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Request for material recommendation for cam and follower

Request for material recommendation for cam and follower

Request for material recommendation for cam and follower

(OP)
Hello everyone,

I would like to ask if anyone could please help me with the following situation.

I have a very small radial disk cam with an oscillating roller follower that has high contact stress between the cam and roller. Everything on the design is "locked in" i.e., I cannot make the cam or roller larger (except for cam thickness & roller length), I cannot increase cam versus follower displacement, or decrease follower versus cam displacement, and I am using a Parabolic curve, which should give the best minimum radius of curvature and lowest contact stress of just about any curve that is located between two dwell points. This is a very slow moving cam oscillated manually by hand, so I don’t have to worry about the dynamics of the curve, vibrations, etc..

The maximum contact stress between the cam and roller using a 3/8” thick cam & 3/8” long roller is  331,228.24 PSI. I used the formulas in the cam design manual by Clyde Moon to calculate the contact stress along the curve, with the aid of a spreadsheet. I downloaded the design manual from http://www.camcoindex.com/svcman/moonbook.pdf.

It’s difficult to make the cam thicker than 3/8” due to various design constraints, but there is a small chance I could go to a thickness of 7/16” or possibly ½” at the very extreme. This would give a maximum contact stress of 306,657.76 PSI & 286,852.07 PSI respectively.

The maximum contact stresses occur at the point of maximum angular cam displacment, and 90% of the time the cam is not rotated that far. The average maximum contact stress that the cam sees 90% of the time is probably in the range of 220,000 to 252,000 PSI depending on cam thickness. Still, it seems I should design for maximum stress along the entire cam profile.

If the device fails there is a zero percent chance that anyone would get hurt or injured. I don’t think I have the luxury of working with normal safety factors (if any), since the design is on the edge.

My main concern is that I need to avoid plastic deformation, and I need to be reasonably sure that any elastic deformation of the cam or roller will not cause the roller to roll rough or slide, i.e., if the pressure causes a large enough flat spot on the roller, there would be sliding or rough rolling. I am more concerned about these two factors than wear or fatigue, since the cam rotates so slow and intermittently.
 
Can anyone please recommend a material and hardness combination for the cam and follower that would withstand this type of contact stress? I want to use something that is as cost effective as possible to machine, heat treat, and work with. What metal properties do I need to be most concerned with ? I would think compressive yield and shear strength would be the two most important properties to consider, along with how easy the material is to work with.

I found the following materials listed below on www.matweb.com that have compressive yield strengths of over 300,000 and 400,000 PSI, depending on how hard they are (usually between 60 & 64 Rockwell C). However, I am not sure how difficult they are to machine and work with prior to heat treatment. The site gave no machining rating, but said the ASTM 897 grade 5 machines well.

The cam is a very small “rib” cam that has two rollers. One roller works on an inner profile and one roller works on the outer profile. The stresses listed above are for the inner profile, since it has the highest stresses. The cam rib gets thin right at the cam high point (about a .120” wide rib over a short span) in case this could be a problem during heat treatment.

Materials Found on www.Matweb.com

UDDEHOLM VANADIS 6® Hot Work Tool Steel
Carpenter Speed Star® High Speed Steel (Red-Hard) (AISI M2)
Spray Formed Grade ROLTEC SF Cold Work Tool Steel
Spray Formed Grade WEARTEC SF Cold Work Tool Steel
ASTM 897 Grade 5 (230-185-00), Austempered Ductile Iron
UDDEHOLM ELMAX® Powder Metallurgy Stainless Mold Steel

Regarding the cam follower roller, I will be pressing the .1875” OD roller onto a 2mm OD hardened steel dowel pin so the roller “rotates with” the dowel/ shaft. Each end of the shaft is then supported by a low friction self lubricating bushing. I think this arrangement will allow the roller to roll well without sliding between the roller OD and cam profile. I was going to use stock tool steel (i.e, A2, D2, 0-1, W-2 etc.) drill rod for the roller since it already comes in the OD I need and is held to close tolerances. There will be no lubricant between the cam profile and roller OD. I have also considered glass bead blasting the cam profile to increase friction between the cam profile and roller OD, to help insure that the roller always rolls well with no sliding between the cam and roller OD.

My concern with the roller is finding stock round 3/16” OD bar that can handle the high contact stress. It seems to me that it probably needs to be hardened to handle this type of stress. However, when the center of the 3/16” OD rod is drilled out so that it can be pressed onto the 2mm OD dowel, it leaves a thin wall. I am concerned that the roller will distort or crack during heat treatment. I need to make the rollers as cost effectively as possible, and due to the way they are assembled, I cannot make the roller and shaft as one piece.    

The parts are so small I don’t think material cost is a big issue, I am worried that the high strength materials will be hard to work with. I would appreciate any recommendations on the most cost effective materials (easiest to work with) I could use for the cam and follower, and the best heat treatment method for small parts that have thin walls.   

Thank you for your help.

Sincerely,
John

RE: Request for material recommendation for cam and follower

I just finished a roller system very much like a car roller lifter but much smaller. I had to stick with stainless steel for use in the food industry. I think you could use tool steel for the roller and a needle bearing as I did. You need the largest diameter needle bearing you can fit in and the widest. If you look around INA bearings you can buy just the roller needles alone. You could possibley optimize the design this way.  I used 440 c stainless steel for the shafts and retained them with TRU Arc  retaining rings.  I durability tested this unit for over 2.5 millinon cycles and it was still going strong. I only lubed it once a day with food grade gear lube which is nothing as far as good oil. The test cam which in reality represented the fixed machine cam was 6" in dia and the roller had constant contact under load so it was far in excess of what real life was.  The machine currently has 1/2 million cycles on this system now and runs 24/7

RE: Request for material recommendation for cam and follower

I don’t know of any materials that can stand up to these stresses reliably.  Gears and rolling element bearings see the highest contact stresses of any application I am aware of, and the stresses are not as high as yours, but not too much lower.

Conventional highly loaded gearing is carburized, case hardened, and ground.  Commonly used steels for this are SAE 9310, SAE 8620, and 17CrNiMo6.  In manufacturing, 8620 is not controlled for toughness, so good performance for gears requires special metallurgical controls.  This may not be an issue in your application.  17CrNiMo6 is popular in Europe, but not in the states.

Nitriding is another possibility.  It creates an extremely hard case, but the case is relatively thin and brittle, and not at all good for impact loading (which may not be an issue for you).  The thinness of the case may be a problem, though, with those very high contact stresses.

Another possibility is 52100 bearing steel.

Timken makes a wide range of bearing and tool steels and probably could steer you in the right direction.

RE: Request for material recommendation for cam and follower

(OP)
Hi Bentwings1 & Philrock,

Thanks for your replies,

Philrock, you mention my contact stresses are a little higher than those seen in rolling element bearing and gearing applications.

What type of high side contact stresses are generally seen with gearing and rolling element bearings ? This may give me an idea of what to shoot for, as I try to make the best compromises for the design.

The contact stresses are probably going to be high no matter what I do, but perhaps I could get them down to the high side of contact stresses seen with gearing and rolling element bearings.

The Parabolic curve gives the lowest contact stress of most of the standard curves, i.e., Modified Sine, Cycloidal, Harmonic, Modified Trapezoid, and most of the standard Polynomial curves.

I got a demo of the cam design program Dynacam, and it had a few curves that seemed to produce much lower stresses (179,000 to 181,000 PSI). The curves were Thoren, Stoddard, Duddley, Berzake-e, Berzak-d, and "Cycloid first half" & "harmonic first half".

Responsiveness between cam versus follower movement after leaving the dwells is important on the design. The demo program won't let me export anything, so I cannot really compare the responsiveness of the above mentioned curves with the Parbolic. At the very least I would like to superimpose the curves over the Parbabolic & standard curves in AutoCAD, just to get a feel for them.

If I had a displacement output file showing cam angular displacement versus follower angular displacement for each degree of cam rotation, that would tell me what I need to know.

At a price of over 2,000 the Dynacam program is a little pricy for one job.

I wish I had some way to explore the Thoren, Stoddard, Duddley, and Berzake curves, before making a final decision on material. The cam design software I have, won't produce these types of curves. It could be that the Parbolic is already the best compromise, but it would be nice to be sure.

Thanks again for your help,
John

RE: Request for material recommendation for cam and follower


Have you condidered the torque it will take to turn the cam? Sooner or later, this seems like it would be a problem area being of such small diameter.

Is lube striclty out of the picture? At such slow speeds, I'm not so sure a non-lubricated rolling element assembly can take more pressure or last any longer than a lubricated non-rolling single-piece follower.

RE: Request for material recommendation for cam and follower

(OP)
Hi Fabrico,

Thanks for your reply.

The torque required to rotate the cam is fine. The cam follower rotates on low friction self lubricating bushings, but there will be no lube between the cam profile and roller follower OD. I could use something if I had too, only once at assembly, but the device needs to be maintenance free.  

I don't want lube between the cam profile and roller OD because I think it will just lead to more sliding between the cam and roller OD.

I have actually considered ways to increase friction between the roller OD and the cam profile, such as using a coating of "belt dressing" on the cam profile and possibly glass bead blasting the cam profile, to help insure the roller always rolls without sliding.

As far as using a non-rotating follower, I thought of using a 3/16" OD hardened chrome plated steel pin, but I am afraid there will be sliding or rubbing noise. Plus, I don't want the user to have to worry about re-lubricating the cam profile.

Thanks again,
John

  

RE: Request for material recommendation for cam and follower

Nominal allowable contact stress for carburized steel is 225 ksi, per AGMA 2001, but for gears this is further reduced by many factors which don't have equivalents in your application.

When I first responded to your question, I overlooked the point that there will be no lube in the contact zone.  Gear calculations assume adequate lube.  I see this as hopeless without lubricant.  You might be better off using lubricant and letting the roller slide if it wants to.

Lubed or not, I think longevity will be a problem.

FYI, what makes elastohydrodynamic lubrication (as in gears and rollng element bearings) work, is the fact that the viscosity of oil increases sharply under great pressure.  The consistency of oil in the contact zone of heavily loaded gear teeth is roughly that of solid nylon.

RE: Request for material recommendation for cam and follower

Torque to turn the cam is very real!!! Just as a side note in the old days of top fuel drag racing it used to take about 50 ft lbs to turn just a loaded roller cam over when assembling a motor. This is about what it took to turn the crank and pistons too.
MY test noted above started out with a 1/2 hp ac motor and controller. It could not rotate the cam profile roughly 6.0 inches in diameter and .75 lift at 120 rpm smoothly. In fact it woudn't run at all. Even a 1/2 hp dc motor wouldn't do it. We finally set the whole thing up in a Bridgeport and ran it there with out problems. I ran as long as 18 hours with any lube other than what was applied by hand. The entire test lasted about 4 weeks and I only used about 4 oz of food grade lube.
To address your question. The roller won't have a problem rotating against the cam if there is even a little pressure.
In my initial test with the roller, I had only a used roller and we had to get something going quickly. This ran about a million cycles before it failed. It happened on a weekend so how long it ran with the bearing destroyed is unknon but I would guess at least 50,000 cycles without lube. It did tear up the cam some but we didn't have time to replace it so I simply cleaned it up and put the new roller assembly on and reset the counter. The cam was just soft 1020 steel and there was 80# spring pressure on the roller.  The pin was 17-4 ph stainless and about rc40 so far from very hard. My new ones are much harder and smoother. There really was not much wear onthe cam itself. The machine cams have no visible wear after 500k cycles.
Plastic bushings in the rollers lasted less than 4000 cycles but the needle bearings were still running way after 2 million.  I just checked with my customer and they are at 500,000 cycles now and still running with out problems.

99 Dodge CTD dually.

RE: Request for material recommendation for cam and follower

As I mentioned before, you must explore other profiles. The parabolic is not the best for situation and you must remember, most cam programs are more concerned with dynamic forces which are virtually zero in your case. Again give us a sketch of your system and let us see what we come up with to minimize the stresses which are statically induced. You may not need a fancy program to ascertain a better design curve.

RE: Request for material recommendation for cam and follower

(OP)
Hi Bentwings1,

Thanks for your additional feedback.

Bentwings1 wrote:
>Torque to turn the cam is very real!!!<
 
John2004:
I did not mean to imply torque was not something to consider. I just meant that I have calculated it, and it is not a problem at all in my application.

Do you have a rough idea of what the contact stresses were for the cam and roller used in your test ? Just curious.

My cam follower roller normal force goes from about 97 to 130 pounds Maximum, but the cam and roller are so small, it's causing high contact stress. Just to look at this thing, you would think a 3/8" thick cam or at least a 1/2" thick cam would support a roller normal force of 130 pounds without failure or any problems, especially hardened.    

Here are all the specs on the cam which gives you an idea of how small it is...

This is a rib or blade cam with an outer cam follower roller that works on an outside cam profile and an inner cam follower roller that works on an inside cam profile. The inside cam profile is the profile with the high contact stress, due to its smaller radius of curvature.

The inside cam profile is the profile closest to the cam rotation axis, the outside profile is the profile furthest away from the cam rotation axis.  

The outside cam follower roller pushes towards the cam rotation axis (like most normal cams) and the inside cam follower roller pushes away from the cam rotation axis, into the inside cam profile. The rollers create opposing torques on the cam.

Inner Cam Profile:

Curve types between dwells = Parabolic / constant acceleration
Inner profile base circle radius (cam rotation axis to low-point dwell) = 0.8527"
Cam rotation axis to mid-point dwell = 0.9959"
Cam rotation axis to high-point dwell = 1.0949"
Follower swing arm pivot point to roller center = 0.8221"
Cam rotation axis to follower pivot = 1.1475"
X,Y, coordinates of follower pivot = X = 0.8616 and Y = 0.7579 (with cam rotation axis x,y = zero)
Cam follower Roller OD = .1875"
Follower Swing arm start angle at cam low point = 41.33647 degrees(angle between line of centers of the roller center and swing arm pivot, and the line of centers between the cam rotation axis and follower pivot).

Outer Cam Profile:

Curve types between dwells = Parabolic / constant acceleration
Outer Profile base circle radius (Cam rotation axis to low-point dwell) = 1.0117"
Cam rotation axis to mid-point dwell = 1.1332"
Cam rotation axis to high-point dwell = 1.2142"
Follower swing arm pivot point to roller center = 0.7417"
Cam rotation axis to follower pivot = 1.1475"
X,Y coordinates of follower pivot = X = 0.8616 and Y = 0.7579 (Cam rotation axis x,y = zero)
Cam follower roller OD = .1875"
Follower Swing arm start angle at cam low-point = 67.73626 degrees (angle between line of centers of the roller center and swing arm pivot, and the line of centers between the cam rotation axis and follower pivot).  

Cam & follower angular displacements from low-point to high-point:

Cam & follower displacement (Low point to Mid-point dwell)
CCW cam rotation of 18 degrees = CCW follower displacement of 10 degrees

Cam & follower displacement (Mid-point dwell)
Follower dwells for one-degree of CCW cam rotation at mid-point

Cam & follower displacement (Mid-point to high point dwell)
CCW cam rotation of 14 degrees = CCW follower displacement of 7 degrees

Cam and follower displacement (high point)
Follower dwells for 3 degrees of CCW cam rotation at the cam high point.

The cam oscillates between the start of the low and high point dwells. The dwells at the cam low-point and high-point are just there for safety, and are not really used. At the resting or neutral position, the rollers are in the center of the one-degree mid-point dwells. Most of the time, the cam  is oscillated about 8 degrees CW from the mid-point dwells and back to neutral, or 8 Degrees CCW from the mid point dwells and back to neutral. The high contact stress occurs on the inner cam profile near the cam low-point.

At the mid-point dwell or resting position, the inner and outer rollers have equal forces. The spring rate for the outer roller force is 8.67 pounds per each degree of follower swing arm pivot. The spring rate for the inner roller force is 4.072 pounds per each degree of follower swing arm pivot.

Inner roller force goes up with CW cam rotation and outer roller force goes down with CW cam rotation. At the cam low-point, the inner roller force is at maximum and the outer roller force is at minimum. At the cam high point, the outer roller force is at maximum, and the inner roller force is at minimum.   

I chose the Parabolic curve because it had the largest minimum radius of curvature and lowest contact stress of the standard curves. However, the Thoren, Stoddart, Duddley, Berzake-e, Berzak-d, "Cycloid first half" & "harmonic first half" may produce lower stresses, but my software does not produce those curves.    

Thanks for your help.
John

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

Thanks for your post, I did not see your reply as I was posting may last reply.

Please let me know if the additional information on the cam geometry, displacements and dimensions in my last post helps. I can give more information or provide a CAD drawing along with my cam design software ouptut if this can help.

I would appreciate any advice you or anyone else can give on a profile that may yield lower contact stresses than the parabolic curve.

Thanks again,
John

RE: Request for material recommendation for cam and follower

(OP)
Hi everyone,

Zekeman, I just uploaded some DWG and duplicate DXF drawings, along with duplicate JPEG images of the cam, to Rapidshare. I also included output text files from my cam design software.

You can download the files from the link below...

http://rapidshare.de/files/16364719/Rib_Cam_Eng-Tips.zip.html

I would appreciate any advice from you or anyone else on finding a cam curve that will have lower maximum contact stress than the Parabolic.

If there is a curve that will produce lower contact stress, it would be very very helpful to get a CAD file of the curves that I could superimpose over the parabolic curves in AutoCAD for comparison.

Software output showing cam angular displacement versus follower angular displacement for each  degree or preferably each 0.25 degree of cam rotation would also be very helpful as I could use this with my spreadsheet to compare responsiveness of the follower to cam displacement, after leaving the mid-point dwell, to the parabolic curve.

Thanks again, I really appreciate your help.

Sincerely,
John

RE: Request for material recommendation for cam and follower

(OP)
Hi everyone,

I think that one way to solve this problem may be to start with the Parabolic curve (since it produces a farily large minimum radius of curvature), and then increase the minimum radius of curvature further by decreasing both acceleration and deceleration of each curve segment. This would be done at the expense of creating a larger maximum pressure angle, but as long as it's reasonable for an oscillating follower that's no problem.

Even with the existing standard parabolic curve, the outer profile has a maximum stress of around 181,373.00 PSI with a 3/8" thick cam, which is much more reasonable than the inner profile. I would like to shoot for a cam no thicker than 3/8" if at all possible.

I would not want the curve to accelerate and decelerate any slower than necessary to achieve acceptable maximum contact stress, since this decreases responsiveness between cam rotation and the start of follower motion after leaving the mid-point dwell, which is on thing I need to consider on the design. It seems to me a small decrease in the acceleration / deceleration of the curve may do what I need.

Unfortunately, I presently see no way to do that with my cam design software.

Well, this was one possible solution that came to mind I thought I would mention.

Thanks again guys, I appreciate your help.
John

RE: Request for material recommendation for cam and follower

(OP)
Regarding my comments on the Parbolic curve in my previous post, I forgot it already has the lowest possible peak acceleration / deceleration for a given motion, I suppose reducing it further won't be possible.

Perhaps it's still possible to somehow increase the minimum radius of curvature of the cam profile at the expense of a higher maximum pressure angle or some other trade off, I'm not sure. Hopefully there is some curve that will give lower contact stress.  

Thanks
John

RE: Request for material recommendation for cam and follower

(OP)
Hi everyone,

I wonder if a "modified constant velocity curve" might be the best solution for this problem.

With this curve you take a constant velocity curve and then put a radius on the ends of the curve where it blends with the low and high dwells. It will have to be at least equal to the roller radius to avoid undercutting, but I will make it as big as I can.

After leaving the dwell, I want the roller to start moving a reasonable amount within a reasonable time. It may also be desirable to make the follower halfway to its' maximum angular displacement at the same time the cam is halfway to its' maximum angular displacement(like a standard curve).

I checked with my spreadsheet and using a constant velocity curve for the inner cam profile produces a maximum contact stress of about 145,000.00 PSI at the cam low point, but I did not check right where it blends with the dwells. This will depend on what size radius is used. There is a 0.195" minimum radius of curvature on the existing outer parabolic profile(which has a maximum contact stress of about 181,000.00 PSI) so perhaps I could shoot for a little bigger than this.

Perhaps blending a constant velocity curve between the low and high dwells, with a .250" or so radius, is the way to go.

I just wonder if there is a better option.

Thanks
John

RE: Request for material recommendation for cam and follower

I think that using an asymmetric constant acc/dec curve might be of some consequence. You start with a higher acceleration for more than half the period followed by a higher  deceleration, the second half for openers but you must look at the reduction in concave curvature at the start.Now that I think about it, you already have a radius of curvature of -.11 at the start,  so that suggestion won't work.
 I am looking into this problem and I must agree that 2000 bucks for that program is not worth it (unless you can get the government to pay). For this problem, it should be fairly staightforward. I will post the equations you need,shortly.
Also you might think of reducing the spring forces where they are significant at the start of the inner segment. As a suggestion, what about constant force springs?

RE: Request for material recommendation for cam and follower

I have run some calculations and cannot corroborate your 300,000 psi stresses.
I get around 200,000 psi at the first curve based on your tabulated data on a cam radius of curvature of .114 and a cam diameter of 0.1875" and width of 3/8".
The Hertz stress should be
sqrt(.35P[1/(rc-rf)+1/rf)]/(2*.375}/E
=sqrt(.35PE[1/(.114-.0937)+1/rf])/(2*.375)
P=130*.419/1.1475=47.5 lbs normal force (noting that cos of small pressure angle is nearly unity.
I got 199,000 psi. at that point which  I think is near the max stress.
If you think that is not the max stress, please post otherwise.

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

Thanks allot for your reply, I appreciate it.

>Zekeman:
>I think that using an asymmetric constant acc/dec curve >might be of some consequence.

John2004:
I tried an asymmetric Parabolic profile, problem is, I have a small radius of curvature at each end of the curve, so if you change the symmetry factor to make the radius of curvature better at one end, it gets worse at the other end.

>Zekeman:
>I am looking into this problem and I must agree that 2000 >bucks for that program is not worth it (unless you can get >the government to pay). For this problem, it should be >fairly staightforward. I will post the equations you >need,shortly.

John2004:
I would sure be grateful for anything you can do, and I appreciate what you have done so far. Thanks !
 
>Zekeman:
>I got 199,000 psi. at that point which I think is near the >max stress. If you think that is not the max stress, please >post otherwise.

John2004:
That is surely the point on the profile of maximum stress, however, after checking my spreadhseet, I may have made an error entering the sign of the radius of curvature in the spreadsheet but using either + or - sign the calculations are still different than yours. This brings about several important questions listed below. Please see below and let me know what you think.

Regarding your calculations, the 130 pound inner roller force at the low point of the cam is created by a linear extension spring. It's not a torque force like the force shown on the outer roller. Therefore it would not be multiplied by the .419" spring moment arm distance.

The .419" distance is the moment arm that the outer roller spring force acts through for the "outer roller only", and is only applicable to the outer roller. As noted in the drawings, the inner and outer rollers use different springs, spring forces, and different spring rates, but I should have noted that the inner roller force is purely linear and does not act through a moment arm, whereas the outer roller force is a torque force acting through the .419" moment arm. The outer roller force is still generated by an extension spring, but it acts through a moment arm. The inner roller is mounted on a slider, but the slider pivots about the follower pivot point with the follower / swing arm.  

This is one reason why your contact stress calculations were different than mine.

The inner and outer roller forces are exactly as shown in the drawings. Please note that the drawings show the different spring rates for the rollers, and show the roller forces at the low, mid, and high point dwells. These are the actual roller forces, and the normal force is the roller force divided by the cosine of the pressure angle, but it usually only adds a few pounds. At the .114" radii point, normal force and roller force are basically the same since the pressure angle is only 2.98 degrees.  

Even if the force of the inner roller were a torque force, I am not quite sure why you are using the 1.1475" dimension for roller force calculations, since this is the distance from the cam rotation axis to the follower pivot point. You would actually use the distance from the follower pivot to the roller center, in place of the 1.1475" dimension, but again, for the inner roller, the spring force does not act through a moment arm, only the spring force for the outer roller acts through the .419" moment arm, which is then divided by the .742" length of the outer roller swing arm, to get the force of the outer roller.    

The formulas I used for contact stress were from the cam design book by Clyde Moon. You can download a PDF file of the book for free here...

http://www.camcoindex.com/svcman/moonbook.pdf

I think the contact stress formula is on page H-1.

I checked my contact stress calculations at the dwell positions with the values given by the program Camtrax www.camnetics.com and they were the same. The Camtrax program will not make a Parabolic curve for an oscillating follower, and that's why I made the spreadsheet to calculate the contact stress for that curve. They will give you a 10 day trial of Camtrax if you want to check it out, It's fully functional for 10 days, then stops running.  

I used the same formula for the entire profile in the spreadsheet, (adjusting it for a positive or negative radius of curvature)so I figured that since I got the same contact stress as the camtrax program at the dwells, my formula was "probably" valid. Now, I'm not sure.

The contact stress formula (in it's simplified form) according to the Moon manual is...

Contact stress = (10)^3 * the square root of (Pn * C) / (L * M)

Where...

Pn = Normal force
C = Radius factor
L = length of cam and roller contact
M = Material Factor = 0.190 for steel according to Moon

For a convex surface...

C = Rc / ((Rc-Rf) * Rf)

For a concave surface...

C = Rc / ((Rc + Rf)* Rf)

C = 1/Rf for a flat surface

Where...

Rc = Radius of curvature of the pitch curve
Rf = Radius of roller foller (.09375)

From the moon book...

"If the radius of curvature (Rc) and the normal force (Pn) are the same sign (Positive or negative) the surface is convex; if they are of opposite signs, the surface is concave. If Rc is infinite, the surface is flat".

Regarding my software output, it appeared to me that for the outer profile the negative radius of curvature is concave, and the positive radius of curvature is convex. I assumed this would be the same for the inner profile. So I used the above (C) formula for a convex profile when I had a positive radius of curvature and the (C) formula for a concave profile when I had a negative radius of curvature. This also seemed to be consistent with a diagram in the Moon PDF file, but the diagram did not seem 100% clear to me.

When Clyde Moon made the comment quoted above about the normal force being positive or negative, I assumed that since the roller is always pushing into the cam profile, the normal force must always be positive. I am not quite sure what he meant by a "negative" normal force. I felt negative must mean the roller is pulling away from the profile, which is never the case with this cam.

At the negative 0.114" radius of curvature point (this is right where the inner curve meets the low point dwell), I calculated a radius factor of -.114 / ((-.114 + .09375) * .09375) = 60.049

C = 60.049

So, using the simplified formula given by Clyde Moon above...

130 * 60.049 = 7806.42 / (.375 * 0.190) = 109563.78

The square root of 109563.78 = 331.004

10^3 * 331.004 = 331,004.00 PSI .

Should I have used the absolute value of the radius of curvature in the formula, and ignored it's sign ? If so, then the contact stress (using a concave radius factor) would be...

.114 / ((.114 + .09375) * .09375) = 5.853

C = 5.853

(130 * 5.853) = 760.91 / (.375 *.190) = 10823.76

The square root of 10823.76 = 104.04

10^3 * 104.04 = 104,040.00 PSI

It seems a little low to me though, given the small cam radius of curvature at that point, and a fairly high load of 130 pounds for a 3/16" OD roller.

I need to be exactly sure when to use the convex or concave radius factor (C) in relation to a + or - radius of curvature listed in my software output, and whether I should use the signed radius of curvature, or ignor the sign and use the absolute value of the radius of curvature in the formulas. Can anyone please shed some light on these issues ?

The formula Moon gives for the material factor (M) is...

M = ((Ec + Ef)/(0.35 * Ec * Ef)) * 10^6

Where Ec is the modulus of elasticity of the cam and Ef is modulus of elasticity of the follower (I used 30,000,000). This worked out to .190 just like Moon said.

Moon also gives a longer formula where Sc^2 = contact stress...

Sc^2 = (0.35 * Pn) * ((1/Rc ± Rf) ± 1/rf) / L * (1/Ec + 1/Ef)

I am assuming that both cases of the ± signs in the formula above are + for a concave profile and - for a convex profile ?  

I have a couple other books on cam design by Rothbart and Reeve and they seem to use slightly different contact stress formulas.

Now I'm really getting confused.

Regarding the dwells, the outer profile dwells would surely be convex, but what about the inner profile dwells? The software output lists both inner and outer dwells as a positive radius of curvature.
 
If Moon's simplified formula is valid, it seems best to use it to avoid the chance of error with a more complex formula.

I would like to reduce the contact stress as much as possible in any case, but perhaps the situation is not as bad as I thought. I can't be sure until I can find out more about the validity of the calculations.

Can anyone please help us verify contact stress for the following scenario ?

Inner cam profile as shown in drawings, roller at cam low-point
Pressure angle = 2.98
follower Pitch curve radius = negative 0.114"
Roller radius = .09375"
Normal force = 130 pounds
Cam thickness and roller length both = .375"    
 
I can email the spreadsheet I created to calcualate contact stress to you. I would also try to copy the page from the moon book related to contact stress.

You can email me at (johnjmechanical   @   Yahoo.com)  lose spaces before and after @ sign. This is the email I use for forums, online, etc..

I sure hope I can get straightened out on this problem.

Thanks again Zekeman, I really appreciate your help.

Sincerely,
John

RE: Request for material recommendation for cam and follower

John,
1)You are quite right. I goofed on the cam force by using the wrong dimension-should have used the arm length which is academic to your problem.
2) The fact that ( I missed it) you already give the force at 130 lbs on the cam and by my calculation, I used about 49 lb is the cause for the whole discrepancy, since the hertzian stress is proportional to the square root of the force. Therefore, your stress should be sqrt(130/48)*199,000, confirming your numbers.
3)You are correct in assuming that the point in question has a convex curvature and when that occurs, the cam radius of curvature( at the cam, not the pitch curve,rc) is equal to the pitch point radius of curvatue minus the roller follower curvature rf, or rc-rf.Please note that rc is not the radius at the cam.
4)So, you have a large force pushing on the cam with a radius of about 0.02 inches, prctically a pencil point, and therefore the stresses you get are correct and very large.
5) Why can't you do something about this 130 lbs spring force-- or can't you?
6) If you were to make the cam nose radius of curvature .05", yielding a pich cam radius of 0.14, it might get the job done, but the acceleration in the first phase wouls have to be reduced. I will look into this shortly and post back.
 

RE: Request for material recommendation for cam and follower

(OP)
Hi everyone,

Thanks for getting back to me on those stress calculations Zekeman.

>Zekeman:
>Why can't you do something about this 130 lbs spring >force-- >or can't you?

John2004:
I tried to use an extension spring with a lower rate, problem is, the springs can only be .390" OD max, and if I go with a smaller rate the spring gets longer, and then it cannot really fit in the space.

Even if the force of the spring would not increase at all, (rate of zero or constant force spring) it would at the very least have to be equal to the 89.28 pound roller force at the cam  mid-point. I don't think I have room for a constant force spring, and must use extension springs.

It looks to me like changing the cam curve as you suggested is the only thing to do. So far, it looks like a modified constant velocity curve may be best, but I am not sure of that. If I take a constant velocity curve and put a radius on the curve ends that is large enough to get the stresses down to acceptable levels, that might work. The stresses should be OK throughout the constant velocity portion of the curve. The ends will be the critical points, which is controlled by the size of the end radii.  

>Zekeman:
>3)You are correct in assuming that the point in question >has a convex curvature and when that occurs, the cam radius >of curvature( at the cam, not the pitch curve,rc) is equal >to the pitch point radius of curvatue minus the roller >follower curvature rf, or rc-rf.Please note that rc is not >the radius at the cam.

John2004:
Actually, since the software output listed the radius of curvature of the pitch curve as negative 0.114", I assumed the actual profile was concave due to the negative sign, and I used the radius factor (C) formula given in the Moon manual for a concave profile.

For a concave surface...

C = Rc / ((Rc + Rf)* Rf)

The profile "looks convex" and I guess for all practical purposes it really is convex in relation to the roller rolling over it. However, in relation to the pitch curve of the roller, I think it matters whether you are talking about the inner or outer profile.

For example, the outer profile (actual cam profile) is the pitch curve minus the tangent to the roller radius. However, the inner profile (actual cam profile) is the pitch curve plus the tangent to the roller radius.

Going from the low-point to the mid point, or mid-point to high point, the software output list the acceleration phase of the curve as a negative radius of curvature, and the deceleration phase of the curve as a positive radius of curvature, regardless of whether it's the inner or outer profile.

This kind of confused me, because the -.114" radii point looks convex in relation to the roller rolling on it, but the software seems to say it is concave.

The Clyde Moon book says that if the normal force and radii of curvature have the same sign (positive or negative) the surface is convex, and if they are of opposite signs, the surface is concave. Whether the surface is concave or convex, determines which radius factor you use in the contact stress formula.

Since the rollers always push into the cam profile, my roller forces are always positive, correct ?

I am afraid I cannot use non-circular gears as Unclesyd suggested.

Thanks again guys, I appreciate your help.

John

RE: Request for material recommendation for cam and follower

John,
I programmed the acc/dec design for the skew case of 10 deg positive and 8 deg negative acc with the result that the convex radius of curvature goes to 0.1284, yielding a minimum cam surface radius of 0.1284-.09=.0384". When combined with the roller contact radius of 0.09 ( I took the liberty of reducing the roller to 0.18" dia to improve the stress picture) , I get the Hertz factor of
1/rcam+1/rroller=1/.0384+1/.09=37 compared with the previous factor of 60. Since we have a square root relationship, this reduces the stresses you got by Sqrt(37/60)=.78, so that your 330,000psi stress is now .78*330,000=257,000 which is manageable.
I will send you the the EXCEL work if you can provide your email address with sheets 1 and 2. Sheet 1 is the normal case and sheet 2 is the skew case. I have not checked them though I see a slight difference with your results on sheet 1, but if you have a commercial program,you  might try to verify my findings.
The programming should work for a swinging system with any input of phi,phi' and phi" you might put in. My a and b entries are for the x,y coordinates of the pivot where I took the liberty of orienting its initial positon
I might add that the confusion of signs stems from the fact that commercial software usually assumes external cams. For internal cams you must change the sign of the curvature results.

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

Thanks a million for all your help, I really appreciate it.

You can send the Excel spreadsheets to (johnjmechanical @ yahoo.com). Delete the space before and after the "@" sign.

Other than what you have come up with, the only other thing at present that looks hopeful is just to use a straight line constant velocity curve, and make the radii at the ends where it blends with the dwells as large as I can.

I thought this would be real easy to make in AutoCAD, but then I realized that I need to make sure the inner and outer roller have the exact same angular displacement at each point of angular cam displacement.

I may have to array the followers around the cam profile, and then run a spline or arc that is tangent to the rollers to create the curves. I don't know if the fact that it's a dual roller system will place any limitations on the size of the end radii I use, since the inner and outer rollers both need to have the same angular displacement, per any given angular displacement of the cam. This may limit the size of the end radii of the outer curve, based on what is used for the inner curve, I am not sure.  

Once again, Thanks a million Zekeman, I really appreciate what you have done !

Sincerely,
John

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

Regarding the skewed curve you mentioned in your last post, is this a skewed Parabolic curve or some other curve type?

The reason I ask is because when I made the Parabolic curve an asymmetric curve (i.e., going from the low point dwell of the inner cam profile to the mid point dwell, when I increased the duration of the acceleration phase to provide a larger radius of curvature at the cam low-point, it then decreases the duration of the deceleration phase, which decreases the radius of curvature at the cam mid-point). So, when making the curve asymmetric, the stresses at the cam low-point go down, but then they go up at the cam mid-point, it's like a catch 22.

I wonder if it's not best to just forget the dwell at the cam low point (it's just there for safety anyway) and extend a constant velocity curve for 1-degree past the maximum CW rotation of the cam (for safety so the outer roller never rolls off the cam or the inner roller never hits the inside cam track). This should take care of the stress problems at the cam low-point, and then free things up so there is more flexibility with what can be done at the cam mid-point to reduce stress there.

Second option:

If there is some way to increase the minimum radius of curvature at both ends of the curve (at the sacrifice of a larger maximum pressure angle in the middle of the curve), this may also be an acceptable way to get stresses down.

Please let me know what you think.

Thanks
John

RE: Request for material recommendation for cam and follower

John,
The curvature at the the midpoint dwell, while getting numerically smaller with the asymmetric curve  is concave at the inner cam and thus there is no problem with stress there. If you mean the outer cam stresses increase you are right and I haven't looked there. Also my characterization of skewness is the same as your asymmetric constant acceleration curve.
On another note,,how can the two follower arms be locked together as you stated and each have individual spring loading rates? Am I missing something?
I now think that a trapezoidal acceleration curve may serve your puroposes well since it moderates the starting curvature may be better than my proposed one.
Finally, if you are only concerned about preventing the outer cam from falling off, why don't you simply  capture the inner roller in a  large enough slot and be done with it.
P.S  Will send the EXCEL info after I look at the trapezoid.

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

Thanks for your reply.

I have uploaded two Excel spreadsheets I used to calculate the contact stresses to Rapidshare, one for the inner profile and one for the outer profile. I created these with the free 602 PC suite...

http://www.software602.com/products/pcs/

The spreadsheets should work fine with Excel and I know they work OK with "Microsoft Works" becasue I checked them in that program. They were saved as actual Excel files.

I also included new cam design software output that divides the curve segments up in 0.25 degree increments of cam rotation. This may be easier to look at than the output I uploaded with the original CAD drawings, which had many more calculation points along the profile. Plus, it corresponds with the contact stress spreadsheets.     
 
Is there any chance you could please take a quick look at the spreadsheets ?

I am worried the +/- sign of the curvature of the inner profile (whether it is considered convex or concave) has confused me to the point where I did something wrong, although our stress calculations at the low point of the inner curve seem to agree (but perhaps that is because the point under consideration is so close to a dwell and has a small pressure angle). I divided the cam rotation up in 0.25 of a degree increments on the spreadsheets. I used a symmetric Parabolic curve for all profiles.

My main concern is whether I used the correct "Radius factor" formula in conjunction with a (+) or (-) radius of curvature as it pertains to the inner profile.

Here is the Rapidshare link to the spreadsheets...

http://rapidshare.de/files/16994195/Contact_Stress_Spreadsheets.zip.html

Here is a column key for the spreadsheets...

A = cam rotation
B = follower rotation
C = pressure angle
D = Radius of curvature
L = Material Factor
M = Radius factor(cell formula depends on +/- sign of curvature)
N = Stress along entire profile
O or P = Maximum stress for inner or outer profile
J= Cam thickness

The rest of the columns are clearly marked. You can change cam thickness or radius of curvature and the contact stress automatically update.  

>Zekeman:
>The curvature at the midpoint dwell, while getting >numerically smaller with the asymmetric curve is concave >at the inner cam and thus there is no problem with stress >there. If you mean the outer cam stresses increase you are >right and I haven't looked there.

John2004:
Actually, I was worried about the inner curve, I think the outer curve can handle a change in the symmetry, since the stresses on the outer curve already look fairly decent.

Important questions:
If a convex and concave surface have the same contact stress, is the stress less harmful on the concave surface ? I had thought that I needed to be equally concerned about the contact stress whether the surface was convex or concave. How do I determine when contact stress becomes a problem for a concave surface ?

>Zekeman:
>On another note,,how can the two follower arms be locked >together as you stated and each have individual spring >loading rates? Am I missing something?

John2004:
The inner roller is on a slider, and the slider is mounted to the follower, so that the slider pivots with the outer roller.  

The outer roller is mounted in a yoke. The yoke is notched to receive another yoke which supports the inner roller. The inner roller yoke can slide on two 1/8" OD dowel pins that  have their ends pressed into the outer roller yoke. Any sliding is only a few thousandths of an inch, equal to any manufacturing tolerances on the width of the cam rib, which would cause the rib to bind in-between the two rollers if both rollers were rigidly fixed.

>Zekeman:
>Finally, if you are only concerned about preventing the >outer cam from falling off, why don't you simply capture >the inner roller in a  large enough slot and be done with >it.

John2004:

Both the inner and outer low-point and high-point dwells are just there in case the cam is rotated further than it is supposed to be rotated, due to manufacturing and assembly tolerances.

Preventing the outer roller from falling off the outer cam profile at it's extreme CW displacement from the centered neutral position is not the purpose of the inner roller.

Referring to the drawings, the two rollers create opposing torque’s on the cam after leaving the centered dwell position. After the cam has been manually displaced Clockwise from the mid-point neutral position dwell via a lever connected to the cam, and then the lever is released, the sole purpose of the inner roller force is to return the cam to it’s centered neutral position dwell.

After the cam has been displaced Counter-Clockwise from the centered neutral position dwell via a lever, and then the lever is released, the force from the outer roller returns the cam to its centered neutral position dwell.

The roller forces actually bring the cam back to the start of the one-degree mid-point dwell and then opposing extension springs connected to the cam return the cam for the 0.5 degree distance to the center of the one degree mid-point dwell. This is done since the roller forces can’t create any torque on the cam once they are in contact with the dwells.

I had thought of using a single roller in a cam track or groove, but decided against it because I thought I would have problems with clearances between the roller OD and the track. That’s usually not a real big issue, but on this design I thought I might have problems.

Due to manufacturing tolerances on the cam groove and roller OD, I will have a minimum and maximum amount of play between the roller OD and the inside of the groove or cam track. What is the smallest minimum amount of clearance necessary to insure that the roller never binds in the track as the cam is rotated ? It seemed to me that even a very small amount would cause problems with my design, plus the manufacturing tolerance adds to this, so I went with the rib cam, and put the inner roller on a slider.

Referring to the drawings, I don't think I have the space to use a single roller in a slot, because then, I would have to put another cam profile on the left side of the inner roller and this would have an even smaller radius of curvature than we are working with now. I cannot have anything to the right of the existing front roller, I just don't have the space.
 
>Zekeman:
>I now think that a trapezoidal acceleration curve may serve >your puroposes well since it moderates the starting >curvature may be better than my proposed one.

John2004:
The first two degrees of acceleration or last two degrees of deceleration of a Modified Trapezoid curve have a larger radius of curvature than the Parabolic, but after that, the Parabolic has the larger radius of curvature and the largest minimum radius of curvature.

I am not sure about a Plain (non-modified) Trapezoid curve, my software won't produce that. You may well be on to something as it seems the stress is at it's worse where the curves meet the dwells. However, the Camtrax software demo www.camnetics.com, showed very very high maximum contact stress for a Modified Trapezoid curve.  

The spreadsheets I designed allow the stress to quickly be calculated along the entire profile, to get an overall picture of which curve may be best. It's just a matter of whether I used the correct formula in the "radius factor" column, as it pertains to a + or - radius of curvature on the inner profile. Even if it's used wrong, it just a matter of changing the + or - sign in the radius factor formula cells to correspond properly to the sign of the radius of curvature.

Thanks again Zekeman, I really appreciate your help.

I sure hope that with your help I can get the stresses down to acceptable levels.

Sincerely,
John

RE: Request for material recommendation for cam and follower

I looked at your spreadsheets and found where you have a
problem with the radius factor which is:
1/rc+1/rf
rc=cam surface radius of curvature
rf= follower radius
Now rc=rp-rf where
rp= pitch radius of curvature
If the sign of the rp is positive it means convex, if negative then it is concave. By this convention
1/rc+1/rf=1/(rp-rf)+1/rf=rp/rf(rp-rf) which is ok for the outer cam surface since that has the proper signs on the output .However as I pointed out previously, the program results for the inner cam surface has the wrong sign. Then if we change the sign of rp (to correct its representation) in the above equation, the radius factor for your inner cam becomes
1/(-rp-rf)+1/rf=rp/(rp+rf)rf.
Now this formula must be used throughout the as is without any regard for signs. It takes that into account. If you do this you will find that there is no stress problem near the midpoint dwell.
As far as my trapezoid, it is worse than the symmetric as well as the asymmetric curve. The most promising thus far is the asymmetric case which should be made more and more asymmetric until excessive pressure angles. The second half of that curve yields negative curvature which is makes the radius factor less than 1/rf . I will send you my corrected version of your results and subsequently my own.
In the meantime, if you have access to the cam programs  try to develop the more pronounced asymmetric acc/dec curves











You changed the formula at varius points in the program

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

I received the corrected spreadsheets you sent to my Yahoo email, thanks for sending them.

Using your suggestion to take another look at a skewed Parabolic curve, I think I may have come up with a suitable compromise on the design.

By return email, I have sent a new DWG which superimposes the new curve (shown in red) over the old curve (shown in green). The new curve is also on a seperate drawing layer. You have to zoom in to really see the difference in the curves. I have also attached new software output and I pasted the new curve data into the corrected spreadsheets you previously sent me.

Referring to the spreadsheets, the stresses look pretty good to me but I am interested in your opinion. I did not want to get to carried away with a large curve skew, because I have other things to consider on the design, such as responsiveness between lever / cam rotation and follower movement after leaving the mid-point dwell.  

Basically, I just eliminated the 1-degree low-point dwell which allowed me to increase the CW cam rotation from neutral(increased from 18 to 19 degrees). Then I put a very slight skew on the curve going from the low-point to the mid-point. This kept the same radius of curvature at or around the cams mid-point, but made a larger radius of curvature at the cam low-point.

Regarding the curve going from the mid-point to the high point dwell, I increased the CCW cam rotation from neutral (Increased from 14 to 15 degrees), and reduced the high point dwell from 3 to 2 degrees. The curve going from the mid-point to the high point dwell is still symmetrical.

I won't actually use the added one degree of CW and CCW cam rotation, the curve just extends 1-degree beyond the actual cam rotation, in the place of a dwell. The dwell was just there for safety anyway, in case the cam were rotated further than it was supposed to. I have stops on the lever, but due to manufacturing, assembly, and adjustment tolerances, I still needed a little safety factor at the cam ends.

Please let me know what you think about the new arrangement.

www.camcoindex.com told me they use 4140 steel or tool steel hardened to Rockwell RC50 for their cams, and never exceed 180,000 PSI contact stress. They say this produces near infinite life for the cams with periodic replacement of the followers. Most of their cams are used in oil baths. However, these cams rotate for millions of cycles and at high speed. My cam is oscillated by hand at very slow speeds and only needs a life in the order of thousands of cycles, not millions.

Given the fact that my cam is oscillated by hand slowly and intermittently, and given the lower contact stresses with the new arrangement, do you think I can get away with not using lubricant between the cam curve and roller OD ? If I do use lube, I could only do it once at assembly, and then never again, unless the customer does it, which I don't really want them to have to do.   

The cam follower roller axle will be supported by plastic bushings (www.igus.com) at each end, and I will put a drop of oil on the bushings at assembly. This should allow the roller to roll with very low friction and I am hoping that most (if not all) sliding between the cam and roller OD will be prevented and/or eliminated.

I had also considered glass bead blasting the cam curve, and/or using belt dressing to increase friction between the cam curve and roller OD, to insure there is no sliding between the cam and roller OD.

I have to drill the roller center out and press it on a 2mm OD hardened steel dowel pin, which will be the roller shaft.  I would like to make the roller out of stock drill rod from www.mcmaster.com since it's held to close OD tolerance and comes in the OD I need. I was considering using A2, D2, 0-1 or W-1 tool steel for the roller and possibly the cam as well. Another option would be 52100 bearing steel.

I would appreciate your thoughts on the best material for the cam and roller, and the best method of heat treatment?

My main concern is with the thin wall of the roller cracking or the roller OD deforming during heat treatment. Perhaps 52100 bearing steel is best for the roller (since bearing races are thin walled) but I don't know if 52100 comes in 3/16" OD round stock.

If the stresses look OK now, and I can find the right material and heat treatment method for the cam and roller, I  think I will be out of the woods.

It's amazing how just a slight curve skew and only one more degree of cam rotation made a difference. You can see by the superimposed drawing that there is only about 0.006" difference between the new curve and the old one.

I really appreciate your help on this Zekeman, with the very limited space on this design, I kind of started to panic when I saw the high stresses, and was not quite sure what to do. I did not think such a slight curve change could bring the maximum contact stress down 100,000 PSI.  

Thanks again,
John

P.S. I tried both Symmetric and Asymmetric versions of the Harmonic, Modified Sine, Modified Trapezoid, Cycloidal, 2-3 Polynomial, 3-4-5 Polynomial, 3-4-5-6 Polynomial, 4-5-6-7 Polynomial, and also tried blending constant velocity with the aforementioned curves.

The best was the Parabolic, and the closest to the Parabolic was the Modified Trapezoid and the 3-4-5-6 Polynomial, with the 3-4-5-6 Polynomial perhaps being a little better than the Mod-Trap.   

RE: Request for material recommendation for cam and follower

The stresses you now have are well within any margin of safety, considering that the allowable stresses you see in the literature take fatigue, wear and dynamics into consideration and assume 10^8 cycles. Moreover, there is ample empirical data that suggests that the load is inversely as the 1/n power of the life. So, for example, if you design for say 10^6 cycles then the allowable stress in your case would be about
S=So*(10^8/10^6)^1/n
So= allowable Hertz stress  for 10^8 cycles life.

I have seen n to be about 3 from bearing literature so this would suggest a factor of 4 which is probably a little too high in your case However, Timoshenko, in his  book on strength of materials( partII, 3rd edition) makes the comment that static Hertz stresses of over  450,000psi are not unreasonable for "hardened steel owing to the fact that  at the center of the ellipse of contact, the material is compressed not only in the direction of the force buit also in the lateral direction".

RE: Request for material recommendation for cam and follower

Correction,
I mentioned that the load is inversely as the nth root of the  life . Since the Hertzian stress is propoetional to the square root of the load then, the stress should vary as the 2nth root of the life. Therefore, for the example given, the 2nth  root becomes the 6th root and the stress factor is about 2, not 4 as written.

RE: Request for material recommendation for cam and follower

(OP)
Hi Zekeman,

I should be able to assemble the cam and follower with reasonable accuracy, but of course in the real world you can never have "perfect" alignment between the cam and follower. Even if it assembles with perfect alignment, you always have some deflection due to load.

Since the stress calculations are based on perfect alignment, it's nice to have some degree of safety factor to account for a very small amount of misalignment between the cam and follower under load.

The comment from the Timoshenko book also makes me feel better about the situation.

I am going to use the new Parabolic curve, I think it should work out well. As time allows I'm going to try to check into other curves just as a matter of interest. I will report back what I find.

Thanks for all your help on this, I appreciate it.

Sincerely,
John

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