Request for material recommendation for cam and follower 1

[John2004]

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

http://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

http://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

 

[John2004]

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

 

[zekeman]

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.

 

[John2004]

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

 

[John2004]

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

 

[zekeman]

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.

 

[John2004]

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

http://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

 

[zekeman]

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

 

[John2004]

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.

http://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 (

http://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

http://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.

 

[zekeman]

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".

 

[zekeman]

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.

 

[John2004]

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