather large (90-94" DIA), and so what I've done is add
cylindricity to the hole and to the bushing and then again another
cylindricity to the inner diameter of the bushing after it is
installed to ensure no high or low out of cylindrical spots after
it's fit in. My question is, how do I calculate the geometric
tolerance for the hole and bushing before and after installation? Do
the fixed and floating equations work in this scenario?
Iíve read this email a couple of times and Iím not certain what
geometric tolerance you are trying to calculate. Cylindricity
tolerances control three dimensional form and are contained within
the size limits.
So, if the MMC of the hole and the shaft are compatible, all the
cylindricity control will do is allow the fit to maintain more
If you are trying to calculate some other type of geometric
tolerance for each, like their positional tolerances for clearance
fits, the fixed and floating fastener formulas will work fine.
Floating Fastener Formula:
- MMC shaft (or screw)
Geometric Tolerance for all holes
Fixed Fastener Formulas:
- MMC shaft (or screw)
Geo. Tol. to be divided between the two mating features (parts)
Virtual Condition hole (MMC Concept)
- MMC shaft (or screw)
Geo. Tol. for shaft
- Virtual Condition shaft (or screw) (MMC concept)
Geo. Tol. for hole
If the type of fit you are trying to achieve is an interference fit,
then the minimum amount of interference is the difference between
the least material conditions of the shaft and the hole. The maximum
interference will be the difference between the virtual condition of
the shaft (or its outer boundary) and the virtual condition of the
hole (or its inner boundary), if they have a geometric tolerance
that is not contained within the size limits of each (such as
position tolerance or runout tolerance).
ē Virtual Condition for holes (or their inner boundary)
MMC of the hole minus the geometric tolerance at MMC = Virtual
ē Virtual Condition for shafts (or their outer boundary)
MMC of the shaft plus the geometric tolerance at MMC = Virtual
If the only geometric tolerance the hole and shaft are being
assigned is a cylindricity control, then the maximum interference is
the difference between the maximum material conditions of the shaft
and the hole.
I hope this helps.
Topic: Clearance vs. Transition vs. Interference Fits
What if the difference between the MMC of the hole and the shaft is
larger than the overall size tolerance that we need to impose?
Doesn't this defeat the purpose of cylindricity? The reason is
simply that when you subtract the MMC of the hole minus the MMC of
the shaft you get a negative number is this ok?
That is the nature of an interference fit (or a transition fit, if
the difference between the LMCís isnít also negative). I donít think
it defeats the purpose of cylindricity. As I said in the last email,
what cylindricity does in these situations is to make the surface
contact more uniform, as opposed to hitting harder at one end than
the other. Cylindricity controls the roundness, straightness and the
taper of the diameters, making their interference more uniform.
Still, cylindricity is a far less used control than the usual
profile of a surface (which requires a basic dimension for the
diameter and tolerances size, shape---and even angle and location if
the proper datum references are used) or position (which uses a plus
and minus tolerance for the diameter and doesnít control surface
shape, instead controlling angle and location). It all depends on
what it is you are trying to accomplish.
Topic: Tolerancing Basic Size Dimensions, What Cylindricity
Controls, and Measurement Force
I hope you don't mind me asking you another question...And this is
the type of question I hope your Geometric Dimensioning and
Tolerancing book based on ASME Y14.5 2009 will answer... but we have
a stainless steel machined hub that will mount to a motor shaft. The
hub I.D. is critical and we want it dimensioned so as to insure that
it will fit onto the motor shaft with little or no force.
Is it proper to dimension the I.D. as a basic, then add a
cylindricity callout to control size?
For example: the current dimension is as follows;
dia. 7.935 +0.025 mm - 0
If I wanted to imply the use of a pin gage to check the hole
I dimension it this way?
dia 7.96 BASIC
concentricity 0.025 @MMC
Would that be a correct method to achieve the tolerance we want?
Also, is there a method to specify that the "GO" pin should fall
thru the part
on its own weight?
Would very much appreciate your advice.
It isnít proper to give a size dimension as a basic dimension unless
you are going to use profile of a surface to tolerance both the size
and the form. Also, a gage isnít suited to easily check the two
tolerance zones (inner and outer) that profile generates. It will
check one, but not the other.
If you dimension the ID to have a diameter of 7.935 plus 0.025 and
minus zero, the GO gage pin would be produced at 7.935 (the ID
maximum material condition) and it would have to pass through the
hole with zero measurement force (per ASME Y14.43-Dimensioning and
Tolerancing Principles for Gages and Fixtures). The least material
condition would be inspected at cross sections. The ID would have to
be cylindrical per Rule #1 (in Y14.5) to within its size tolerance
and you must maintain perfect form (cylindricity) if the entire
feature is produced at its MMC.
The more the feature of size departs
from MMC (as produced), the greater the form error that is allowed.
Still, the variations in form and size may not violate a perfect
cylinder that is simulated at its MMC by a GO gage, and the LMC may
not be violated at cross-sections. There are exceptions to this rule
of ďperfect form at MMC must be maintainedĒ, such as for cylindrical
features that are:
2. controlled with straightness of
the derived median line,
3. dimensioned and toleranced using the
4. stock size in the as furnished condition,
features given size dimensions and tolerances that are identified as
applying as an average (AVG) dimension, and
6. features that are
controlled with a feature control frame that includes the least
material condition symbol.
Topic: Implying or Specifying Measurement Techniques
Thank you very much. Just got your book last night and will be doing
some studying of it in the near future.
But, in regards to my question below; how would you dimension an ID
if you wanted to imply the use of a gage pin to inspect it? Again,
this is a hub for a blower wheel that fits onto an electric motor
Thanks a million,
There is no definitive way to imply a gage pin will be used to
inspect an ID. The use of the word BOUNDARY beneath a position
tolerance feature control frame is an implication that the virtual
condition boundary is to be verified, which is what functional gage
pins do. The use of the MMC and MMB symbol in the control (circled
Mís after the position tolerance and any datum features of size)
would make gages easier to use.
The easiest way to insure the use of a gage pin is to write a note.
Something like ďSEE NOTE 2Ē beneath the position tolerance, then
note 2 stating that ďthe position tolerance is to be inspected with
a gage pin of the following size and configurationĒ should do the
Another option is to write a measurement plan that states how the
part is to be measured. Iíve got a chapter in my gray book on how to
write one (pages 436-441). There is also a B89 standard on
I hope this helps.
Topic: Gages That Measure Positional Boundaries
I hope you doing well Ė I really enjoyed the class I took with you
in Las Vegas back in March.
I have been tasked with making a gage that checks the outside
contour of a hose with several ďbend intersections.Ē
I remember you showing us an example of one of these on the
projector at class.
Do you cover the basics of doing one of these in your latest book? I
have been wanting to get one of your new books ordered anyways.
If you have any suggestions, or advice you can pass along to me, I
would really appreciate it.
The illustrations below are explained in my GD&T textbook (the gray
one on my website
The gage dimensions will
come from your CAD model or the complete part drawing requirements.
The size of the trough in the gage will be the virtual condition of
the outside diameter of the hose (MMC plus the positional
tolerance). The tolerance on the trough gage will be all minus on
the trough (so that no bad parts will be bought by the inspector).
10% of the part tolerance is recommended for use on the gage. The
page numbers for the illustrations Iíve sent you are from the grey
textbook. I hope this helps.
Examples of Specifying Positional Boundary Controls on Hoses, Pipes
Gages to Measure Positional Boundaries
Received your latest GD&T textbook yesterday Ė browsed through it,
and it looks like a great resource. Thanks for the care and hard
work you put in to your teaching and textbooks.
Topic: Independency vs. Envelope Principles and Symbols
Howís life been treating you? Itís been awhile since our last
contact. I left the company where I took your classes last year and
moved to the Salt Lake City area. Are you still in the Nashville
The engineers here are having GD&T discussion and I wanted to get a
second opinion on something. It has to deal with form control within
size and the independency rule in the new Y14.5-2009 standard.
1. Is the circled E symbol now required on prints to enforce the
envelope principle of size controls form? I recall this as being an
ISO thing, but Iím not up to speed yet on the new Y14.5 standard.
2. What is the rationale behind the circled I symbol? I know you
have good stories from the ASME committee. What has the committeeís
logic behind the rule? When should it be used?
Thanks for the help and keep in touch.
P.S. I need to get my hands on copy of your latest book. Iíve tried
to find other sources for the new GD&T rules, but no one does it
better than you.
Yes, still in Nashville. Life keeps on going and business is good.
Iíve been to Salt Lake City a few times. Itís clean and the people
1. The circled E is still an ISO thing. ASME Y14.5 doesnít use the
symbol. Perfect form at MMC is still required (other than the few
exceptions; straightness of the derived median line, flatness of the
derived median plane-used to be straightness of the derived median
plane, stock size in the as-furnished condition, use of the circled
L in feature controlled frames, flexible features and now use of the
circled I). The circled I symbol stands for Independency, and means
that the size tolerance and the form tolerance are independent of
one another. It means that size tolerance does not control the form.
2. The rationale behind the circled I symbol was mostly that the
virtual condition boundary (when using the MMC symbol) is the
functional worst mating boundary and the MMC boundary of perfect
form is rendered unnecessary by the virtual condition boundary. So,
the use of the circled I is basically to allow inspectors to measure
MMC easier (at cross-sections) instead of trying to simulate a GO
gage. Since the virtual condition boundary will require verification
of the worst mating condition anyway (perhaps even with a functional
gage), there is no reason to check the MMC boundary of perfect form.
The circled I comes with a warning label. If there is no geometric
tolerance that controls the form used with it and no feature control
frame that generates a virtual condition boundary, the only control
on the form being infinitely bad is that other tolerances (such as
part length) might be violated. Think of a shaft that bows like a
banana. If it bows enough, eventually the length is very short (and
the size limits on the length might be violated).
Thanks for the compliment on the books. It was nice hearing from you
and where you ended up. Maybe weíll cross paths again in the future.
Topic: Least Material Condition and Least Material Boundary
Quick question. In your book on page 275, it shows a 10.7-11.0
diameter with a position tolerance. Imagine instead of MMC it was
referenced at LMC. Is that practical: a feature control frame with
the geometric tolerance @ least material condition, a primary datum
reference, a secondary referenced with least material boundary and a
tertiary with least material boundary? I have never seen this.
Itís used to protect wall thickness, material strength and seals.
The seals wonít be uniform, but they will seal. For uniform seals,
the RFS and RMB concepts are better. Also, there are many ways to
protect material strength and wall thicknesses. The LMC/LMB
principle is only one. It makes parts hard to measure. It changes
the rule on how to measure size limits from a perfect form at MMC
envelope to a perfect form at LMC envelope. MMC is checked at
cross-sections, while LMC is checked for an envelope.
I use it for casting drawings, where my major concern is to preserve
material to machine away in subsequent machining operations.
Topic: Positioning Hole Patterns with Basic Angles vs. X and Y
A question has repeatedly surfaced with respect to positioning holes
with a single datum reference frame. The confusion is associated
with the fact that 5 holes are positioned with basic angles (30)
associated with an axis which has not been identified as a datum.
An alternate approach under consideration would be to establish the
axis as a new datum via the proper datum feature identifier. A new
feature control frame would be implemented (for 5 holes only) with
the new datum axis as part of the datum feature references for the
angularly positioned holes.
Is it legal and / or sufficient to leave the existing Feature
Control Frame positioning all the holes with the basic angles or are
two FCFs necessary? Iím not opposed to either format. I just wish to
be as accurate as possible and compliant with the intent of ASME
As a side note, I look forward to the possibility of meeting you
again at the ASME Y14.5 committee meeting.
Thank you for your consideration.
I thought for certain that I had responded to this, but I can find
no record of it. So, I guess I looked it over and answered it in my
head, but never wrote it down.
Anyway, the attached drawing that you
sent me is well done. Assuming the primary datum feature is A and
that A has been considered for a flatness tolerance, B is then
controlled to A and C is controlled to A and B. There are basic
dimensions traceable from B to all holes and since C generates a centerplane datum, we can surmise the basic location of every hole
from datum centerplane C. The 30 degree typical is fine (not
something we do much anymore, preferring instead to state the number
of times in front of the basic angle that it will be repeated). Then
all 12 holes are positioned to A, B and C.
It may have been more clear if the five holes you are referring to
had basic dimensions given directly from datum planes B and C,
instead of making us deduce their location by giving the basic bolt
circle and the basic 30 degree angles. Still, what has been done
allows anyone to calculate the basic location dimensions from datum
plane B and datum centerplane C. The real question to ask in these
situations is, ďCan you calculate the basic hole locations from the
given information?Ē I think the answer to that here is clearly that
you can. There is no datum axis that anything is being measured
from. The basic bolt circle and the basic angles are given as a
vehicle to allow one to deduce the basic location from datums B and
C. With basic dimensions, there is no accumulated tolerance error
unless you switch datums in mid-stream. That hasnít been done here.
The only thing I would make sure is done is that all of the radii
have tolerances on them. Plus and minus tolerances are fine to use.
It would also be fine to make the radii on the part periphery basic
radii and then give each a profile of a surface tolerance.
To address what options you are considering, I would not choose to
make the axis of the (what?) outside FULL Radius a datum feature.
Using the datum structure you currently have seems much more stable
I also see no need for two feature control frames. You have 12 holes
and they all have the same MMC, LMC and position tolerance
(hopefully one that has been calculated using the correct fixed or
floating fastener formula). The only reason to use two feature
control frames would be if you had two different sets of holes with
different sizes or different position tolerances.
This is one of those rare instances I occasionally encounter where
the holes have been properly positioned. If some feel uncomfortable
with this technique, my only suggestion would be to calculate the
basic dimensions that are equivalent to what is currently to be
deduced from the basic radii of 4.500 and the 30 degree typical
dimensions and put those on the drawing instead. These basic
dimensions could originate at datum plane B and datum centerplane C
and lead us directly to the center of each hole.
I hope this helps. Iíll see you at the next meeting.
Topic: Continuation of Previous Discussion and a Tolerance Stack-Up
Question with Threaded Holes
Thank you very much for your timely response. With 20/20 hindsight,
I believe I recall you had mentioned difficulties with your email
service, and I should have remembered to forward my note to your
Upon reflection, your discussion is very sound and logical, as your
thoughts usually are! This reaffirms my notion that the orthogonal
datum planes in the given example are sufficient for the noted
positional tolerances and creating an auxiliary datum axis would not
be wise. You made a very good point that the ďdatum featureĒ, a
surface dimensioned by a radius (not even a feature of size!), which
would have been used to establish the axis, would be a very poor
candidate. The subsequent inspections would have been difficult due
to the instability of the part and the datum selections for the
datum reference frames.
Another point that I had missed, would be the incursion of tolerance
stack up due to the two different datum reference frames employed
within the two feature control frames for a hole pattern (5
As a side note, I did finish working all of the problems in the GD&T
in 2007 workbook, however, I am still pondering the complexities of
the tolerance stack up of the 5 piece assembly of problem 55. The
discussion within the text book is exceptional (chapter 25),
however, the notion that the 4 threaded holes should be considered
at their LMC to accommodate the analysis perplexes me at this time.
Let it be noted that I havenít given up, just taking a little more
Thank you once again,
As far as the tolerance stack-up problem you refer to. Itís not the
LMC of the threaded hole. Itís the LMC of the screw mounted in the
threaded hole. As my long deceased teacher used to say, ďThreaded
holes controlled with a position tolerance and a projected tolerance
zone arenít to be considered holes in a tolerance stack up analysis.
They are considered vehicles to move screws around.Ē In other words,
he was saying to look at threaded holes and consider them shafts.
They are screws that are allowed to move by the projected position
tolerance of the threaded holes they are mounted into. I call them
mounted screws. In this situation, what we are looking for is the
clearance between the smallest screws and the clearance holes that
fit over them and also the movement that is allowed by the position
tolerance of the threaded holes.
As a side note, the threaded holes are, in theory, allowed some
bonus tolerance due to the circled M in the position control, but
that amount is hard to quantify because of the self-centering effect
of the screw entering the threaded hole, the depth of the threaded
hole and the class of fit. It is a negligible amount that we usually
donít try to put a number on. If we did try to calculate the bonus,
it wouldnít be based on the size of the screwís outside diameter. It
would be based on the holes allowed pitch diameter growth and how
much that allowed the mating screwís pitch cylinder to move.
Nice to hear from you again.
Topic: Formed Tools and Countersinks
I wanted to ask you a question on formed tools.
If you have a hole with a countersink on both sides of that hole and
it is created by a formed tool (the tool creates both countersinks
and the hole at the same time), do you apply the same controls as if
it were create by individual tools or is there some kind of
expression/note that you attach to the datum of that hole?
It can be treated as though it was a counterbored hole. There are
three ways we usually deal with them.
One is that we show the diameter of the pilot hole and its
countersink size and size tolerance specifications together, then
beneath that we show one position tolerance that applies to both.
Two is that we show the pilot holeís diameter and its size tolerance
with a position tolerance below (for example to A, B and C) and then
we add a datum feature symbol below that states that the pilot hole
is a datum feature (for example D). If there are four pilot holes,
we write 4X INDIVIDUALLY after the datum feature symbol (D). Then we
show the countersink with a position control that positions it to D
and below that we write 4X INDIVIDUALLY. This is to note that there
are four different Dís and that each counterbore or countersink is
positioned to its own D. It would also be allowable to call each
pilot hole a different datum feature letter (for example D, E, F and
G) and position each counterbore or countersink to its own datum
The third way is to spec the pilot hole and the counterbore
separately and give each its own position tolerance to whatever
datums you want (usually to the same datums-for example A,B and C).
This allows you to use different amounts of position tolerance for
the pilot hole and the counterbore or countersink.
Some say that a countersink should be controlled with a profile
control, as we would control other tapered diameters, with basic
angles and basic sizes and profile of a surface tolerance to a datum
reference frame. I personally think that is overkill, but it
wouldnít be wrong.
Subject: Gauge Design Question
I am working with a gauge designer that wants to do something
unconventional with the datum- feature relationship in this absolute
pessimistic gauge and I would like to get your take on it. The
picture below shows a portion of the inspected part drawing with the
feature we are trying to inspect for position and the datum it is
called out to.
Normally, we would call out the datum B gauge to have a VC boundary
of .2760. The gauge designer wants to widen it (see below) and
subtract the widened amount from diameter that measures the feature.
Although we understand that this will reject more good parts, we are
not sure if this is capable of accepting bad parts. We also came
across a section in your book that states only the VC or MMC of a
datum should be used and thought you might be able to tell us why.
Any assistance is greatly appreciated. Please call or email if
anything is missing or unclear.
Itís an interesting approach that he wants to take. Itís also not
correct. The Y14.43 standard states that with an absolute gage the
tolerance on the gage holes may only be minus from the virtual
condition or the MMC or maximum material boundary (MMB) as
applicable. Your gauge designer wants to go one way on the datum
feature simulator (plus only) and the other way on the gauge element
that simulates the constrained feature (starting the size at a
smaller size than the virtual condition and giving it a plus
tolerance also, but not enough to make it exceed the .501 virtual
I have no idea why he or she wants to do that. Itís creative, but
there is a much simpler and more correct way to do it. Just make the
datum feature simulator a diameter of .276, its MMC, and make the
diameter of the larger gage hole a diameter of .501. Make the size
tolerance on each all minus tolerance (as much as 10% of the total
tolerance on the feature being simulated). For example, the datum
feature simulator could be a diameter of .276 and have no plus
tolerance, but a minus tolerance of as much as .0001. Then make the
other simulator hole a diameter of .501 with no plus tolerance and a
minus tolerance of .0003, with a position tolerance of zero at MMC
and the datum feature referenced at regardless of material boundary
(or as itís called in the older 1994 Y14.5 standard, regardless of
This gage would be capable of measuring the MMC of the datum feature
and the position tolerance of the larger part diameter. The gage he
or she has designed could accept a part with a datum feature B that
is oversized. Also, what heís trying to do involves transferring
some of the tolerance on the larger diameter to the datum feature
simulator. This gets tricky when the diameters arenít produced just
out-of-coaxiality, but are produced at angles to one another. Notice
the datum feature is a lot longer than the larger diameter. If the
diameters are out-of-parallel to one another, transferring tolerance
from one to the other wonít be calculated as a one-to-one
proportion. It will be trigonometric.
The ASME Y14.5-1994 standard had a similar math error in its
formulas for calculating position tolerance for multi-diameter
mating features, but when I pointed it out (as described above) they
changed it in the ASME Y14.5-2009 standard.
Hope this helps.
Thank you, that is in line with what I was thinking and I appreciate
One more thing. The gage I described uses a zero position tolerance
at MMC. That makes it a Practical Absolute gage, which could, in a
far-fetched theory, allow a bad part to pass.
To make it an Absolute
gage that wonít even in theory accept a bad part, call out the gage
position tolerance as zero at LMC. Or start the size of the gage
element at .5007 and then give it a minus only size tolerance of
.0003 and a position tolerance of zero at MMC. That would make the
outer boundary of the gage hole .5007 (LMC) plus .0003 bonus
position tolerance equals .501 (which is the same as you would get
if you called the gage hole out with a .501 LMC and a zero position
tolerance at LMC).
In either case, they are both Absolute gages.
Some Other Trainer Wrecked Our Design Manual
About a year ago, we asked you for a quote to come to our company
and train a large group of our employees in GD&T. Ultimately, we
decided go with one of your competitors. This guy came in and
trained our people, but during the training in which he was supposed
to be using the Y14.5 standardís rules and symbols, he said the
Y14.5 committee had made many errors and he had a better way of
doing things. So, he proceeded to teach us symbols and rules he had
apparently made up.
Then he talked our management into ďcorrectingĒ
our internal design manualís tolerancing section by inserting these
rules and symbols which donít appear in any standard we can find.
After our designers created a bunch of new design drawings that used
what heíd taught, we sent our part drawings out for bid, and none of
our suppliers knew what any of the symbology we used meant.
short, this guy taught us wrong and changed our design manual to be
wrong and we made a lot of new product drawings that are wrong. And
they are so wrong that our suppliers say they are meaningless.
Youíve got to do something about this! We want to know what you plan
to do to correct this mess.
Just so I understand. You hired someone other than me. He wrecked
the tolerancing section of your design manual and taught your
employees rules and symbols he plucked out of mid-air. Now, none of
your new drawings can be interpreted by your suppliers. And you want
to know what Iím going ďto do to correct this messĒ?
Iím going to advise you to choose your trainers more wisely in the
future, and to ask me for another quote to correct the problems the
other guy created.
Subject: Weldment Question
I would like your opinion on the GD&T for a weldment. I have a roof
that is welded together with two different sheet metal parts and I
would like two rectangle opening to line up after welding so we can
install an air duct. See attach PDF for my idea on how to line up
the rectangle openings. The more that I look at it the more that I
think that it is not right. Also attach is a BMP of the Solidworks
assembly. What do you think?
Alignment is location. Youíve assigned perpendicularity controls
everywhere, and although this will make the sides well oriented, it
doesnít do anything for where they are. If it is warping that
concerns you, complying with the flatness and perpendicularity
controls after welding would help, but if these controls apply only
at the stage before they are welded, even they donít help much. The
rule since Y14.5-1994 is that tolerances only apply at the level of
the drawing they are depicted. They donít apply to the part at other
stages, such as after welded.
It seems to me that you need to consider applying alignment
controls, such as position before and after welding. Working
assembly drawings are allowed to contain information regarding size,
form, orientation and location tolerances that canít be described
I hope this helps.
Subject: Step Datums, Continuous Features, Compound Datum Features
and Patterns of Datum Features
Dear Mr. Meadows,
My name is Brad, and Iíve taken the GD&T course and the Tolerance
Stack-up course from you at University of Milwaukee of Wisconsin.
You mentioned that if we have GD&T questions to contact you and Iíd
like to take you up on that. I have two questions at this time.
The first has to do with combined datums (ie, B-C, or whatever the
proper term for it is) within a feature control frame. But, I guess
me basic question is, does B and C have to be coplanar along one
axis or another? I have an electrical board designer that wants to
put them in opposite corners of a board and an engineer that says
they both need to be along the x-axis or y-axis. I tried to find an
example in your 2009 book, but couldnít find anything.
The second question has to do with continuous feature of size. In
the 2009 text book on page 177 you specify to inspect the CF
(continuous feature) MMC with a GO GAGE. Does this mean that the
surface being specified needs to be the maximum surface of the part
in order to use CF? I have a heatsink that has 50 fins on one side,
but only 45 of them mount a PCB, which are below the height of the
remaining 5 fins. To me, these lower 45 fins are the critical
surface and need to be controlled, but they are not the MMC
surfaces. Can I still use the continuous feature of size callout for
the 45 fins? Or what would be the best way to control these to be
coplanar? Let me know if youíd like an example, ie. 3d step file, 2d
drawing pdf, etc.
I hope things are going well for you and your family and are
surviving the storm battles there in Tennessee.
Itís perfectly legal to call out two features that are not coplanar
or coaxial to each other as a compound datum. In terms of two
non-coplanar surfaces it is called a step datum. In a fixture, it
requires one fixturing block to be taller than the other. In terms
of two holes or shafts that are not coaxial, it is exactly the same
as calling both holes or shafts a datum feature pattern (for example
positioning both holes to other datum references or just to each
other and then hanging a datum feature symbol below the position
control that states both features comprise datum feature pattern B).
As for the continuous feature type of control, they can be any group
of features of size (widths or diameters) that you would like to be
treated as one continuous feature of size. They donít have to be the
outermost (largest) features of size on the width. On the other
hand, if you just want the tops of the fins to be coplanar,
profiling them all might be the best way to accomplish that.
Thanks for the concern. My wife and I are fine. This group of storms
didnít do us any serious damage.
I Need to Prove My Boss Wrong
My boss thinks I donít know anything. He treats me like a hangnail.
He belittles my work, my clothes, the way I talk and even makes fun
of the way I look. He said my eyes are so far apart I look like some
sort of alien sea life. He says I walk like a penguin with jock
itch. He said if my head was any bigger, it would need its own zip
What with jobs being so hard to find these days, Iím keeping my
mouth shut. Iíve got two masters degrees and heís treating me like I
wear my underwear on the outside (which come to think of it is kind
of a trendy thing to do now). I need to impress this guy. Iím
thinking that our company product drawings (and CAD models) are
somewhere near the technical correctness of drawings on a cave wall.
If I could pass the ASME Geometric Dimensioning and Tolerancing
Professional exam and be certified as a GD&T expert, I think that
would change his attitude toward me.
Do you have any advice or materials on how best to pass this exam? I
hear it isnít easy.
I think the way to pass any test is to study the material. Take GD&T
courses. Read GD&T textbooks. Buy a copy of the Y14.5 standard and
take it everywhere you go for a few months and read from it every
chance you get.
The senior level exam was 100 questions when I took
it. The technologist level exam was 125 questions. Recently, I was
asked why I didnít teach a course on how to pass the GDTP exam. I
told the person that it wouldnít be possible to teach such a course,
unless you told exactly what was on the exam and taught the students
the answers. That, of course, is not something ASME would look
Someone later asked what materials might be possible for me to
create that would make a good study guide. It got me to thinking.
So, what I did was to write a test that is more than three times as
long as the test ASME gives. The
test and its answer book are for
sale on my website. I sincerely believe that if someone can pass my
test, they can easily pass the ASME exam.
Good luck with the exam and with your boss.
And when you get a
chance, send me that zip code.